ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
156 Rollers were chosen primarily to mitigate the potential for jamming. Jamming will occur if the torque applied  by the actuator cannot overcome the frictional forces developed within the chuck mechanism in its  preloaded, engaged state. The use of rollers significantly reduces the friction present at the interface with  the chuck balls. The rollers are mounted onto flexures machine d into the cam. These flexures provide a  compliant connection between the rollers and cam, allowing for better load distribution amongst the balls  as well as reducing changes in preload due to temperature induced dimensional changes and part and  assembly dimensional deviations.  

Figure 7: Cross sectional detail view s of cam roller s, both in the engaged and disengaged states .  

Flexures and Load Distribution   Another concern with preloading balls is the potential for uneven load distribution amongst the balls. The  number of balls in the chuck  was determined by performing an analysis for contact stress under off -nominal  load cases such as rover slip and bit pull out. The goal of the analysis was to determine the minimum  number of balls needed to avoid brinelling at the balls and interfacing parts. The analysis  indicates that six  balls is the minimum needed to react all design loads without yielding assuming a fairly even and predictable  load distribution. However, once machining and other errors are accounted for, the load distribution is  unlikely to be even or predictable and could result in very high loads being applied to a small number of  balls.  

To mitigate high ball loading, the cam rollers were mounted onto flexures cut into t he cam. The flexures  allow for a small amount of compliance at the cam rollers, allow them to take up any difference in relative  radial position due to machining and assembly errors. This helps to spread the load more evenly amongst
157 the balls, resulting in lower maximum ball loads. The flexures were sized so as to provide enough  compliance to spread out loads amongst the balls but be stiff enough to allow for a very stiff bit -to-chuck  connection to obtain good quality cores.  The cam also includes hardstop p ins which serve to limit total  flexure travel. Under pure tensile axial loading on the bit, the cam rollers will move radially out under the  action of the balls being pushed outward by the bit sleeve. The hardstop pins serve to limit the radial  deflection of the flexures for such cases so as to prevent plastic deformation.  These pins are captured in  the cam by cap that is bolted onto the cam front.  

Figure 8: Detailed CAD view of the flexures on the Mars 2020 chuck cam. The cam rollers used to  preload the balls that restrain the bit assembly are mounted onto the flexures via roller pins. The flexures  are sized so as to provide a stiff bit -to-corer connection at drilling loads while allowing the bit assembly to  contact an internal load shunt at high off nominal side loads. Flexure deflections are limited by hardstop  pins to prevent yielding under pullout loading.     

Figure 9: View of the chuck cam with integral flexures. The cam is machined as one piece using standard  manufacturing processes including milling, drilling, turning, and band saw  cutting. Despite the complex  shape of the flexures, there was no need to utilize electrical disc harge machining (EDM).
158 An analysis was performed to determine the torque needed to disengage the chuck given a preload torque  and loads applied on the bit during the unchucking operation (Figure 10). This analysis was performed for  various coefficients of friction. It was determined that even if the coefficient of friction between the roller and  shaft was 1, the chuck actuator had suffic ient torque to disengage the chuck. If one defines jamming as a  condition where two parts cannot be separated  given a certain force or torque, the analysis indicates t hat  the chuck is resistant to jamming.     

Figure 10: Plots of c alculated engagement and disengagement torques as a function of coefficient of  friction. Given the specified chuck actuator torque capabilities, the chuck is capable of being disengaged  even if the coefficient of friction at all sliding surfaces is 1.  

Mechanism Functioning  

The chuck is typically actuated to one of three positions: engaged, disengaged, and loose chucked as  shown in Figure 11. The chuck holds onto a bit in the engaged state and releases it in the disengaged state.  The loose chucked state is an intermediate position where the chuck still captures a bit but allows it  to have  a small degree of angular , axial,  and lateral play. This loose chucked state is important for bit exchange  operations where a certain amount of play is needed to accommodate positional uncertainties between the  coring drill and bit holders.  To mov e to either the engaged or disengaged states, the chuck cam is rotated  by the chuck actuator until a previously defined current limit is reached, indicating that contact with a  hardstop has been made. When moving to the disengaged state, reaching the current limit indicates that  the segmented gear on the chuck has contacted a hardstop surface. When moving to the engaged state,  reaching the current limit indicates either that the gear has contacted another hardstop surface within the
159 corer or that the cam rollers are preloaded onto the chuck balls, depending on whether a bit is inserted into  the chuck. During flight, no operations are planned to occur that would involve actuating the chuck to the  engaged state without a bit present within the mechanism.  In addition to these three states, the chuck can  be commanded to any intermediate position. For example, during bakeout, the chuck was in an “almost  engaged” position with a bit installed. This position was chosen so as to avoid preloading the bushings in  the c am rollers during the high temperatures (~114° C) of bakeout, which could result in creep in the bushing  material. The use of intermediate chuck positions has also been explored for use in recovery operations for  a bit stuck in rock.  

Figure 11: Illustration of three major chuck states : 1) engaged, 2) loose, and 3) disengaged. A bit  assembly is fully restrained and preloadd within the chuck in the engaged state, this is the position the  chuck is in during sampling operations. In the loose chucked state, the bit assembly is still restrained  within the chuck but is not preloaded and has some lateral, angular, and axial play. This play in the bit  assembly is important in allowing for robust and reliable bit exchange that can accommodate for  dimensional and positional deviations. The bit assembly is fully released when the chuck is actuated to  the disengaged state.  

Testing and Results  

Several iterations of the chuck mechanism were assembled and tested, including three flightworthy units.   Various tests to characterize chuck performance and verify capabilities were conducted at the mechanism  and corer level.  Some of more noteworthy tests conducted and their results are summarized here.  

Loose chuck bit ass embly play measurement   As previously mentioned, some degree of axial, angular, and lateral play is needed in the bit assembly  during bit exchange operations while in the loose chuck state. Measurement of this play in the assembled  flight units was accompli shed using a test bit with an affixed laser pointer. By measuring how much the light  from this laser moved on a screen when the chuck was actuated from engaged to loose, it was possible to  calculate what play was allowed. Subsequent bit exchange tests on a  qualification unit confirmed that the  play allowed for by the chuck enabled nominal functionality.
160 Bit Pullout Static Test   When sizing the components within the chuck, one of the major driving load cases was bit pullout. This  covers a situation where a bit is stuck in the rock and the full margined pullout load capability of the corer  feed mechanism is applied on the bit. When including anticipated pullout load needed to free a bit, force  margin, and potential feed actuator torque overages, this pullout l oad sums up to 6514 N. To verify the  capability to react this high load, one of the flight chuck units was subjected to a static test to verify that it  could retain a bit under high pullout loads without damage. Figure 12 includes plots of the applied forces  during this static test. After the static test, the chuck mechanism retained full functionality with no detrimental  yielding on any parts observed.  

Figure 12: Plots of force data during a static test of a Mars 2020 coring drill chuck. The applied forces  represented the worst case loading for the chuck, which is pull out on a bit with maximum applied load  from the coring dr ill feed mechanism. The lower plot is a magnified view of the force data with further  detail of the loads actually reached. After the static test, the chuck mechanism retained full functionality  with no detrimental yielding on any parts observed.
161 Qualifica tion Model Dirty Testing (QMDT)   The qualification model  of the entire Mars 2020 SCS, which includes the chuck, is currently undergoing  environmental testing as part of the Qualification Model Dirty Testing (QMDT)  campaign currently ongoing  at the time of writing. As the name implies, QMDT involves dirty testing, with the qualification model of the  SCS being used to core, abrade and collect regolith from test rocks in a representative thermal environment.  This testing is being used to verify numerous subsystem and mechanism capabilities, including ability to  function in a dirty environ ment , and is informed by previous dirty testing performed using prototype units  [4]. Ability to function in a dirty Martian environment drove several aspects of the chuck design, including  the use of rollers and flexures, the decision not to cover the outside of the cam in order to facilitate the flow  of dust and debris out of the mechanism, and exposed part fitments. QMDT thus far has demonstrated the  ability of the chuck to nominally operate in a Mars representative dirty environment, including the ability to  restrain a bit assembly to enable quality sample collection and the ability to actuate the various required  states. Actuator current draw and position when driv ing to engaged, disengaged, and loose states have  been the major metrics used to evaluate chuck performance during tests. Example plots of this data can  be seen in Figure 13. 

Figure 13: Example plots of chuck cam angular position (blue plots on the top row) and chuck actuator  current draw (red plots on bottom row). In this particular test, the chuck was actuated to intermediate  positions betwe en fully engaged ( -0.46 rad) and loose (-0.2 rad) . High current draws at the start and end  of motion are a result of the chuck actuator driving through its magnetic detent brake.
162 Challenges and Lessons Learned  

Several challenges were encountered during the design of the chuck mechanism as summarized here.  

Hertzian contact stresses   Hertzian contact stresses generated during high load cases  were a major driving factor in material selection  and component design. These high contact stresses required the selection of high strength materials , such  as Maraging 300 steel,  and platings , such as electroless nickel,  for parts such as the chuck balls,  rollers,  and bit sleeve. Contact stresses also played a role in the selection of the number and size of chuck balls.  A higher number of balls would lessen the amount of load reacted per ball but  would also increase contact  stresses due to the smaller radi us of each ball. High contact stresses also drove the material and surface  treatment selection in the bit sleeve assemblies.  The geometry of the cam rollers was also optimized so as  to minimize contact stresses, specifically by adding a concave race to the central portion of the rollers which  preload against the balls.  

Figure 14: Detail view of cam rollers and concave central races . The rollers preload ag ainst the balls at  these races. A concave race was added to each roller so as to minimize the Hertzian contact stresses  generated under worst case loads such as would occur during rover slip or bit pullout.  

Flexure design   The design of the cam flexure mounts for the rollers was a highly iterative process, requiring multiple finite  element analyses and design changes to optimize stresses and deflections in the flexures. Constraints on  flexure geometry included limitations imposed by manufacturing processes, total size envelope allocated  for the overall  chuck mechanism, number of balls needed in chuck, and stresses developed under worst  case loads.  

Contamination control accommodations   Contami nation control requirements  placed restrictions on the type of bearings and lubricants that could be  used in the mechanism. Fortunately, since no high -speed rotations occur in the chuck mechanism, the use  of ball bearings and grease lubricants was avoided. Instead, plain bearings were used throughout. To  accommodate dimensional changes due to temperature and material  differences, the larger plain bearings  had a split cut into them to minimize changes in fitment.  

Chuck actuator design and implementation   Challenges were also encountered in the design and implementation of the actuator, particularly regarding  the use of  a magnetic detent brake and the issues this caused in controlling actuator torque output.  The  great advantage of a magnetic detent brake is that it is a passive brake. Whereas an active mechanism  such as a solenoid or toothed brake would introduce an addi tional potential point of failure, a passive detent  brake does not pose such an issue. However, when applying torque, the actuator must overcome this detent  in addition to producing the needed torque output. This complicates the controlled application of torque into  the mechanism and results in potentially high overages in the torque delivered versus what is desired.
166   Figure 1. (a) Corer in turret structure (b) “Corer body” (c) Percussion mechanism in context (d) Flight  percussion mechanism und er test  

Mechanism Architecture  

Principle of  Mechanism Operation   The basic principle of operation of the hammer in the percussion mechanism is a classical base- drive,  single degree of freedom vibration system. The hammer is mounted through springs to a base that moves  with fixed amplitude, but variable frequency . As the base oscillates sinusoidally at increasing frequenc y,  the hammer and base begin to oscillate out of phase, energy is stored in the spring, and the motion of the  hammer is amplified - provided that the base drive frequency does not exceed the natural frequency of the  system . The percussion mechanism adds a fixed stop to this system (the anvil) that the hammer will contact  once a certain amplitude of motion is reached. After the initial contact between the hammer and anvil, the  system has a new, non- linear, dynamic characteristic , that can be ful ly described using vibro- impact theory   [1]. As the base is driven to higher frequencies beyond the point of initial hammer anvil contact, the hammer  will impact the anvil with greater velocity, producing larger  percussive forces  for the drill . Therefore, the  system has a practical range of frequencies at which it will percuss and those frequencies have associated  varying levels of impact energy.  

This system can be parametrized non- dimensionally by describing th e drive frequency  (𝜔𝜔) as a fraction of  the spring/mass natural frequency  (�𝑘𝑘/𝑚𝑚) and the static gap between the hammer and anvil before the  system was put into motion (Δ) as a fraction of amplitude of base motion (𝐴𝐴𝑑𝑑) – see Figure 2. Proper  selection of these non- dimensional parameters will ensure that operation will occur in a regime where a  stable vibro- impact process will be maintained and that the system will always operate with an impact  frequency equal to the base oscillation frequency , as opposed to regimes where impacts occur as integer  multiples of the base oscillation frequency.
167   Figure 2. (a) Block diagram for the percussion hammer mechanism, a base driven spring mass with a  stop (b) The physical components represented in the bloc k diagram (c) A plot of hammer and driver  motion from a simulation. Note that the hammer amplitude grows as driver frequency increases until first  hammer/anvil contact at 25 Hx (indicated by red asterisk on the plot). Subsequent impacts at higher  frequenci es will have higher velocity and will generate more force.  

In the case of the Mars 2020 percussion mechanism, the hammer mass is 200 g, the total spring stiffness  is 11 N/mm, the single- sided amplitude of motion at the base (i.e., crank shaft throw)  is 5 mm and the static  gap to the anvil is approximately 9 mm. The result is a system that first impacts with base motion (and  impact frequency) of approximately 25 Hz and the corer typically operates the percussion mechanism  between 25 Hz and 40 Hz . Impact velocities are typically between 0.5 m/s and 3 m/s but will depend on the  effective coefficient of restitution of the system and other factors.
170 that both springs are always in compression throughout the full range of driver and hammer motion.  Assembly of the springs and driver plate onto the hammer is possible because the hammer shaft is a  separate part that is threaded into the hammer . The hammer rides on split -bushings made from Vespel  SP-3 (polyimide doped with MoS 2) that slide without additional lubricant against a guide shaft made of  15-5PH H1025 steel with a fine ground finish.  

Figure 5. (a) Hammer assembly (b) Single bearing, bevel pinion, crank shaft, and roller ball  

The Percussive Chain as a System –  Hammer / Anvil / Bit  The effectiveness of the percussion mechanism at the drill level depends heavily on the design of the  downstream components in the percussive chain, specifically the anvil and drill bit . To transfer the impact  loads efficiently, the parts must be axially stiff , ideally with a solid column of metal running directly form  under the hammer head down to the rock . The anvil and bit have numerous other functions besides  transferring impact loads , so their design must be a compromise. In general, designs that transfer impact  loads well will avoid abrupt changes in the cross section s normal to the drilling axis  (i.e., gentle tapers, not  deep sharp- cornered grooves) and will avoid highly compliant features that extend out radially, as these  will parasitically sap percussive energy when set ringing by an impact  (i.e., no thin, large radius flanges) .  Although the anvil, bit, and ground should all be in firm contact before an impact from the hammer is  delivered, so as not to waste energy sending parts into rigid body motion, testing on the Mars 2020 corer  has shown that approximately 80 N to 100 N was sufficient weight -on-bit preload to achieve efficient force  transfer and that additional preload did not have a benefit for that system. Contact surfaces pairs (face of  hammer, back to anvil for example) were made curved, to minimize sensitivity to misalignment as parts  came in and out of contact, and also conformal  (sphere in cup) , in the hopes of creating a more uniform  stress wave through the part cross section.
175   Figure 5. Typical results from a force dyno test performed at the mechanism level (force measured under  the hammer). (a) Peak impact force statistics gathered from 3 seconds of operation at each frequency  from 28 Hz to 40 Hz. (b) By us ing the time stamp from each peak force point, the frequency of impacts   can be calculated independent of actuator telemetry. Results show that impact occur at a very consistent  rate given that the hammer is driven through a spring and that each individual impact has a number of  uncontrolled parameters in terms of the motion and orientation of the contacting parts.  

Challenge: Achieving Durability in the Crank Shaft and R oller  The interface between the crank shaft pin and roller ball is a grease filled plain bearing that will experience  as many cycles as there are hammer impacts . The estimation of the loads experienced by the interface is  complicated by the fact that the load is shared by three crank shaft s with an unknown load distribution and  that i t bears a cyclic load due to the oscillating inertia of the hammer assembly combined with hammer  impact loads, although the impact  loads should be smoothed by the springs . In retrospect, more attention  should have been paid early in the design better understand the loads at this interface and thus properly
177   Figure 6. (a, b) Heavily worn crank shaft pins from early prototype. Wear was so severe in some cases  that the pin became too thin and broke off. (c) Comparison of the design after improvements were  introduced to reduce wear between the roller ball and crank shaft pin  

After a comprehensive investigation that eliminated all other sources of drag, the bearings were found to  be the cause of the large increase in drag at - 70°C . The solution was to drastically reduce the grease fill to  approximately 3%, and th us ultimately reduce the drag to acceptable levels . Improved warm up routines  were also devised that eliminated starts and stops until the mechanism had gradually accelerated and then  run for about 1 minute. The mechanism life tests mentioned above were all conducted using this reduced  grease fill and at the end of the test there was no sign of bearing performance degradation. The lesson  here is well worn, but worth repeating: early life testing in a relevant environment is invaluable and the  increase in the viscosity of Braycote 600EF grease is extremely non- linear in the - 50°C t o -70°C range. A  large number of bearings and relatively low -torque actuator (relative to power) in this mechanism  exacerbated the sensitivity to the thickening of the grease.  

Conclusions  

A percussion mechanism that will enable the Mars 2020 coring drill  to core and abrade hard rocks in the  Martian environment was successfully designed, built, and tested and is scheduled to arrive on the surface  of Mars in February 2021 and begin operation soon thereafter . The novel mechanism, with a conventional  rotary actuator drive, a base -driven hammer, and triple Scotch yoke architecture made it possible to perform  variable output force percussing whilst subjecting only a simple and robust subset of mechanism hardware  to the large, dynamic impact loads.  

Acknowledgement  

The research was carried out at the Jet Propulsion Laboratory, California Institute of Technology, under a  contract with the National Aeronautics and Space Administration (80NM0018D0004).
178   Figure 7. Hammer bushing and retainer. (a) Origi nal GGB DP4 bushing wore quickly and (b) Bushing  retainer failed catastrophically soon after 1X life. (c) LS -Dyna analysis revealed the severity of compression  / tension reversing loads at the top end of the hammer shaft. (d) The best retainer was no retai ner but  instead to keep the bushing in a blind groove, which was enabled by (e) A Vespel SP -3 split bushing that  could be installed into the blind groove and exhibited better wear performance.  

References  

[1]  V. I. Babitsky, Theory of Vibro -Impact Systems and Applications, Berlin: Springer -Verlag, 1998.   [2]  A. Siddens, "Validation and Predictions of Explicit Dynamics Simulations for the M2020 Percussion  Drill," in Spacecraft Launch Vehicle Dynamic Environments Workshop , El Segundo, 2018.   [3]  Y. Bar -Cohen and K. Zacny, Drilling in Extreme Evironments, Darms tadt: Wiley, 2008.   [4]  T. G. M. S. S. e. a. Spohn, "The Heat Flow and Physical Properties Package (HP3) for the InSight  Mission," Space Science Review, no. 214, 2018.