ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Damper Rotor Driveshaft Damper Housing (Stator)  Spring Arbor  GSE Drive  Hex Hex Drive  Diaphragm  Flexure (Spring mounting Barrel  Washer and Nut)  Figure 7. Spring/Damper Driveshaft  The two damping regimes in fact came in very ha ndy. The Spring/Damper had a full stroke of about   100 degrees, while each Mechanism only deployed about 60 degrees. By biasing the deployment more  within one damping regime or the other, the total deployment time and be havior could be tuned. To  facilitate this adjustment, the Driveshaft was designed with a vernier rotational indexing pattern between the Axle’s hex drive and the Damper Arbor’s 22.5-degree bolt pattern (Figs. 6,7). The resulting 7.5-degree  increment (3*22.5 – 60 = 7.5) allowed the room-temperature deployment time to be systematically  adjusted from about 10 s to over a minute in ~10-s  steps. This was finer re solution than practical for a  splined shaft, and much easier to fabricate. The Driveshaft was also designed to account for  misalignment between the Spring/Damper a nd the Axle. The diaphragm flexure accommodated up to 0.5  mm lateral / 0.2 degree angular e rror with very little side load, while the potential for handling damage  from over-travel was eliminated by the shaft snubbi ng against the Damper Rotor.  A vernier pattern was also used for locking the Spring/Damper in nearly any rotational position. A pin  inserted between holes in the Spring Arbor and Damper Housing (Fig. 6) provided 1 degree position  increments, essential when aligning the Driveshaft and during assembly and testing operations.  Figuring out how to assemble the Spring Motor was its own R&D effort. The trouble was how to uncoil the  unwieldy and incredibly stiff Spring Stack in order to install the bolt that mounted the Stack to the Spring  Arbor. The solution was to assemble the Stack over a spreader bar that held the coil open and allowed  the bolt to be installed. Both Stacks were mounted to the Arbor, and then hooked  over the Roller Pins.  With a big hand tool, the Arbor was then cranked in to position and attached to the Rotor. This system  worked, but resisting the torque of the Springs while trying to align the Arbor and drop it onto the shear  pins in the Rotor was tricky, and could result in damage to the interface. Later, a much easier and safer  procedure was developed where the Arbor/Sprin g assy and Rotor were mated first, the Springs were  retracted with a special tool, and then the Roller Pins were installed. The lesson learned was that  assembly considerations and tooling design are just as important as the hardware design. Telemetry The Deployment Mechanisms were equipped with microswitches that provided telemetry about the  stowed/not stowed, not deployed/deployed, and not latched/latched states (Fig.  8). Like the Latch, these  Switches were implemented such that they were depressed in the stowed state. The Stowed and  Deployed Switch Actuating Levers rode on  cams and toggled within 1 degree of the Axle stowed and  241
deployed positions. The Axle had no stop in the stowed direction, so the use of cams allowed Axle overtravel without damage to the Switches. The cams also mitigated the potential of the deployment being  stopped by a misaligned Switch [2].  Since the Switches were considered ancillary, they were not given appropriate design attention early on.  This resulted in less-than-ideal placement from mounting, functional, and cabling standpoints. Their  relative proximity to the Pin Puller also raised concerns about shock sensitivity. A ceramic button in the Switch had a record of failing due to pyroshock. Live firings were performed to measure the shock levels  across a breadboard Hinge, resulting in a program to qualify the Switches to 6000 g’s. While the Switches  themselves passed, the standard two-bolt mounting system failed to keep them from shifting. They had to be adhesively bonded to Brackets, with the Brac ket then given an adjustable in terface to the hardware.  The lesson learned was that testing will uncover problems that were never imagined. Stowed  Switch  CamsSTOWED DEPLOYED Switch  Bracket Latched  Switch Deployed  Switch Adjustment  Interface  Figure 8. Telemetry Switches  Cable Spool Electrical cabling to the Reflector and UDM had to pass across the hinge-lines. This was accommodated  by a flexing section of round cable held within the Cable Spool (Fig. 9). The section was wrapped snuggly  around the Spool for launch, and unwrapped as the mechanism deployed. The fixed ends of the Cable  were rigidly attached to the Clevis and Axle via Brackets, so cable motion was controlled and repeatable.  This was done with lacing cord for “simplicity” of adjustment, but was actually rather tedious to implement.  Axle  Bracket Lacing  CordFlexing  Cable Section Axle  Bracket Lacing  Cord Clevis  Bracket Figure 9. Cable Spool  242
Inner Spring  Laminations  Buckling Large OD  Roller  w/Braycote  Figure 10. Actuator Spring Buckling  Reconstructing the system used to fill the Dampers was again another big task. Filling was done under  vacuum, with a gas-actuated “syringe”. The fluid was first outgassed, then drained into the piston of the  syringe and injected into the Damper. Everything was heated to expedite the process, since the fluid was extremely viscous. Learning the idiosyncrasies of the system took practice, and some redesign was  required to get it working properly. One unexpected difficulty was simply finding a location for the bell jar.  Intentionally heating and outgassing silicone oil was not something folks wanted anywhere near their clean rooms. Instead, a clean tent had to be built. The first usage was filling heritage actuators to  characterize how fluid viscosity affected deployment rate. As the viscosity was increased, the damper  began having erratic behavior, similar to that reported by Stewart, Powers, and Lyons [7]. It was suspected that the fluid was cavitating, or the damper was not filled completely. Having not anticipated  such a problem, the fill volume had never been pr ecisely characterized, nor the empty damper weighed  for comparison. Disassembly might destroy the ev idence. The damper was x-rayed, but the energy  required to penetrate the thick aluminum housing left the fluid nearly transparent. However, by standing  the damper on end for a while, and then flipping it on its side and taking a series of x-ray images, a very  faint line was seen to slow ly move (Fig. 11). Since no free movement was possible in a full damper, the  line was assumed to be the surface boundary of the fluid where the meniscus was tangent to the view.  This discovery led to the realization that the heritage filling pressure was far too low, and that the fluid  heating was inadequate. These issues were remedied, and there were no further filling anomalies.  Figure 11. Sequential X-ray Images of a Partially-filled Damper Showing Fluid Movement  245
Assembling the Hinges and Latches went relatively smoothly, once all the known dimensional  discrepancies had been corrected. Special tools were made for pressing in the Bearings and Stub Axles,  so that no load was applied across the balls. Fortunately, disassembly had also been  considered in the  design phase. This came in very handy when one of the Stub Axles was installed improperly and needed to be removed. Simple design featur es allowed what could have been a real catastrophe to be only a  minor inconvenience. Lesson reinforced: always consider disassembly, especially for items with press fits. Figure 12. Assembled LDM  Boom To meet alignment requirements, the Boom was bonded using the flight LDM and UDM as fixturing. After  functionally verifying the Mechanisms, they were mounted to a jig that represented the deployed Primary  Structure and Reflector interfaces, and the Boom was assembled between. All piece-part fabrication  tolerances were thereby taken up in th e bondlines, and the end-to-end alignm ent was as good as the jig.  Testing Temp/Atm Deployment and Stiffness Testing The start of mechanism testing turned out to be a real eye-opening experience. The objectives of the first  Temp/Atm Deployment Test were to functionally demonstrate deployment and latching of the Mechanisms at temperature, and to satisfy thermal cycling requirements. Stiffness Testing was then  designed to apply a moment to the Hinge, at temperature, and measure the resulting rotational deflection.  The Clevis was mounted to a rotation stage having a horizontal axis, and a large weight was cantilevered off the Axle to apply the moment (Fig. 13). The resulting deflection was measur ed by linear variable  displacement transformers (LVDT’s). Slowly rotating the stage through 360 degree s resulted in one  complete load cycle and a mapping of the hinge stiffness.  Regrettably, test preparations were neglected due to all the fabrication and assembly issues. In the panic  to catch up, the test engineer took many shortcuts and wound up making many mistakes. In the end, far more time and resources were spent fixing the resulting problems than if testing had been delayed to  allow for proper preparation. Without going into all the gory and embarrassing details, the following list  highlights a number of lessons learned. The most important lesson of all was that test development  requires the same level of attention, scrutiny, review, and verification as the hardware to be tested. /g120Be sure all test requirements are well established and communicated to all supporting parties  /g120Verify the complete test setup prior to testing. That means functionality, mechanical interfaces,  electrical cabling, thermal control, and especially the data acquisition system are all thoroughly demonstrated under environmental conditions  /g120When designing GSE to interface with flight hardware, simply duplicate the flight interface rather  than try to re-engineer it from scratch  /g120It’s better to have two simple test fixtures than one complicated one  /g120Make accurate predictions of the expected test performance 246
Deployment  Mechanism  LVDT Deflection  Sensor - measured stiffness across the  BearingsSupport Stand Rotary Stage Cantilevered  Weights LVDT Deflection  Sensor - measured stiffness in  the deployment  direction Figure 13. Hinge Stiffness Test Setup  The only hardware issue of any significance wasn’t uncovered until well after testing. When the test report  was delivered, it was discovered that the post-test deployment time of each Mechanism was roughly 30%  longer than the pre-test deployment time. The Mechanisms had undergone enough rework since the  testing that studying the hardware would yield little insight. Even worse, the associated data files couldn’t be located. After subsequent testing to rule out some possibilities, the leading suspect was increased  bending resistance of the Cable in the Spool due to the thermal cycling. The investigation into this  anomaly is ongoing, including life-testing of the actuators. The lesson learned is that even if a test seems  successful, review the data immediately and make sure to back it up. Mockup Inertia Margin Testing The objective of Mockup Inertia Testing was to verify and characterize the Mechanisms under flight-like  conditions, and to demonstrate actuator margins. Each Mechanism was bolted to the wall of the lab, and  an insulated thermal chamber was built around it (Fig. 14). An inertia mockup representing the mass and  stiffness properties of the flight deployable was attached to the Axle. Deployment was horizontal, and  gravity offloading was provided by a cable that pivoted about a point on the axis of rotation. The Axle was held in the stowed position by an electromagnet. For the UDM, the sep aration Kick-off Springs were also  represented. Axle position was recorded with an encoder, Telemetry Switches were monitored, and the  final deployed and latched position was measured for repeatability with a dial indicator.  Deployment  Mechanism STOWED DEPLOYEDThermal  Chamber  (Door Removed)Gravity  Offload Cable Electromagnetic  Restraint (& UDM  Kickoff Spring)  Dial Indicator Mockup Inertia (90kg)Safety Cart  Figure 14. Mockup Inertia Margin Test Setup  247
With spring-driven actuators, demonstrating torque margin was more complicated than just turning down  a power supply. The Spring Motor produced torque from two stacks of spring laminations. One stack was  decoupled to reduce the torque output by half. Successful deployment under the worse-case-drag  conditions therefore demonstrated at least 100%  margin. The Latch was similarly handicapped by  removing one of the redundant springs, and 100% margin was demonstrated in a maximum deployment  velocity scenario.  With a new test engineer, and having learned a lot from the previous testing, things went much more  smoothly, but not perfectly. The biggest issues were with the data acquisition system, which still had a  number of bugs and other problems. There was also a lot to learn about constructing a thermal chamber. While thermal regulation with heaters and LN 2 was fairly straightforward, preventing condensation was  not. Water molecules are extremely mobile, and even positive pressure along with purge rates over 15  volume exchanges per hour didn’t eliminate frosti ng of the hardware. Also, directing GN 2 into an LN 2 stream can result in snow as the moisture in the GN 2 freezes.  Conclusions  The Aquarius Deployment Mechanisms enabled the  radar Reflector to be stowed for launch and  deployed on-orbit. While the Mechanisms were  relatively simple latching hinges, the challenging requirements demanded extreme scrutiny of every  design detail. Use of a heritage Spring/Damper  Actuator faced the dilemma of trying to maintain heritage while simultaneously having to adapt to the  new implementation. Unending fabrication,  procurement, and as sembly issues caused test  preparations to be neglected, which  in turn led to  terrible testing difficulties. Despite all the challenges  and setbacks, the Deployment Mechanisms ultimately exceeded all of their performance  requirements. Many important lessons were learned  in the process, as de scribed throughout the paper.  References  1. John R. McGuire and Joseph A. Yura,  “Advances in the Analysis and Design of Constant-Torque Springs”, 30 th Aerospace  Mechanisms Symposium, May 1996, pp. 205-220 2. Robert L. Fusaro, “NASA Space Mechanisms  Handbook”, NASA-GRC, NASA/TP—1999206988, July 1999, pp. 145-152 (damper), pp.  94-95 (telemetry switch) 3. Jack D. Harper, “Viscous Rotary Vane  Actuator/Damper”, 10 th Aerospace Mechanisms  Symposium, April 1976, pp. 198-207 4. “L4 Aquarius Mechanisms Subsystem  Requirements, Initial Release”, JPL Document D-35155 Boom UDM LDM Figure 15. Final Deployment Boom Assembly 5. Alexander H. Slocum, “Precision Machine Design”, Society of Manufacturing Engi neers, Dearborn MI,  Jan 1992, pp. 464-468, 539-543 6. Alan R. Leveille, “Bearings Analysis Program Version 10 C”, Licensed by The Aerospace  Corporation, El Segundo CA 7. Alphonso Stewart, Charles Powers, and Ron Lyons, “Improvements for Rotary Viscous Dampers  used in Spacecraft Deployment Mechanisms”, 32nd Aerospace Mechanisms Symposium, May 1998 248
A Pro/E CAD model of the port-side ETD including the DDM and UM in its closed and latched  configuration is shown in Figure 3. Actuators are not shown. 

UM Linkages  DDM Linkages 

Figure 3. Port Side ET Door Configuration including DDM & UM 

Analysis Approach  During the Space Shuttle mission STS-118 of August 2007, ground telemetry data indicated that the ETD  did not fully complete their travel before the uplatches were commanded to engag e. Because the DDM  actuator has a fail-safe brake that engages when no power is applied, the mechanism is constrained at its input link as the door is forced to close. This constraint at the input link effectively creates deflection in the  DDM linkages during uplatch operation and will therefore induce an associated stress.  

FEA is frequently used to calculate stress for su ch loading conditions. Within this approach, a single  analysis will describe the static loading of a given mechanism configuration. Additionally, the analyst must create a mesh of the entire mecha nism, and then re-mesh for each configuration of interest. This process  can be both labor intensive an d time consuming, and most importantly, the mechanism dynamics are lost  in such quasi-static analysis.   In this paper an alternative modeling approach is proposed for use in the multi-body dynamics software  MSC/ADAMS. This approach utilizes lumped parameter approximations to model the compliance of individual parts within a mechanism. To represent each flexible part, single degree of freedom linear  springs were introduced for links in tension and compression, while torsional springs at the base of rigid  cranks were used for lin ks in bending. With the door drive input crank fixed, the door was then forced to  close, thereby simulating the conditions seen by the actual mechanism. To validate the representative  equivalent stiffnesses, each was iteratively updated until conve rgence using the ratio of FEA stresses  calculated based on spring deflection and spring force.   Modeling  Simulation to Obtain Deflections

Normal operation requires that two sets of indicators, at the uplatch hooks and in the door drive actuator,  are both observed during door closure before running the uplatches. However, during the STS-118  mission only the first set of these indications was initially obtained. Therefore the door was within the  251
capture envelope of the UM (51 mm or 2 in from complete closure), but had not fully reached its intended  latching location. Because the DDM’s actuator brake is on during uplatch operation, the proposed  analytical approach is used to find the induced deflections in the DDM linkages due to this unusually long  uplatch stroke. Based on these deflections, or alternatively the associated spring force s, linkage stresses  can be computed. For elastic deformations, cycling the ETD’s DDM and UM mechanism will return the  door to its final closed configuration with respect to the Shuttle as expected. If plastic deformations have  occurred then the door may form gaps or a step with respect to the Shuttle.   The Pro/E model of the DDM shown in Figure 3 had been previously constructed at JSC and LaRC, and  was used as a starting point for subsequent analysis. This model was imported into ADAMS, with the appropriate piece-parts merged into functional moving parts. As shown in Figure 4, the door rods within  ADAMS are a simplified geometric repre sentation of the actual linkage. For the purpose of the  mechanism analyses, we chose to repre sent these links as rod elements with revolute joints at each end. 

Door Rod  Aft  Outboard Inboard  Fore  Figure 4. ETD Door Drive Mechanism 

For deflection analyses, the linkages of the DDM are represented in ADAMS as spring or compliant members, while the door and shuttle frame are considered rigid. Figure 5 shows each part of the linkage  and its spring equivalence depending on loading: linear springs are used for tension or compression and  torsional springs (at the base of a rigid crank) for bending. For the “follower 1” part, both compression and bending loads exist, thus both linear and torsional springs are used respectively. 

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Follower 1  Rod 1 (tension/comp.)  Crank 1 Follower 1  (bending)  Crank 2  Actuator Crank 

Figure 5. Spring/Compliant Representation of the DDM   To find initial equivalent stiffnesses for use in this lumped parameter model, a representative cross  section of each part was taken near its midpoint and used to find values of cross sectional area (A) for linear springs and area moment of inertia (I) for torsional springs. Linear stiffness was found using  equation (1) for a rod under axial load [2], and torsional stiffness was found using equation (2) for a  cantilever beam with end load, assuming small deflections.   /g187/g188/g186 /g171/g172/g170/g32mmN lEAkeq (1) /g187/g188/g186 /g171/g172/g170/g152/g184 /g185/g183/g168 /g169/g167/g184 /g185/g183/g168 /g169/g167/g32degmmN 1803 32/g83 lEIl keq  (2)

 To run the dynamic simulation, the actuator crank is fixed (motor brake is on) while the opposing edge of  the door is forced to move in the closing direction 25.4 mm (1 in). According to the door’s rigging  specifications, actuation of the DDM must leave the door between 25.4 mm (1 in) and 44.45 mm (1.75 in) of the fully closed and latched configura tion under gravity [3]. Therefore, if the door were to stop short at  50.8 mm (2 in), then the maximum amount of additional displacement would be 25.4 mm, if the door were  rigged to its minimum sag of 25.4 mm. Figure 6 shows this requirement, as presented in the ETD  installation and rigging procedure [3].   Actuator Rod  Actuator Follower  (hidden) Follower 2 Rod 2  Torque Tube  Door Rod  ET Door  253
g  Figure 6. Port-side Door Looking FWD, Door Sag Specification 

Iteration of Stiffness Approximations   If FEA is used to find stress in each compo nent based on the results of dynamic simulation, this  corresponds to a switch from a lumped paramete r model in MSC/ADAMS to a distributed model in  MSC/NASTRAN. While the resulting deflections and forces are analytic within ADAMS, where force and displacement for each link obey Hooke’s law for a given equivalent stiffness, stresses computed in FEA  using boundary conditions (BCs) based on these displacements and forces will be different unless the  equivalent stiffness is the same in both models. In a linear static analysis, stress is proportional to force; therefore the ratio of stresses in these two load cases is equal to the ratio of applied force respectively.  Equation (3) shows this relationship, where the subscript “disp” signifies an FEA model with displacement  BCs, and “force” signifies an FEA model with force BCs. Similarly, subscripts “f” and “ds” indicate FEA and dynamic simulation respectively. After running an initial dynamic simulation and calculating the  associated stresses, k f is the only unknown in equation (3). 

ds dsdsf dsf forcedisp forcedisp xkxk FF FF/g32/g32/g32/g86/g86  (3)

An iterative process is employed to ensure that the equivalent stiffnesses k f and k ds are in fact the same.  By updating the value of k ds, the stress ratio shown in equation (3) can be manipulated. When this ratio  becomes equal to one, k ds is therefore equal to the unchanged value of k f. After an initial set of deflections  and forces is obtained from the ADAMS model, these results are used as the input to a NASTRAN model  of each flexible component. Using the stress ratio described above, a new equivalent stiffness for each  link is computed according to equation (4). This result is then used to update the ADAMS model, thereby completing the loop of a process that can be repeated until convergence.  

][ ]1[ ,,ik ik k ki i forcei disp ds ds forcedisp f /g184/g184 /g185/g183 /g168/g168 /g169/g167 /g32/g14/g159/g184/g184 /g185/g183 /g168/g168 /g169/g167 /g32/g86/g86 /g86/g86  (4)

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