ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
261 initial testing of the development generator, abnormalities in the generator torque were noticed during inprocess run-in. The generator was disassembled and it was found that the rotor band had been rubbing the generator stator inner diameter, reference Figure 11.    In order to determine the cause of the rubbing, all the dimensions were verified and a rotor deflection analysis was performed. The result of the analysis was that everything appeared correct. The only abnormality noticed in the test data was that the no load generator torque measurements indicated higher than expected torque. No load generator torque measurements are an indication of the losses in the generator magnetic core and bearings. From the higher than expected generator torque, it was determined that approximately 50 W of power was being generated in the rotor band at the higher speeds due to varying magnetic field. In a normal generator design, the magnetic field in the rotor band does not vary to the extent that any significant losses are generated in the rotor band. However, the optimization done to the generator stator design, in order to minimize inductive reactance, created larger than normal magnetic field variation in the rotor band. The rotor band is a thin metal band which is primarily in direct contact with the rotor magnets. The rotor magnets are Samarium Cobalt, which has relatively low thermal conductivity. Due to the relatively low thermal conductivity of the magnets, the 50 W of power generated in the rotor band caused the rotor band temperature to increase. Due to this, the band expanded significantly which caused it to rub the generator stator inside diameter. Normal operation duration for the Descent Brake is seven seconds, which does not generate enough heat to cause the interference. This issue did not show up until the run-in was performed for 1 hour in each direction at 6 rad/s. 

 Since the rotor band is not exposed to significant varying magnetic field in a normal generator design, it is typically made from a material that has appropriate mechanical properties for the application. In this case, the material initially chosen was beryllium copper. Power losses due to magnetic fields are a function of the electrical conductivity of a material. Beryllium copper has an electrical conductivity of approximately 20% that of pure copper. To solve the band power dissipation issue, the beryllium copper rotor band was replaced with a one made from Inconel. Inconel has an electrical conductivity of approximately 2% that of pure copper. The Inconel band resolved the power dissipation issue.     Development Rotor with BeCu Band  Development Rotor After Run-In   Development Rotor with Inconel Band  Development Stator After Run-In  Figure 11 - Rotor Development Issue
262 Although the operation time of the Descent Brake is relatively short, understanding all the environments  the device is going to be used in is critical to achieving a robust design.  Commercial Components  Another issue encountered in the manufacturing process was the use of commercial needle roller bearings. Roller bearings were implemented in the design to support the planet gears on the second and third stage carrier assemblies due to the high radial load and speed requirements at these locations. These bearings were procured from a commercial vendor using their standard materials. The 440C Stainless Steel rollers were fine, but the carbon steel cages had corrosion issues even though they were coated with a proprietary silver plating. Strict corrosion prevention process had to be implemented to ensure parts with corrosion did not end up in the assembly.   Conclusions 

Ultimately, the program achieved successful completi on of validation testing of the Descent Brake design  for the qualification unit. High level collaboration between JPL and SNC led to the successful assembly and test of the flight unit. The flight unit has been integrated into the rover and completed system level testing. This included a system level drop test in January, 2011. The total drop time during this test was within 0.1 second of the predicted time. The Curiosity rover is currently in transit to Mars, it was launched November 26, 2011, and is scheduled to land on Mars on August 6, 2012.   Acknowledgments 

This work was funded by JPL under Subcontract 1293381. The authors wish to express appreciation to Ted Iskenderian, JPL Mechanisms Engineering Manager, for his work on the Descent Brake and their assistance in the production of this paper. BUD and Curiosity images courtesy of JPL.   References   1. Gradzial, M.J. and Holgerson, K.J.  “Mechanisms for Lowering Tethered Payloads: Lessons Learned  from the Mars Exploration Program”  IEEE Aerospace Conference Paper No. 1030, Big Sky, MT, March 2008  2. Mobley, J. “D21507 Final Report: Lightweight Gearbox Technology Program Phase II SBIR” SNC
263 Mars Science Laboratory Differential Rest raint:  The Devil is in the Details 

Elizabeth Jordan* 

Abstract 

The Differential Restraint, a mechanism used on the Mars Science Laboratory (MSL) rover to maintain  symmetry of the mobility system during the launch, cruise, and entry descent and landing phases of the MSL mission, completed nearly three full design cycles before a finalized successful design was achieved. This paper address the lessons learned through these design cycles, including three major design elements that can easily be overlooked during the design process, including, tolerance stack contribution to load path, the possibility of Martian dirt as a failure mode, and the effects of material properties at temperature extremes.    Introduction 

The Differential Mechanism, a series of linked, passive pivots, is a component of the Mars Science  Laboratory’s rocker-bogie mobility design. The differential acts as a motion reverser for the suspension system. The Differential Restraint was designed to prevent the rotation of the Center Differential Pivot (see Figure 1) of the Mars Science Laboratory mobility system during the launch, cruise, and Entry Descent and Landing (EDL) phases of the MSL mission.   The Mars Science Laboratory is JPL’s next generation Mars Rover, known to the public as “Curiosity”. The mission launched on November 26, 2011 on an ATLAS V rocket and is scheduled to land on Mars on August 5 th, 2012. The rover mobility system incorporates the heritage rocker-bogie suspension design  which was invented at JPL for the first Mars rover missions. The large size of the rover, approximately nine feet in length, requires the mobility system to be folded during the cruise stage of the mission. During the Skycrane phase, in which the rover is lowered to the ground via bridles and a descent stage (see Figure 2) the rocker arms of the mobility system are released from their stowed positions and allowed to rotate about passive pivot joints until the four rockers latch in their ready-for-touchdown positions. The Differential Restraint is the last of the mobility restraints to be released, via a pyro, approximately 2 seconds before rover touchdown.                                                    * Jet Propulsion Laboratory, California Institute of Technology , Pasadena CA 

Proceedings of the 41st Aerospace Mechanisms Symposium, Jet Propulsion Laboratory, May 16-18, 2012  © 2012 California Institute of Technology. Government sponsorship acknowledged.
265 Differential Restraint Design Development and Description 

The Differential Restraint is a passive mechanism which relies on the compression of a series of Belleville Springs to limit the degree of rotational motion allowed at the Rover Mobility Center Differential Pivot during the Cruise and EDL phases, as well as to absorb and dampen the energy associated with the violent mobility deploy and latch events. Since the rover mobility system is also the vehicle’s landing gear it is critical that the 6 wheels are in, what is termed, their “ready-for-touchdown” state, or as close to 6-wheels flat as possible when the vehicle impacts the Martian surface. The Differential Restraint enables this goal by limiting the rotation of the passive center pivot joint, known as the Center Differential Pivot (CDP), to a small angle, keeping the two sides of mobility, port and starboard, nearly balanced. Once the chassis dynamics, namely, chassis pitch and roll from the effects of the mobility deploy and latch events, have settled out the Differential Restraint is released, via a pin-puller pyro just prior to rover touchdown.  Design History  The driving load case for the Differential Restraint is derived from the near-simultaneous deployment (and end of travel latch events) of the mobility rockers during the Skycrane phase of EDL. In the event of actual simultaneous latch events of the port and starboard sides, large shear loads are expected at the CDP. Alternatively, if the deployment and latch events are staggered from port to starboard large moments are created about the CDP, which the Differential Restraint must absorb. The original Differential Restraint design was a rigidly pinned interface between the externally mounted mobility Differential and the top deck of the rover chassis which prevented nearly all motion, except for the small amount of slop between the pin puller pin and interfacing monoball (see Figure 3).   Approximately one year after the mobility Critical Design Review the increasing maturity of the ADAMS dynamic model produced a drastic increase in the predicted moment and shear loads at the CDP. This had the effect of creating significant negative margins in the original piece parts of the Differential Restraint; including the AerMet pin of the 3/8” Pyro, the shear pins and bolted joint which interfaced to the Rover top deck, as well as the Rover top deck itself. In late 2008 it was concluded that a completely new restraint design was required. The new design was to be analyzed, built, tested and integrated before the (then) planned 2009 launch. In addition to the tight timeline constraints, the re-design effort was further complicated by the fact that the hardware interfaces were already fixed, the available space was incredibly limited, and the 3/8” pin puller was the only feasible option for pyro devices.
266   Figure 3: Original, Rigidly Pinned, Differential Restraint Design Concept  Since the rover top deck interfaces were already built, and the pyro device was already qualified and  delivered the new design had to find a means by which the loads that the Center Differential Pivot would impart on these critical interfaces could be significantly decreased. The new design relied on the principle of applying a (theoretical) torsion spring at the center of the Differential Pivot which would absorb the large impact energy. After running idealized dynamics simulations, a spring rate of approximately  200,000  N-m/rad was selected as the stiffness that would reduce the impact loads enough to prevent  negative margins in mobility and chassis hardware but still restrain the mobility system from excessive motion during launch, cruise and the initial stages of EDL enough to be able to ensure that the mobility system was in its “ready-for-touchdown” state in time for landing. The new restraint was designed to accommodate the worst case combined loading event as predicted by a 2000-run Monte Carlo ADAMS dynamic model of the MSL EDL system.    Implementing a torsion spring at the Center Differential Pivot (CDP) was impossible due to packaging and size issues so it was decided that the Differential Restraint would use a 3-bar crank-slider approach to turn rotation into linear actuation (see Figure 4). Combinations of large Belleville washers in series and parallel were implemented to achieve the extremely high spring rate that was required in the limited space available. A rod, known as the Spring Plunger, wa s used to actuate the two stacks of Bellevilles. One  stack is compressed from positive moments while the opposite stack is compressed by negative moment loads at the CDP (see Figure 5). The Spring Plunger, made from Titanium, was threaded at one end to allow it to be joined to a Clevis, which in-turn was connected to a Linkage with a Monoball in each end (see Figure 6 and Figure 7).
268   Figure 6 : Cross Section of Differential Restraint Mechanism 

Figure 7: Differential Restraint Mechanism Nomenclature Overview 

Tolerance Stacks and Stress Analysis 

Test Failure  With the delay of the MSL launch the new Differential Restraint hardware was built but the planned component level testing of the hardware was delayed in an effort to conserve funds. In July of 2009, the Differential Restraint was integrated to the mobility differential system (see Figure 1) on the Dynamic Test Model of the MSL rover, in preparation for the differential system static test. With the restraint in place, the test was designed to load the entire differential system and generate a large moment load about the  Threaded Interface    Clevis
269 Center Differential Pivot by pushing on an additional lever arm connected at the Main Differential Pivot  Interface. At 90% of the full test load the measured load began to drop rapidly and an audible snap was heard. Upon investigation it became apparent that the Spring Plunger had catastrophically failed at the point of thread termination at the interface between the Spring Plunger and the Clevis (see Figure 8).       

Figure 8: Failed Differential Restraint (Left) Disassembled from Clevis (Right)    Determining the Cause of Failure  Since the component level tests of the Differential Restraint had been skipped a wide array of failure causes existed that needed to be examined. Linear sliding mechanisms are often avoided at JPL on principle. This fact made the theory that one of the sliding surfaces between the Spring Plunger, Internal Guides, and Bellevilles had jammed, thereby preventing actuation of the mechanism, a prime suspect. Another theory was that the Spring Plunger, which was made from Titanium, had been made by a fraudulent vendor, and that imperfections in the material caused the crack initiation. The lack of a thread relief callout on the Spring Plunger drawing was also considered as a possible cause of failure, as was the large bending load on the Spring Plunger.             The leading failure theory, that the sliding surfaces had jammed and prevent motion, was disproven in two ways. First, a line of Braycote, the lubricant used inside the Belleville Spring Assembly, was left on the Spring Plunger from where it had entered the Belleville Spring Assembly. By measuring the distance between the grease marks left on the Spring Plunger, it was possible to approximate the distance traveled between the start of test and the time of failure to be 5.46 mm (0.215 inch). At the load applied at the time of failure the total compression of the Belleville Spring stack was predicted to be 5.59 mm (0.220 inch). The close proximity of these values was strong evidence that the Spring Plunger did in fact slide along the Internal Guides and compress the spring stack as intended. Furthermore, the shape of the load deflection curve measured during the test exactly matched the prediction (see Figure 9), including the inflection point where the stiffness of the Belleville Stacks drops by one-half due to only one stack being compressed after the initial motion during which both stacks are in the load path. If the mechanism had  jammed the applied load versus displacement would have spiked significantly rather than following the predicted linear line before dropping off when the Spring Plunger yielded and eventually snapped.      Failure Theories:  Jammed sliding surfaces  Fraudulent Titanium  Improper thread termination on Spring Plunger  Excessive bending load on Spring Plunger
270   Figure 9: Applied Load vs. Pull-arm Displacement measured during failed Static Test 

Upon inspection of the material certs for the Spring Plunger it was confirmed that the material was from Western Titanium. However, upon SEM inspection of the failed threads (see Figure 10) it was determined that the failure was ductile and there was no evidence that a material imperfection initiated the crack.    

Figure 10: 10kX magnification of Spring Plunger Fractured Surface  Bending Loads   The Spring Plunger was originally analyzed for a bending load which would be applied as the Center Differential Pivot rotated. Since the Differential Restraint mechanism was designed to limit the rotation of the Center Differential Pivot to ±2°, the total misalignment between the line of action of the Linkage and the Spring Plunger was calculated by assuming a worst case rotation of the CDP to be 5°. If the CDP were allowed to rotate 5 degrees the misalignment between the line of action of the Linkage and the  Point of  Failure  Expected  change in  stiffness
271 Spring Plunger axis caused by this rotation would be 0.5°. With these assumptions the Spring Plunger  was shown to have very small, but positive, margins of safety at the root of the threads. What was thought to be a very conservative analysis, (doubling the expected rotation of the CDP) was in fact not nearly conservative enough.  

Figure 11:  Line of action misalignment due to rotation of Center Diff. Pivot    In the post failure analysis the total angular misalignment between the Linkage force vector and the Spring Plunger was further scrutinized. It quickly became apparent that the misalignment caused by the rotation of the CDP was actually one of the smallest contributors to the overall misalignment. More significant misalignments were due to piece part misalignments due to imperfect shimming of the mechanism, the vertical slop of the Linkage in the mouth of the Clevis and Horizontal Swingarm Bracket, the slop of the Center Differential Pivot in its own bushings, and the elastic deformations of the Chassis Bracket and the Rover Chassis top deck. The maximum possible misalignment was calculated to be 2.2 degrees (see Figure 12), which was over four times what was initially accounted for in the Spring Plunger stress analysis.       +X  +Y
272   Figure 12: Differential Restraint Angular Misalignment Calculations 

Recovery From Failure  Immediately following the failure of the Differential Restraint many senior engineers began taking a closer look at the restraint design. They highlighted the fact that the restraint poorly implemented the concept of a crank-slider because the point at which the crank grabbed the slider was not at the center of the slider but rather it was offset- the link between the Clevis and the Spring Plunger was at the end of the Spring Plunger rather than the middle of the Spring Plunger (see Figure 13). Many engineers advocated redesigning the restraint to properly implement the theory of the crank slider; however limited schedule and funds made a complete redesign impractical. Instead, the design concept and most of the original hardware was kept intact, but key components were changed from Ti-6Al-4V material to significantly higher strength steel. By remaking two components and modifying two additional ones it was possible to quickly and relatively cheaply make the restraint design work.  ∆X ∆Z ∆RX ∆RZ [in.] [in] [ °] [°] 0.00458 0.02 0.02 0.0233 0.0094 0.020518 0.03211 0.022 0.013 0.0448 0.0031 0.0053 0.629 0.0822 0.0045 0.0290 0.0188 0.0442 0.629 0 0.0822 0.629 0 0.08220.07492.1 RSS X_Z 3.0 3.3 Clevis Elastic  Deformation 3.4 Spring Plunger Elastic  Deformation Angular Misalignment (1.0,2.2,3.0) 2.2154 0.71120.7112 Angular Misalignment (1.0,2.1,3.0) 2.06143.1 Chasis Bracket Elastic  Deformation (8443 lbf) 1.2 HAS Shim 1.3 Chassis Bracket Shim 1.4 Monoball to HSA 0.02555 RSS (1.0,2.2,3.0) 0.08340.0084 3.2 Sum_X_Z 0.0335 RSS (1.0,2.1,3.0)Angular Misalignment  (Include RSS where applicable) 3.2 Top Deck Elastic  Deformation (8443 lbf)RSS1.0 1.0  RSS_X_Z 0.03811 3.1 Sum_X_Z1.5 Monoball to Clevis 2.02.1 CDP TranslationMisalignment Source 2.2 CDP Rotation about X/Y1.1 Sprng Stack Deflection  (.215" )
273 Using the interaction formula:   ௕ଷయൌ1  ݏݏ݁ݎݐݏ	݁ݐܽ݉݅ݐ݈ݑ/	݈݀݁݅ݕ	݋ݐ	ݏݏ݁ݎݐݏ	݁ݒ݅ݏݏ݁ݎ݌݉݋ܿ	݂݋	݋݅ݐܽݎ   	ݏݏ݁ݎݐݏ	݁݉݅ݐ݈ݑ/݈݀݁݅ݕ	݋ݐ	ݏݏ݁ݎݐݏ	ݎ݄ܽ݁ݏ	݂݋	݋݅ݐܽݎ   ݏݏ݁ݎݐݏ	݁ݐܽ݉݅ݐ݈ݑ/	݈݀݁݅ݕ	݋ݐ	ݏݏ݁ݎݐݏ	ܾ݃݊݅݀݊݁	݂݋	݋݅ݐܽݎ   it was possible to predict the allowable side load, and hence the misalignment angle between the Spring  Plunger and Clevis that a steel Spring Plunger could withstand, if the Flight Limit Load was applied axially. After calculating the misalignment angles that would produce yield and failure for a Spring Plunger made from steel it was decided that the values calculated were too close to the worst case possibilities predicted by the tolerance analysis to make simply changing the material a comfortable approach. Therefore, in addition to changing the material, the Spring Plunger diameter and thread diameter was increased. With the larger Spring Plunger diameter, and the material swap from Ti-6al-4v to steel, the new Spring Plunger would be able to withstand 4.24° and 4.29° misalignments before yielding and ultimately failing respectively. Additionally, a thread relief was added to the Spring Plunger thread termination (see Figure 14) to prevent excessive stress concentrations.       

Figure 13: Comparison of “proper” crank-slider design methodology vs. Differential Restraint  Design 

Figure 14: Spring Plunger Thread Termination Callout  Key Findings:   Tolerance stacks and shimming procedures affect load paths- Make sure to account for this in  stress analysis   The location of pivot joints in kinematic mechanisms is extremely important- Failing to follow best  practices early-on can cause unexpected difficulties down the road   Proper thread relief callouts on drawings is essential for limiting the effect of stress concentrations