ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
248 In order to answer the question of minimum acceptable guide length and material choice, manual testing  was performed using the test assembly shown in Figure 12. The plates at the end of the assembly determined the radial clearance to the housing bore and the plates in the middle region controlled the effective length of the test guide.  

Figure 12.  Test Guide Assembly - Used to qualitatively determine  the best length and radial clearance combination using Delrin ® disks   The test guide assembly was inserted into the test housing and the operator used the mandrel diameter to stroke the assembly within the bore. As the test guide was moving, the operator applied a moment to the assembly with the thumb and index finger. As moment was applied, some combinations of length and radial clearance would solidly jam, some would tip over freely, and others would get tight in the bore but not be self-energizing. The results of the testing are listed in Table 1.   Table 1.  Results of Qualitative Manual Testing of Test Guide Assembly  as a function of radial clearance and guide length using Delrin ® disks 

  Clearance, Ø [mm]  Length [mm] 0.025 0.127 0.254 0.381 0.508 0.635  6.35 FAIL - FLIP FLIP FLIP FLIP FLIP  12.7 FAIL FAIL - FAIL - FAIL- FAIL/FLIP FLIP  19.1 FAIL FAIL - FAIL FAIL  FAIL FAIL-  25.4 FAIL FAIL - FAIL FAIL+ FAIL+ FAIL  31.8 FAIL FAIL FAIL+ FAIL + FAIL+ FAIL+  38.1 PASS - PASS - PASS PASS PASS PASS -  44.5 - - - - - PASS  FLIP = Piston could freely flip through 360 degrees, i.e. provided no tip restraint  FAIL - = Jammed badly, i.e. became stuck with little or no effort  FAIL = Would jam, but only when a moment was applied or was pushed at edge FAIL + = Would jam, but took a lot of effort.  Would not recover on its own  PASS - = Judged not to jam, but considered to be closer to jamming than others  PASS = Could not be made to jam
249 The manual testing indicated that a 38.1 mm minimum guide length was required; no length/clearance  combination less than 38.1 mm was successful. However, the maximum possible length of the guide that could be implemented with the housing and structural constraints was 25 mm. The test results demonstrated that a Delrin ® solution was not acceptable.   Since the longest possible guide length was 25 mm, a new set of tests were implemented with different guide materials and two different guide designs. The first test guide design was a solid cylindrical guide that was 25 mm long on a stainless steel mandrel. The second design was the same cylindrical guide with a mandrel and a slotted guide, with the slot in the circumference of the guide, similar to automobile engine piston ring designs. The slot allowed the CTE mismatch between the guide and housing to be accommodated in the expansion/contraction of the slot, rather than in the radial clearance between the guide and housing. This was crucial for high CTE materials, since a solid guide design would have required the guide to actually be smaller in diameter than the bellows themselves. The slotted guide was also placed on a stainless steel mandrel to provide support for the bore of the guide. Figure 13 shows the virgin Teflon ® version of the guide.        

Figure 13.  Solid and Slotted designs of the guide on Test Mandrel and inside the Test Housing  with the largest possible radial clearance. Note that the slotted design springs outward, closing  the radial clearance under all conditions.  (the terms Solid and Slotted refer to the circumference of the guide)    The different materials tested using circumferentially solid and slotted guides on a mandrel were    Torlon ® 5030   Vespel® SP-3   Delrin® 100 AF   Teflon® 25% glass filled   Virgin Teflon®   The testing was done manually with the same pass/fail criteria as was used in the previous Delrin ® guide  testing (see Table 1). The test guides were fabricated to diameters representing a nominal radial clearance at 25°C and the maximum radial clearance associated with the lowest temperature condition and the smallest manufactured outside diameter (Least Material Condition, or LMC). The results of this testing are listed in Table 2. Note that a solid Delrin ® version was not tested. This was the version tested  previously and represented in Table 1. For the Torlon® and Vespel®, the slotted versions were not tested
250 because the CTE of these materials did not require them to be slotted and the solid versions failed. Both  Teflon® versions were successful, except that the solid version with virgin Teflon® had too high a CTE and  lost all guiding function at cold temperature and least material condition.   Table 2.  Results of Qualitative Manual Testing of Test Guides using circumferentially  slotted and solid test items as a function of radial clearance [Nominal and Least Material  Condition Cold (LMC Cold)], and material 

 The test results demonstrated that both glass filled and virgin Teflon ® were acceptable candidates, though  only a slotted design was acceptable with the virgin Teflon®. Additionally, the glass filled Teflon® showed  signs of abrading the bore of the housing by capturing fine particles of housing material within and on the wear surfaces of the guide. This abrasion was deemed undesirable and the virgin Teflon ® slotted design  was chosen to proceed forward. The final design of the guide and hard stop for the flight assembly is shown in Figure 14 along with a representation of the original design, emphasizing the changes made to mitigate the demonstrated failure mode.      Figure 14.  Original and Final designs of the end of the bellows assembly. The figure shows the  bellows stroked to the maximum fluid condition and against the end-of-stroke hard stop for both  design versions    Material    CTE Solid  Slotted       [ppm/C] nominal LMC Cold nominal LMC Cold Torlon 5030 glass fill 16 FAIL        Vespel SP-3 MoS2 52 FAIL        Delrin 100AF Teflon fill 100     FAIL    Teflon 25% GF glass fill 100 PASS PASS PASS PASS  Teflon, virgin   151 PASS FLIP PASS PASS  Old Hard Stop New Hard Stop  added  guide length
253 Development of the Descent Brake Mechanism for the Mars Science Laboratory 

David Dowen*, Jeff Moser** and Jeff Mobley** 

Abstract   This paper will describe the design and testing of an electromechanical damper assembly for the Mars Science Laboratory as well as provide the current program status. Unique test equipment was developed for verification of flight requirements and is presented. Included in this paper are the problems that arose during design and testing that are peculiar to a device intended to dissipate energy. Also discussed are the lessons learned relating to assembly and test anomalies and the resulting corrective actions.   Introduction 

The primary mission of the Mars Surveyor Lander (MSL) program is to deliver the 900-kg Curiosity rover to the surface of Mars in the year 2012. Its “Skycrane” landing system uses a free-flying rocket-decelerated descent stage to set Curiosity down on the surface, which hangs on a 7.5-m-long bridle cord, without ever landing the descent stage or fully powering down the engines. This architecture enables safe delivery of Curiosity onto rugged terrain. A key component of the Skycrane landing system is the device which allows the rover to be lowered by a distance of 7.0 m to fully extend its bridle. The Descent Brake allows this deployment to happen quickly and with controlled speed that decreases toward the end of deployment for a soft stop at the full bridle length. The amount of energy dissipated by the Descent Brake during the event is approximately 24 kJ.  

  Descent speed control is achieved passively by deploying the bridle cord from a tapered spool. The tapered spool allows an ending descent speed that is less than the average speed. The radius from which the cord deploys is decreased as the cord is deployed. Initially the vertical velocity of the rover is high relative to the angular rate at the input shaft. As the bridle deploys, the diameter of the spool decreases and the angular rate, at the shaft, increases relative to the rover’s vertical rate. For this design, the torque is initially low at the Descent Brake, and then rapidly increases to slow the vertical rate of deployment, reference Figure 2. The design chosen to manage the drop converts the rover’s potential energy into heat with an electrical generator, driven by the bridles wrapped on the spool and dissipating the rover's energy into a bank of resistors. This technique has the advantage of being completely passive, and unlike previous friction technologies, offers highly consistent performance over a wide temperature range. 1                                                           * Sierra Nevada Corporation, Space Systems Group, Louisville, CO  ** Sierra Nevada Corporation, Space Systems Group, Durham, NC   Proceedings of the 41st Aerospace Mechanisms Symposium, Jet Propulsion Laboratory, May 16-18, 2012  Figure 1 - Sky Crane Landing Sequence
254  

Figure 3 - MSL Deployment Mechanism 

Descent Brake Requirements   Damping Performance  The initial nominal performance specification for the MSL Descent Brake required a nominal linear speed dependent drag of 10.6 ± 2.3 N•m•s/rad over a temperature range of -35°C to +35°C.  Figure 4 shows the performance envelope allowed as it relates to shaft speed and torque. The performance of the Descent Brake design needed to remain in the performance envelope over all variations in manufacturing tolerances and environments.  Descent Brake  Spool Device  Confluence Point Bridle Guide Umbilical Device  3X Bridle Exit Guide  Figure 2 - Output from MSL Lowering Device Simulation1
255  

Figure 4 - Descent Brake Performance Requirement  Design Limits  Several other requirements key to the design were specified as design limits. These requirements from  the initial specification were:   Output shaft peak speed of 28 rad/s    Instantaneous output shaft peak torque of 340 N•m    Event duration of 7 seconds maximum   Power dissipation capability of 9.5 kW    Approximate envelope of Ø265 mm (Ø10.43 in) maximum outer diameter by 230 mm (9.05 in)  maximum length (circular form factor) 

Descent Brake Design   The initial step in the system design of the Descent Brake was to evaluate the driving requirements and establish design goals. The peak power requirement of 9.5 kW drove the design to require the majority of the power to be dissipated by a resistor bank. To ensure consistent performance with variation in temperature the resistor bank was designed around high reliability wire wound resistors with very low change in resistance over temperature (~±20 ppm/°C). In order to minimize the generator winding copper losses (I 2R losses) it was desired that the generator winding current be minimized. To minimize the  generator winding current, the generator voltage was designed to be as high as practical. In order to avoid corona issues, the maximum generator design voltage amplitude of 150 V peak was chosen.   To meet the desired performance requirement of nearly linear damping over the speed range, it was critical that the inductive reactance component of the generator winding impedance be minimized. The inductive reactance component of the generator impedance is a function of the generator winding
256 inductance, the number of generator poles selected and the shaft speed. The shaft speed was  determined from the gear ratio selected. The gear ratio was selected to be as high as practical to balance the generator and gearbox sizes while still surviving the specified load conditions with margin.   

   Late in the design process several of the design limits changed as the Curiosity rover design matured. The damping ratio was modified to 9.45 ± 0.15 N•m•s/rad. This change in damping ratio resulted in the peak momentary torque rising to 583 N•m and independently the peak speed to 36 rad/s. The Descent Brake design was evaluated for performance to these new requirements and fortunately was able to meet them by analysis except for the new maximum speed requirement. Operating at the new maximum speed requirement drove the rotor speed to over 1750 rad/s. This new rotor speed was 20% over the manufacturer’s recommendations for the rotor bearings. A test rotor with bearings was used to show the bearings were capable of withstanding operation at the maximum speed requirements with no issues. The rotor bearing set was tested to speeds that equate to 40 rad/s at the gearhead input shaft. The bearings were inspected at SNC and by the bearing manufacturer after completion of the testing with no detrimental affects noted.  

Descent Brake Components   Generator  From the established generator parameters, the generator design space was evaluated. Several generator designs of various geometry and number of poles were created. Each design was optimized to minimize inductive reactance while meeting all other parameter requirements. A Simulink model of the Descent Brake system was also created. The Descent Brake performance was evaluated for each generator design to determine the best generator configuration to meet the design requirements stated above using the system model. The generator configuration that ended up with the lowest inductive reactance was a 6 pole generator with a 1.1:1 length to diameter ratio. The chosen design showed very near linear operation of the generator damping over the required speed range.  6X Resistor Boards  Jumper Board Output Spline Interface Gearhead  Generator  Figure 5 - MSL Descent Brake Assembly
257  

Gearbox  The gearhead is a high torque density, 3 stage planetary gearbox with an overall gear ratio of 48.8:1. The gearbox mass is 3.73 kg within an envelope of 98.55-mm diameter by 91.44-mm length (excluding flanges and shaft extension). Gearhead capacity was optimized using the gearbox design guidelines developed through the NASA Phase II SBIR Lightweight Gearbox Technology Program 2, including  material selection, design features, and analysis techniques. The gearbox was designed for an operating high torque load case (322.5 N•m at 25 rad/sec), an operating high speed load case (298.8 N•m at 36 rad/sec), a momentary load case (583 N•m at 18 rad/sec) and a static peak torque case (700 N•m). The required life of the gearbox is only 8 cycles of 120 radians (960 radians total) for the operating load cases with the momentary load occurring no more than twice per cycle. With the relatively short life requirement, the gearbox was designed to provide positive margin for the specific loads and duration specified rather than being designed for endurance limit. Because of the high loads and speeds involved, significant attention was paid to imbalance of loads between the planets and the resultant net load applied to the supporting bearings. At the high speeds required, the centrifugal force on the planet bearings was factored into the analysis in addition to the resultant gear forces. Thermal impacts upon the lubricant during high speed operation were also considered. 

Figure 6 - 3 Phase Generator Stator Assembly  Figure 7 - First, Second and Third Stage Carrier Assemblies
258 Resistor Bank/Structure  The design of the Resistor Bank for the Descent Brake was a challenge in packaging. The Resistor Bank  consists of three 1.5-ohm phase resistances wired in a wye configuration. Each phase consisted of two Printed Wiring Boards or PWBs containing 25 resistors on each board. The resistors were set in series/parallel configuration to result in the required phase resistance. This number of resistors was also utilized to ensure the maximum power dissipated in any resistor never exceeded 10 watts. An additional PWB was used to make all the connections between the Resistor Bank and Generator Assemblies. The resistors were arranged in a radial pattern on the donut shaped PWBs due to the circular form factor of the required envelope. The PWBs were mounted onto aluminum chassis components that were optimized for weight under the required structural loads. Individual resistors of different values could be chosen to tune the damping of the assemblies as required to meet the desired requirement in the range of approximately 7.0 N•m•s/rad to 12.0 N•m•s/rad. 

Figure 8 - Resistor Bank Assembly      Test Program/Equipment   Functional testing of the Descent Brake can be thought of as essentially dropping the equivalent of a Mini Cooper from the roof of a second story building on a spooled cable system. Although the requirements could be met with this type of test setup it was not ideal for many reasons. The functional testing needed to be performed in many different configurations including in a thermal vacuum chamber. The test method chosen was to use a brushless DC servo system to supply the required torque at the Descent Brake input shaft and a reaction torque cell to measure reacted torque at the mounting flange. The DC servo was sized to be able to meet all torque requirements without any gearing to keep sources of error at a minimum. This type of test setup allowed for a flexible test rig that could be adapted to many different scenarios. The DC servo system also allowed for flexibility in test profiles with only changes in programming. Figure 9 shows three of the test setups used for in-process, acceptance and qualification testing. The upper left photo shows the test configuration used for component level verification and resistor bank tuning. An additional torque cell is added between the generator rotor shaft and input to the gear head.   This allowed for easy measurement of gear head efficiency and generator performance in the system. The upper right photo shows the ambient test configuration for a fully assembled Descent Brake. This
259 setup was extremely robust and allowed for consistent test results from the Descent Brake. The bottom  picture shows the test configuration for thermal vacuum testing. This setup is essentially the same as for ambient except the addition of a fluid coupling to pass torque through the thermal vacuum chamber wall. Although the fluid coupling does have some drag associated with it, the magnitude is small compared with the torque input so had little affect on the overall test results.  

   The tests performed on the flight Descent Brake included the following:   Structural/Stiffness test   Initial ambient functional test   Vibration test   Ambient functional test   Thermal cycling   Functional testing at thermal extremes   Final functional test  The flight unit completed all acceptance testing without issue. The measured damping rates for the functional tests performed were 9.51 N•m•s/rad during initial ambient, 9.42 N•m•s/rad during thermal vacuum hot extreme, 9.51 N•m•s/rad during thermal vacuum cold extreme and 9.36 N•m•s/rad during final ambient functional testing. The torque vs. speed curves for these tests are shown in Figure 10 and shows the consistency of operation under varying environmental conditions.  Component Verification Configuration Ambient Test Configuration  Thermal Vacuum Test Configuration  Figure 9 - Performance test setups
260  

The Qualification Unit completed the same tests as the Flight Unit with the addition of shock, more extreme temperature limits during thermal vacuum and life testing. All tests were completed with no issues. After completion of the life test the qualification unit was disassembled and all parts inspected. No issues were found during the inspection.   Lessons Learned 

Collaboration and Robust Design Margins  One of the biggest challenges on the program was reacting to changes in requirements throughout the program and in particularly after completion of the design phase of the program. At the point the last requirement changes occurred there was a lot of schedule pressure to complete the program to support launch date at that time. Fortunately, the gearbox and resistor bank were designed to take advantage of the required envelope and had adequate design margins. We were able to perform analysis to the new requirements and show that the gearhead and resistor were still compliant. The generator rotor bearings were the only components we could not show compliant to the higher limits via analysis. This verification was handled through empirical testing and turned out to be a nonissue. The highly collaborative environment cultivated between JPL and SNC throughout the program was instrumental in working through the requirement changes effectively. This type of relationship is a necessity in order to effectively work through a program with dynamic requirements.  Robust Test Set  Besides the design of the deliverable hardware, the test set was also a challenge. The size of the Curiosity rover drove torque and speed test requirements outside our normal range. Using the large direct drive DC servo system worked extremely well. The combination of size and flexibility allowed us to perform many different tests in essentially the same test setup.  Development Test Anomaly  The generator rotor is composed of three main parts, a stainless steel rotor hub, permanent magnets and a thin metal rotor band. The permanent magnets are bonded to the stainless steel rotor hub and the rotor band is thermal fitted over the outside diameter of the permanent magnets. The function of the rotor band is to provide secondary mechanical retention and to protect the magnets within the air gap between the rotor and stator. Under nominal conditions there should normally be 0.25 mm (0.010 in) clearance, typical for this type of device, between the rotor band outside diameter and the stator inside diameter. During  Figure 10 - Flight Acceptance Performance Curves