ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Figure 4.  Deployment Mechanism Se ction View, Deployed Configuration TPS Deployment Mechanism Design Background The initial conceptual development for the TPS Deployment Mechanism began in 2007 by LASP engineer  Ryan Lewis. The design concept shown in Figure 5 utilized a 7-stage stepper gearmotor to deploy and  stow the unit. A latch tower was used to lock the system in the deployed position. Prior to the Critical Design Review (CDR) in 2009, the design evolved and the 7-stage stepper gearmotor was eliminated due  to cost, schedule and vendor response concerns. Other design changes that occurred around the CDR time frame are listed below. 1.  Ball Screw with yoke attachment to the Mast used to give better  mechanical advantage.   2.  Smaller and more readily available actuator chosen to drive the ball screw.  3.  Latch tower eliminated.4.  Bellows used to contain ball screw shaft lubricant.5.  Stiff compression spring/stop tower compressed in the deployed position to preload the system.   6.  Duplex and single radial ball  bearings used to support the mast.  7.  End of travel sensors and hard stops incorporated into the design.8.  Ability to manually actuate the system when the actuator is removed. 9.  Stiff compression spring with overtravel gage added for stowing the system. Figure 5. Deployment Mechanism Initial Concept, 2007. 439
Design Discussion Motor The Deployment Mechanism is driven by a 45°, 2-phase and 4-pole stepper motor with redundant  windings. This actuator includes a detent brake and a 3-stage, 384:1 planetary gear head integral to the  unit. The large gear reduction combined with the 2:1 transfer gears and ball screw results in a large  mechanical advantage for the mechanism. The actuator detent torque, 3.4 N ·m (30 in ·lbf) at the output  shaft, combined with the reduced torque reflected back through the ball screw and transfer gears  eliminates the need for a latching device in the deployed state. The 45° step angle equates to 0.12 /g1014of rotation at the actuator output shaft per step. The nominal  operational rate of the actuator is 128 PPS (Pulses Per Second) which results in an operational speed of  2.5 RPM (Revolutions Per Minute) at the output shaft. At this nominal operational rate, the Deployment  Mechanism requires 31.2 minutes to travel from the stowed configuration to the deployed configuration  and vice-versa. The maximum backlash in the gear-motor was measured at <0.0029 radian (<10 arcminutes). This  backlash is not an issue for the TPS because when the Deployment Mechanism is in the deployed state,  the preload from the stiff compression spring in the stop tower removes all of the play in the system. The bearings in the motor were lubricated with low vapor pressure Braycote 601EFVB grease and the  gears were lubricated with Braycote 601EFVB with 3% MoS 2. All materials in the actuator assembly were  chosen to meet the outgassing specification of <1.0% Total Mass Loss (TML) and <0.1% Collected Volatile Condensable Material (CVCM). A thermal clamp bracket is used to conduct heat away from the actuator and into the ball screw cover as  shown in Figure 6. Polypad 1000, a 0.23-mm (.009-in) thick thermal conducting material, is compressed  between the actuator and clamp to enhance the thermal conductivity between the actuator and ball screw cover.  Figure 6.  Deployment Mechanism Drive Components 440
On orbit, the Deployment Mechanism will only be operated one time, moving the TPS from the stowed to  deployed configurations. During this one-time deploy sequence the Deployment Mechanism is open-loop  controlled. The actuator is commanded to move in the deploy direction and then it will automatically be commanded to stop when either one of the redundant deploy position sensors is triggered. On the ground during Integration and Testing (I&T), the deploy command sequence is closed-loop  controlled, identical to what it is in space. Stowing on the ground is open-loop. The actuator is  commanded to move in the stow direction until it is manually commanded to stop slightly prior to reaching  the stow position. The actuator is then incrementally commanded to approach the stowed position so that the kinematic interfaces of the 3X vee-spherolinder-cones can be closely monitored and dialed-in foroptimal positioning and a flush, slightly preloaded interface. Transfer Gears Both the pinion and spur gear have a 20 diametral pitch and 20 /g1014pressure angle. The 28-tooth pinion gear  is a custom part fabricated from 15-5 PH stainless steel and heat treated to condition H1000. This pinion  gear part includes an integral precision shaft for two radial bearings to slip fit over. There is also a8-mm (5/16-in) hex head at the end of the integral shaft that allows for the Deployment Mechanism to be  manually stowed or deployed without an actuator. The 56-tooth spur gear is a standard off-the-shelf part made from 303 stainless steel. The spur gear is modified to add mounting holes, a precision central through hole and a dowel pin slot. Both gears were lubricated with Braycote 601EFVB. There is a small amount of backlash [<0.0026 radian (<9 arcminutes)] in the transfer gears, but as was stated in the motor  section above, this is not a concern due to the preload in the deployed state that removes all play in the  system.   Figure 7. Focus Mechanism Drive Components, Section View 441
Stop Tower and End of Travel Sensors The stop tower assembly is used to  preload the system in the deployed state. This is accomplished with a stiff  compression spring that is mounted  inside the stop tower bracket. As was mentioned earlier, the spring is compressed by 2.54 mm (.100 in) in  the nominal deployed position, which results in a force of 1,628 N (366 lbf).When this force is opposed by the ball  screw, all backlash in the system is  removed. This results in a stiff design  that meets the solar tracking requirements of >6 Hz in the deployed  configuration. Redundant end-of-travel  sensors are also mounted to the stop  tower bracket. These sensors arephototransistor/infrared emitting diode  pairs and their positions are dialed into the ideal spot during the initial  assembly of the Deployment Mechanism. Interrupt flags are mounted  to the mast as shown in Figure 10. Cover boxes are mounted around the sensors to prevent external light from potentially causing sensor reading errors.  3 rdAxis Bearings The 3rdaxis bearings  that support the mast assembly (see Figures 3 and 11) are  comprised of an angular contact duplex pair on one side and a single radial ball  bearing on the opposite side. The duplex back-to-back pair was  chosen for excellent radial, axial and moment stability. The bearing pair size is: 80mm bore x 125-mmoutside diameter x 22-mm wide and there is a  311-N (70-lbf) preload. The single radial bearing size is: 80-mmbore x 140-mm outside diameter x 26-mmwide. This bearing has an axial preload of 44.5 N (10 lbf) that is applied with  Figure 10. Stop Tower and End of Travel Sensors Figure 11. Deployment Mechanism Vertical Section View 444
Figure 13.  Deployment Mechanism Life Test Data for Cycles 34-40 The maximum current  observed during the life test was 1.2 amps at the  nominal operation voltage  of 28 volts. This correlates  to 0.89 N ·m (7.9 in ·lbf) of torque required by the  actuator to stow or deploy  the mechanism. This  number was calculatedusing the actuator’s torque  constant of 26.8 mN ·m (3.8 in ·oz/watt). During  qualification testing at the  vendor prior to shipment,  the actuator pull-in torque at  the nominal rate of 128 PPS  was measured to be 14.1 N ·m (125 in ·lbf), see Figure 14. From these measurements, and using a safety  factor of 1.5, the actuator torque margin was calculated to be +9.4. Backdriving Test In early 2012, shortly after the life test, the Deployment Mechanism successfully passed the backdriving  test. This test verified that the Deployment Mechanism is able to deploy and stow in any orientation in a 1G environment. A Flotron was used to perform this test as shown in Figure 15. The Deployment  Mechanism had healthy torque margin while deploying and stowing in all orientations. Figure 14. Actuator Speed-Torque Curve at Various Temperatures 447
The backdriving test  also verified that the TPS remains in a stable position in the worst case vertical orientation due to gravity. This was done by measuring the gap between the spring cap and the stop tower bracket(see Figure 10) while the TPS was in the  deployed  configuration. The gap remained unchanged after 24hours, thus verifying that the system did not back drive and that the detent torque in the motor was appropriately  sized. The torsional spring rate of the Deployment Mechanism system in the stowed configuration was  calculated during this test. During the test, gap measurements between the mast and the baseplate were  taken at various times and it was noticed that the gap would change slightly after the Flotron was rotated  to a new position. This gap change was the result of the gravitational force acting on the optical bench mass simulator mounted to the top of the mast as shown in Figure 15. Certain orientations resulted in  more torque imparted to the mast rotation axis compared to other orientations, thus resulting in a smaller or larger gap. The angle between the mast vertical axis and the baseplate is easily calculated from this  gap measurement.  The torsional spring rate in the stowed configuration was calculated with the Flotron in the horizontal  position. A gap measurement verified that the mast angle was at the nominal 15 /g1014. At this point the optical  bench mass simulator plates were removed and the gap was measured again. This time the gap  increased by 1.14 mm (.045 in) which correlates to the mast angle increasing to 15.386 /g1014.U s i n gt h i s   change in angle measurement along the moment torque caused by the mass simulator plates, the  torsional spring constant of the Deployment Mechanism system in the stowed configuration was  calculated to be 674 N ·m (5,967 in ·lbf/degree).  Future Testing A gimbal characterization test will be performed early in 2014 with an optical bench mounted to the  gimbal and representative mass models mounted to the optical bench. This test will characterize the gimbal pointing performance when it is mounted to the TPS. Following that test, the engineering model assembly will be expanded to include the launch towers and resettable separation nuts. This unit will then undergo tests to verify deployment and stow/alignment capability. A modal test is also planned in both  stow and deployed configurations and the results will be correlated to the structural model. Figure 15. Deployment Mechanism Back Driving Test on the Flotron 448
Conclusions  The Deployment Mechanism design has been proven to be a viable method for deploying and stowing  the TPS. The mechanism successfully passed the lifetime and back driving tests. Future testing will undoubtedly reveal more nuances about the design and its features.  Lessons Learned During the life test, the actuator thermocouple produced transient spikes that exceeded the +150C  temperature limit, setting off alarms in the software that resulted in the test automatically stopping and causing delays. This was fixed by increasing the thermocouple sample rate to 200 msec and averaging 5 samples. The peer review process is extremely helpful in producing a successful design. Prior to CDR, two peer  reviews were held at LASP with many LASP mechanical, electrical and systems engineers in the  audience. The insightful questions from the audience and the action items that came out of the reviews  resulted in improvements to the design. Because of these design improvements the fabrication, assembly  and testing of the assembly went smoothly without any major problems. Acknowledgements The author would like to thank all of the LASP machinists, technicians, engineers, scientists and  managers who played a role in developing and building the Deployment Mechanism. The success of this  effort was a direct result of their skill, commitment, hard work and enthusiasm. References 1.  JPSS website. http://www.jpss.noaa.gov/ 2. TSIS mission Quick Facts. LASP website. http://lasp.colorado.edu/home/missions-projects/quick-factstsis/ 449
ExtraVehicular Activity (EVA). The pump module ORU houses the pump, controls, and valves required to maintain proper temperature in the heat rejection loop. A canned motor cartridge pump circulates this  heat transfer fluid through the thermal control system. A typical canned motor cartridge pump lies at the  heart of the system. The pump must run continuously for the 15-year predicted life of the module in order  to maintain safe operation of the ISS. The 2010 Cooling Pump Shutdown Shortly before midnight on July 31, 2010 the ISS cooling Loop A on the Starboard side of the vehicle went  offline when a Remote Power Control Module opened during a high current draw event. As a result the  crew was instructed to reconfigure the ISS thermal management system to rely solely on Loop B.Subsequent remote procedures were unable to restart the pump and very likely caused additional  mechanical damage to the hardware. Over the course of the next 17 days the ISS operated in a zero-fault tolerant mode, relying only on Loop B for all active thermal control. During this time period electrical power input to the habitat was limited and many of the scientific experiments were placed in standby mode. Astronauts Doug Wheelock and Tracy Caldwell-Dyson performed three EVAs totaling over 25 hours to replace the failed pump module with a spare ORU. The EVAs proved to be much more difficult and dangerous than anticipated due to frozen ammonia that jammed the quick disconnect fluid couplings  used to connect the pump module to the rest of the thermal control loop. During the de-mating procedure on the failed unit, frozen ammonia particles were liberated from the connectors and caused both a  micrometeorite type of hazard for the two astronauts and a health concern upon re-entry to the ISS as the  frozen ammonia particles were stuck to the outside of the spacesuits. On August 17, 2010 function of  cooling Loop A was restored and ISS operations returned to nominal. The failed pump module was stored on the ISS external structure for 11 months. The module was  returned to earth on the final Space Shuttle mission, STA-135. During these initial 11 months, a team was assembled and failure investigation was initiated. Early root cause failure paths focused on the electrical power and control system. Initially this approach was logical since the primary data that was available to  the team was the initial over-current RPCM shutdown and the inability to restart the motor due to persistent high current draw. Once the pump module was returned to earth, inspections revealed severe Figure 1. Thermal Control and Energy Balance of the International Space Station 452
mechanical damage to the motor and in particular the graphitic carbon bearings. A new fault tree branch  was then investigated with emphasis on the hydrodynamic bearings, rotordynamic vibrations, and wear life of the bearing materials.  Pump Mechanical Design The pump used in the ISS external active thermal control system is a custom-built and space-qualified  canned motor cartridge pump. A typical canned motor cartridge pump is shown in Figure 3. This type of pump design was selected for several reasons. The main driver for  the selection of this type of pump is  the elimination of fluid seals on the  shaft. High-speed shaft seals are typically the life-limiting component on pumps and a seal failure on the ISS ammonia system poses a great  risk to the vehicle and crew. An additional reason for the selection of  this type of pump is its compact and  lightweight architecture. Finally, this type of pump has a robust heritage in  spaceflight hardware from the early  manned vehicles to the Mars Pathfinder rover; canned motor cartridge pumps have proven to be robust and reliable. The challenge Figure 2. Overview of the ISS Highlighting the Ex ternal Active Thermal Control System Showing  the Orbital Replacement Unit (ORU) and th e Pump Control and Valve Package (PCVP). Figure 3. A Canned Motor Rotor with Centrifugal Pump,  Typical of Actively Pumped Thermal Control Systems for  Space Applications. 453
Figure 7.  Rotordynamic Analysis Results rotor. It is good design practice to specify the  correction planes as near the bearings as is practical. By locating the balance planes near the bearings, the final unbalance recorded by the dynamic balance process can be used to calculate bearing loads directly. For the  specific pumps in question here, the balance  planes are located near the rear bearing and  on the thrust disc near the front bearing. Build records were examined for all motors and pumps, not just for the failed unit, and bearing  loads were calculated. The results indicate that  the imbalance load on the rear bearing goes from 0.11 N at 12000 rpm to 0.17 N at 14700 rpm. These loads are well within the capacity of the rear bearing but do approach 20% of the load capacity. The residual imbalance alone would not overload the bearing and cause damage; however, it may either excite a rotordynamic vibration or combine with other forces to overload the bearing. Damage associated with a bearing overload from residual unbalance would manifest itself around the entire periphery of the bearing. This type of damage was not found during the physical inspection of the failed pump. Therefore, this force could not be the root  cause, as it needs a force that organizes the overload in one specific direction.    Radial Magnetic Asymmetry The drive for the pump is a four-pole  permanent magnet motor. As such, there exists the possibility that radial asymmetries of  the stator or rotating group position could cause a radial load on the bearings due to the negative stiffness of the magnets. During the  design, build, and qualification of the motors no testing was conducted to quantify the radial  force due to asymmetry. However, this quantity  is bounded by the contract specifications to the motor supplier to be no greater than 10.7 Newton. Assuming this load would be equally split between the two bearings since the motor  is straddle mounted, this would impose a 5.35  Newton load on each bearing. This magnitude of load is sufficient to overload the rear journal bearing and cause wear. However, the asymmetry would have to manifest itself through at least two poles of the motor stator in order to cause the angular misalignment damage found during the physical inspections of the failed hardware. A two pole stator asymmetry would manifest itself in reduced motor efficiency.Qualification testing of the motor did not reveal this characteristic. A magnetic asymmetry in the rotating  group would manifest itself around the entire periphery of the bearing while a single stator pole asymmetry would cause the damage on both bearings to be in the same angular orientation. Therefore, this force could not be the root cause, as it needs a force that organizes the overload in one specific direction. 457
Annulus Fluid Inertia Forces Annulus fluid inertia forces were first discovered and analyzed by R. J. Fritz in the early 1970’s. These  forces arise in canned motors where high density fluids fill the rotor-stator gap of the machine. When accompanied by high speed, non-concentric operation, the fluid in this gap can impose a negative  stiffness between rotor and stator, similar to the asymmetric magnetic forces. This causes a bearing load that increases with increasing eccentricity and speed. At 14700 rpm and an assumed eccentricity of 25 micrometers, the load on each bearing would be 2.58 Newton from the annular fluid force. This force is  sufficient to overload the rear bearing under most hydrodynamic conditions. However, these annular fluid  inertia forces for canned motors are well understood for eccentric rotor motion in a cylindrical mode. For  the failed ISS ammonia pump an angular misalignment was found from physical hardware inspections. It is unclear whether these inertial forces behave in a similar fashion for this type of misalignment.Additionally, this force would rotate with the eccentricity causing wear around the entire periphery of the bearing, which is inconsistent with the findings of the physical inspections. It is unlikely that this force was the root cause of the wear because, like the previous forces, it requires an additional force that organizes  the overload to occur in one specific angular location.Figure 8.  A Summary of the Physical In spections of the Damaged Carbon Bearings .  458
Verification was needed to confirm the hypothesis that the lubricant jet force on the rear bearing served  as the root cause for the ammonia pump failure. While analysis of the bearing capacity, analysis of the magnitude of the lubricant jet force, and the findings of the physical inspections were consistent with this fact, experimental validation of the analyses was needed to run the issue to ground. To this end, a flow visualization experiment of the rear bearing lubricant feed geometry was conducted. Still pictures of the lubricant jet characteristics at 11000 rpm and 15000 rpm are shown in Figure 9. In these experiments water was used to simulate the anhydrous ammonia in order to match the kinematic viscosity between the two fluids. In the flow visualization blue dye was added to the lubricant jet in order to distinguish between the fluid in the jet and the fluid resident in the aft bearing pocket. The pictures clearly show a change in the character of the jet force between the low speed operating point of 11000 rpm and the high speed operating point of 15000 rpm. At the low speed point, the jet is slow to emerge from the feedhole and diffuses into the surrounding fluid. These characteristics are consistent with a low imposed jet force.By contrast, at the high-speed condition the jet remains organized and coherent up to the point of impact  with the shaft extension. This characteristic is consistent with a significant jet force imposed on the shaft and rear bearing. Because the bearing load capacity increases linearly with rotational speed but the jet force increases with the square or cube of shaft speed, a crossover point can be identified at which the  rear bearing margin goes negative. Once the bearing margin goes negative, solid surface contact occurs  and wear damage of the carbon bearing begins to accumulate. This initial damage changes the geometry of the bearing, which in turn predisposes the rotor-bearing system to angular misalignment in this direction. These flow visualization experiments provided verification that the lubricant jet force initiated damage to the rear bearing and contributed to the premature failure of the pump. In order to restore operational life of the remaining pump modules, procedures were modified to limit maximum speed to 12500 rpm, minimize changes to pump speed, and minimize start-stop cycles on the pump. Figure 9.  Flow Visualization Images of the Lubr icant Jet Force and its Interaction with the  Shaft Aft of the Rear Bearing .  460