ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Figure 5. Kepler Primary Mirror Mounted onto the Focus Mechanisms.  Focus Mechanism Driving Requirements   Table 1 summarizes the requirements for the Focus Mechanism and shows the performance based on flight unit testing. The focus mech anism has the ability to tip and tilt the primary mirror, but this is not a  mission requirement. Only axial focus displacement of the PMA is required. 

Table 1. Focus Mechanism Key Driving Requirements. 

Requirement Value Performance Axial position knowledge ±1.4 μm 0.5 μm  Operating Temperature Range -55C to +35C Meets  Survival Temperature Range -65C to +45C Meets  Focus mech Motor Case Operational Temperature Range -55C to +60C Meets  Focus Mech Motor Case Survival Temperature  Range -65C to +70C Meets  Range of Travel ±762 μm ±900 μm  Smallest increment of Travel /g1481.5 μm 0.4 μm  Unidirectional Repeatability /g1481.25 μm Meets  Focus Mechanisms shall not require power during launch n/a Meets  4 Year Lifetime equates to mechanism cycles 112 cycles Meets  Mass (1 Focus Mechanism) /g1484.9 Kg 4.13 Kg  1st Mode Frequency, Axial >50 Hz 61.2 Hz  1st Mode Frequency, Lateral >50 Hz 70.0 Hz  Maximum Launch Acceleration 41 g Meets     Focus Mechanism Design Overview   A unique design was incorporated in order to meet the difficult resolution requirement of less than   1.5 microns over a range of at least 1500 microns. This design includes a variety of the classical moving mechanical components that were integrated into an elegant, robust and efficient system. Some of the  key components of the Focus Mechanism include a stepper motor, gears, ball screw, bearings, lever arm  and several flexures. Figures 6 and 7 show the front and aft views of one of the three flight Focus  362
Mechanisms. Redundant LVDT (Linear Variable Differential Transformer) sensors are mounted on both  sides of the Primary Mirror Interface shelf and used for position sensing. The Stepper Motor, which drives  the entire system, also has primary and secondary windings for redundancy. 

Figure 6. Kepler Focus Mechanism, Front View.  Figure 7. Kepler Focus Mechanism, Aft View. 

Function   The Focus Mechanism is driven by a 30° stepper motor with an integral 100:1 gear head. The output  shaft of the gearmotor attaches to a 2:1 pinion-spur gear reduction. The spur gear is attached to the shaft of a 2-mm pitch ball screw which is supported by duplex bearings at the base and a radial bearing on top.  A titanium ‘wishbone’ bracket with flexures attaches the ball nut to the end of a titanium lever arm. As the  ball nut moves up or down, the lever arm pivots on an integral pivot flexure while a 2 nd drive flexure, also  integral to the lever arm, moves the primary mirror interface shelf up or down. A parallel double bladed  flexure provides nearly perfect up and down movement of the interface shelf. The 8.75:1 lever arm  provides mechanical advantage by reducing load requirements and also allows finer displacements.  Redundant LVDT sensors, mounted on both sides of the primary mirror interface, measure  363
displacements. Upper and lower hard stops are attached to the ball screw to prevent damage to the  mechanism. Figure 8 shows a section view of the mechanism solid model. 

Figure 8. Section View taken from Focus Mechanism solid model. 

Focus Mechanism Design Background   The initial conceptual development of the Focus Mechanism began at Ball Aerospace in September, 2002  by Robert Warden. The design con cept (Figure 9) was based on  the components of the Spitzer Space  Telescope (formerly SIRTF) Secondary Mirror Cryogenic Focus Mechanism.  

Figure 9. Initial Concept for Focus Mechanism in 2002.   Numerous developments and improvements were made to the mechanism from late 2003 to 2006. Some  of these developments include: 1.  The parallel double bladed flexure was made more compliant to decrease the torque requirements and  stresses on the drive and pivot flexures.    2.  The main flexure interface ‘shelf’ was widened to accommodate the interface to the primary mirror strut brackets.  3.  The geometry of the lever arm was optimized to minimize bending and maximize mechanical  advantage. 4.  The geometry of the lever arm pivot and drive flexures was optimized to minimize stresses.  5.  The ball nut flexure ‘wishbone’ design was optimized to eliminate ‘wind-up’ and minimize stresses.    6.  Redundant LVDT sensors were ad ded for 16 bit position sensing.   364
Bearings   The bearings used to support the ball screw are a duplex pair at the base and a radial ball bearing on top.  The duplex pair uses face-to-face mounting to allow for slight misalignment and also to eliminate axial  and radial play in the system. The preload in the duplex pair is 13.3-26.7 N (3-6 lb). As shown in Figure 12, the ball screw shaft is supported and rotates with the inner races of the duplex pair. The sleeve and  retainer disk also rotate with the ball screw. The ball screw shaft is allowed to float axially on the upper  radial bearing which prevents possible binding due to CTE mismatch. There is a 0.2-mm radial gap in the mounting design for both bearings to provide a labyrinth seal which keeps dust out and prevents the lube  from migrating. 

Figure 12. Section view of Bearings/Ball Screw    Lever Arm

The lever arm provides a mechanical advantage of 8.75:1 which helps achieve the required resolution and torque margin in the system. The  shape of the lever arm resulted in analysis to optimize the design  for minimum bending at the extents of travel. The bending forces on the lever arm increase as the  mechanism approaches the hard stop positions due to bending of the flexures. This results in a system that is not perfectly linear. An identical amount of motor steps will result in approximately 5% less  movement of the primary mirror at the extents of travel versus at the null position. This non-linearity is  mitigated by the LVDT measurements. The lever arm was fabricated from titanium, like most of the  367
The stepper motor contains 2 sets of 2 phase windings, each capable of independently meeting the  performance requirements. The motor heater foil contains two sets of coils, only one is required for  normal operation. 

Figure 14. Focus Mechanism LVDT Sensor and Motor Heater. 

Analysis   Detailed finite element models of the focus mechanism were produced to predict structural and thermal  stresses along with modes during launch loads. As a result of this analysis, several features of the focus  mechanism were adjusted as the design matured to produce positive safety margins. A surrogate primary mirror was mounted onto the three completed focus mechanisms and vibration tested. The finite element  models were correlated to the vibration test data to more accurately predict the axial, lateral and rocking  modes of the focus mechanism/primary mirror assembly during launch loads.    Testing

Static Load Test

On April 18, 2006 a static load test was performed to measure the stiffness of the focus mechanism in all  3 axes. An MTS Sintech Extensometer was used to apply a force on the focus mechanism in the area of the primary mirror interface and measure the resulting displacement. The forces applied were within the  elastic limit of the unit and the data was used to correlate the finite element model of the focus  mechanism. Figure 15 shows the Y-axis test. 

Required Torque Test

In March, 2007 a required torque test at cold temperatu re was performed on the focus mechanism to  determine the torque margin at the worst case temperature. The test was run in a climatic chamber that  was cooled with liquid nitrogen. The motor was removed from the focus mechanism and replaced with a bearing/pinion gear/shaft assembly that was heated with a foil heater wrapped around the perimeter of  the cylindrical housing. The torque to actuate the mechanism was measured with a torque watch. The  results were within 20% of predicted values a nd the torque margin was acceptable.  369
Figure 15. Focus Mechanism Static Load Test, Y-axis. 

Functional Test   All three of the focus mechanisms underwent and passed extensive functional testing in the spring of  2007. Most of these tests were performed to show that the focus mechanism meets the requirements shown in Table 1. Some of these tests included range of motion, unidirectional repeatability, axial position  knowledge and smalle st increment of travel.    One issue that somewhat slowed testing was the requirement for the motor ca se maximum operational  temperature. This requirement states that the maximum temperature that the motor case can operate at is  +60C. The reason for this is to avoid warming the lubricant in  the bearings. If the  lube becomes too warm,  the viscosity decreases and it can migrate out of the bearings. During testing the motor could only run for  approximately one minute at a time before the case temperature became too high. Temperature sensors  on the motor case were monitored closely to insure that this limit was not exceeded.  

Figure 16 shows actual data from Focus Mechanism functional testing. The graph shows absolute  position vs. LVDT counts for the entire range of motion. The X-axis is the absolute position, measured with a Ziess CMM (Coordinate Measuring Machine) and the Y-axis shows the corresponding primary and  redundant LVDT counts. Each mechanism was fully characterized during this testing.   Vibration Testing

In April, 2007, all 3 focus mechanisms passed vibration testing with a surrogate primary mirror mounted to  them. Figure 17 shows the test setup for one of the axes.  

370
Figure 16. Range of motion test. 

Figure 16. Vibration Test 

Thermal Vacuum Testing   In May, 2007, all three focus mechanisms passed thermal vacuum testing. All three units were tested  simultaneously as shown in Figure 17. Mass simulators, 1/3 the mass of the primary mirror, were  mounted to the interface shelf to simulate the correct loads. Each focus mechanism was mounted to an Invar base plate to simulate the CTE of the composite aft bulkhead. Cooling boxes were mounted over  the motor cover box to increase the cooling rate at the motors and speed up functional testing  371
Figure 17. Thermal Vacuum Test 

Conclusion  The design, fabrication, assembly and testing of the focus mechanism was successful due to the  combined efforts of dozens of engineers, machinists and technicians at Ball Aerospace. It is impossible to list all of the lessons learned during this effort. One issue that caused a slight schedule delay involved the  ball screw. When the ball screws were disassembled and lubricated at Ball Aerospace it was discovered  that the ball sizes were oversized and the pre-load was too high. New balls needed to be ordered from the manufacturer which delayed schedu le. It is recommended up front to order a spare set of balls, 2.5 &  5.0 /g541m (0.0001 & 0.0002”) larger and smaller than the standard ball size to avoid delays.   The focus mechanism has successfully met all design and schedule requirements that it has faced and is  currently operating flawlessly in the Kepler Telescope. 

Acknowledgements  The author would like to thank all of the Kepler machinists, technicians, analysts, engineers and  managers that played a role in developing the focus mechanism. The skill, hard work, passion and  dedication of this team of people were vital to the success of this effort.   References 

1. Kepler mission overview. Kepler/NASA website. http://kepler.nasa.gov/

2. Illustration of Milky Way Galaxy copyright Jon Lomberg.  http://www.jonlomberg.com  

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Precision Linear Actuator for Space Interferometry Mission (SIM)   Siderostat Pointing  Brant Cook*, David Braun*, Steve Hankins*, John Koenig* and Don Moore*  Abstract “SIM PlanetQuest will exploit the classical measuring tool of astrometry (interferometry) with  unprecedented precision to make dramatic advances in many areas of astronomy and astrophysics” (1). In order to obtain interferometric data two large steerable mirrors, or Siderostats, are used to direct starlight into the interferometer. A gimbaled mechanism actuated by linear actuators is chosen to meet the unprecedented pointing and angle tracking requirements of SIM. A group of JPL engineers designed, built, and tested a linear ballscrew actuator capable of performing submicron incremental steps for 10 years of continuous operation. Precise, zero backlash, closed loop pointing control requirements, lead the team to implement a ballscrew actuator with a direct drive DC motor and a precision piezo brake. Motor control commutation using feedback from a precision linear encoder on the ballscrew output produced an unexpected incremental step size of 20 nm over a range of 120 mm, yielding a dynamic range of 6,000,000:1. The results prove linear nanometer positioning requires no gears, levers, or hydraulic converters. Along the way many lessons have been learned and will subsequently be shared.  Introduction SIM will improve “our understanding of the physical properties of stars, determining the mass, including  the dark matter component, and its distribution in our Galaxy, observing the motions of the Milky Way’s companions in the Local Group, and probing the behavior of supermassive black holes in other galaxies” (1). Using three interferometers, 1 science and 2 guides, SIM will deliver a dramatically more accurate mapping of our universe as well as a better understanding of the formation and evolution of other planetary systems outside our own. Accurate mapping using interferometers requires high precision actuation, pushing the limit of both the mechanical positioning realm and the electronics/control realms. The required lifetime of SIM is also extremely challenging with a need to meet performance requirements for no less than 5.5 (with a goal of 10) years of continuous science observation. This is approximately 3 million large angle gimbal moves. Figure 1:  SIM PlanetQuest  While many precision mechanisms are required for SIM to deliver its science data, the Siderostat is the  initial pointing mechanism used to direct starlight into the instrument. The Siderostat tips and tilts a 304.5-mm clear aperture optic across a 15-degree Field of Regard (on the sky) via a two-axis hexfoil flexured gimbal mechanism, with a required coarse accuracy of 1arc-second (as) and a fine accuracy of                                                              * Jet Propulsion Laboratory, California Institute of Technology, Pasadena, CA  Proceedings of the 39th Aerospace Mechanisms Symposium, NASA Marshall Space Flight Center, May 7-9, 2008  373
5 milli-arc-second (mas) on the sky. Implementing a pair of linear actuators approximately 300 mm from  the gimbal axis yields a required linear coarse accuracy of approximately 1 micron and a fine linear accuracy of approximately 5 nanometers. See Figure 2, Potential Siderostat Configurations.  Figure 2:  Potential Siderostat Configurations  This paper will discuss the design, build, and test of a linear actuator capable of repeatedly, over long  periods of operation, performing submicron positioning maneuvers. The design successfully pushes the limits of mechanical positioning while remaining true to the JPL principles of heritage, simplicity, and robustness. Along the way interesting lessons were learned and will be put forth for the reader’s benefit. Mechanical Design  Generally speaking, design is a continuous balancing act to meet competing requirements. It requires the  designer/engineer to carefully balance the design process between many conflicting requirements. In actuator design, if mutually exclusive positioning requirements are equally important a two-stage  mechanism is often the solution. Unfortunately, multiple stages add to complexity, mass and cost. In our design, the need for large fast moves, trump the needs for fine positioning, and vice versa. It was originally thought that no single actuator could meet the large stroke, high speed, small incremental step size (actuator resolution), required by the Siderostat. The original intention of our Precision Linear Actuator, Direct Drive (PLADD) was to fill the role of a coarse actuator.  The design approach taken with PLADD is one of simplicity, heritage, and robustness. The initial design  required at least 2000 incremental positions per revolution of the nut/motor in order to obtain a linear incremental step size of 1 micron. Figure 3 shows a cross section of the actuator.  Figure 3: SID PLADD Cross Section View  374
The bearing housing, as shown in Figure 4, is the center of the actuator, housing the back-to-back duplex  angular contact bearings, brake, and motor. Both sets of bellows are suspended from the bearing housing, while the housing bolts directly to the carbon fiber outer tube for thermal stability.    Figure 4: Bearing Housing  The angular contact bearings are assembled into the bearing housing from opposing ends of the housing.  The housing is stepped, creating a bearing spacer against which the bearings can be preloaded. The bearings are thin section MPB (Miniature Precision Bearings) angular contact ball bearings with a phenolic retainer ball spacer. The bearings have 440C balls and races. The bearings are shown in Figure 5. Figure 5: Timken MPB Angular Contact Ball Bearing  The bearing inner races are mated to the motor drive link. As depicted in Figure 6, the inner race of the  inner bearing seats on a non-sliding preload flexure. The DC brushless motor rotor is assembled to the rear of the drive link via set screws. The ballnut is mounted to the front of the drivelink via bolts. Figure 6: Bearing Preload Flexure and Rotor mounted to Drive Link  The motor is coupled to the NSK ballnut by way of an in-house designed and built drive link. The drive  link is shown in Figure 7. The complexity of the mechanism is reduced and backlash of a gear train is eliminated via a direct drive  approach. By eliminating mechanical backlash in the actuator the control system is able to actively eliminate any mechanical imperfections in the actuator using feedback from the linear glass scale  375
encoder mounted to the actuators’ output. The direct drive approach also greatly simplifies manufacturing  and assembly of the actuator by lowering part count and complexity.  Figure 7: Motor Drive Link  The heart of the actuator is a 12-mm diameter, 2-mm lead, NSK ballscrew mounted onto an NSK Double  nut utilizing a spring preload of 130 newtons. The combination can be seen in Figure 8.  Figure 8: NSK Ballnut  The ballnut is connected directly to the drive link via 8 bolts.  Figure 9: Ballnut bolted to Drive Link  The ballnut drive link combination is assembled into the bearing housing with the angular contact  bearings, preload flexure, and finally the piezo brake as pictured in Figure 10.  376
Figure 10: Housing Assembly  The piezo brake consists of a set of flexured levers, preload springs, and guide flexures machined out of  a single piece of titanium. The drive-link is radially compressed by the machined springs upon power-off. This constrains the drive link, and thus the ballnut, in rotation. The power must be applied in order to release the brake. The braking force is applied by the compressed preload springs via a wedged preload shim. The preload force is eliminated by the application of power to the piezo stacks.  Figure 11: Piezo Brake  The actuators’ driving torque is delivered via an off the shelf DC brushless motor from BEI/Kimco  Magnetics as pictured below. Figure 12: BEI DC Brushless Motor  Long life requirements along with a high correlation between lubricant consumption and life limitation led  the team to immerse the mechanical drivetrain of PLADD in oil via a set of hermetically sealed bellows mounted to each end of the ballscrew. The bellows create a constant volume to house Brayco 815Z oil, and subjects the mechanical system to a continuous oil flush. The bellows additionally are used to hold the screw in rotation, thus allowing the rotating ballnut to produce linear motion at the screw. The continuous motion of oil through the mechanical system is conjectured to eliminate many failure modes. This hypothesis is being tested via life tests that are in progress. The motor windings, piezo brake, and inner bellows are bolted to the outer housing via thru bolts.  377
Figure 13: Inner Titanium Bellows  Figure 14: Motor Windings Mounted  A MicroE linear glass scale is connected to the end screw via a set of parallel motion flexures. These  flexures provide a thermally stable mount for the glass scale via a low CTE metering rod. The scale is allowed to float relative to the inner tip of the ballscrew, minimizing the coupling of thermal drifts in the parts of the actuator not in the position path. The use of metering rods helps minimize false delta readings at the sensor head. The glass scale is mounted on a carbon fiber rectangular rod. A dummy glass scale is mounted opposite the operating scale to help eliminate bending in the scale due to changes in bulk temperature.  Figure 15: Glass Scale Encoder Mount via Flexures  The sensor head is mounted via a set of parallel motion flexures and the composite tube that acts as a  metering rod. This completes the thermally stable sensor mount.  378
Figure 16: Sensor Mount  All component alignment is achieved via sets of self aligning slip fit radial diameter lips between mating  parts. All components are thus axially self aligning to within the tolerances of the slip fits. O-rings are used at all interfaces to eliminate oil leaks.  The actuator connects to ground via a coaxial orthogonal flexure blade flange. The metering rods are  hard mounted to each end joint.  Figure 17: Flexured End Joint  Results The initial resolution requirements for the ballscrew actuator was ±1 micron which was not clearly  achievable based on vendor data for ballscrew position as a function of rotation angle. The selection of the Micro-E linear encoder capable of 5-nm position feedback was expected to give position sensing feedback well beyond that which was needed to determine the limits of ballscrew positionability. The direct drive architecture and a preloaded ballnut minimized actuator windup and backlash that enabled maximal position control. Additionally, rather than control the motor commutation based solely on the hall effect sensor as with typical motor control schemes the team also utilized the output sensor data. Since a 3-phase DC brushless motor is being driven, a PID controller reads the linear encoder and outputs 3-phased sinusoidal voltage commands to the 3 motor windings. Typical trapezoidal commutation uses only the Hall Effect sensors to apply the voltages in a rough way  (24 steps per motor rotation). These 3 discrete sensors decode to 6 distinct points per pole, and since there are 4 poles per motor revolution that provides 24 points per revolution. Sinusoidal commutation makes use of the encoder data to fill in the gaps between the Hall Effect transitions and increase the points per revolution from 24 to 409,600. This allows the sinusoidal shape of the back-EMF waveform to be matched much more precisely.  The geometry of the motor is such that sinusoidal commutation converts all of the applied current into  useful torque (in the ideal case), rather than into undesirable radial forces. Since all the current is being converted into torque, the variation in torque is minimized and thus torque ripple is reduced compared to  379