ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
27   Figure 16.  Overall release velocity in flight  

Conclusions  

The injection of an object into a geodesic trajectory constitutes a relevant space engineering challenge. In this paper we present a ground- based testing approach, which allowed us to provide the qualification of  the mechanism designed and developed to perform the injection of a 2- kg gold- coated test mass into a  geodesic trajectory of unprecedented purity (LISA Pathfinder). The criticality of the mechanism for its test mass injection into geodesic function mainly relies on the possible asymmetry of the two opposed actions  on the test mass. Both the holding forces and following adhesive interaction between the test mass and its holding tips may produce a net impulse at the release, i.e. , an excessive test mass residual velocity,  when the dynamic behavior of the two opposed contact on the test mass results asymmetric.    The testing method and the test results for the Grabbing Positioning and Release Mechanism developed  for the LISA Pathf inder mission are here presented and discussed.

References  

1. F. Antonucci, et al. “The LISA pathfinder mission”. Classical Quantum Gravity, vol. 29, no. 12, pp. 124014- 1–124014- 11, 2012.   2. M. Armano et al. "The LISA Pathfinder Mission". Journal of Physics: Conference Series Volume 610,  Issue 1, 11 May 2015, Article number 012005.   3. B. Zahnd, M. M. Zimmermann, and R. Spö rri. “LISA- Pathfinder cage and vent mechanism  —  Development and qualificat ion”. Proceedings of the 15th European Space Mechanisms and  Tribology  Symposium , 2013, pp. 1– 7.  4. A. Neukom, R. Romano, and P. M. Nellen . “Testing and lesson learnt of LISA GPRM” . Proceedings  of the 13th European Space Mechanisms and Tribology Symposium , 2009, pp. 1 –8.  5. I. Köker, H. Rozemeijer, F. Stary, and K. Reichenberger . “Alignment  and testing of the GPRM as part  of the LTP caging mechanism ”. Proceedings of the 15th European Space Mechanisms and Tribology  Symposium , 2013, pp. 1– 7.  6. M. Benedetti, D. Bortoluzzi, M. D. Lio, and V. Fontanari. “The influence of adhesion and sub- Newton  pull-off forces on the release of objects in outer space”. J. Tribology, vol. 128, no. 4, pp. 828– 840,  2006.   7. D. Bortoluzzi, M. De Cec co, S. Vitale, and M. Benedetti.  “Dynamic measurements of impulses  generated by the separation of adhered bodies under near -zero gravity conditions ”. Exp. Mech., vol.  48, pp. 777– 787, 2008.   8. D. Bortoluzzi, M. Benedetti, L. Baglivo, and S. Vitale.  “A new perspective in adhesion science and  technology: Testing dynamic failure of adhesive junctions for space applications ”. Exp. Mech., vol.  50, pp. 1213 –1223, 2010.
28 9. D. Bortoluzzi, M. Benedetti, and J. W. Conklin.  “Indirect measurement of metallic adhesion force as a  function of elongation under dynamic conditions ”. Mech. Syst. Signal Process,  vol. 38, no. 2, pp.  384– 398, 2013.    10. D. Bortoluzzi, L. Baglivo, M. Benedetti, F. Biral, P. Bosetti, A. Cavalleri, M. D. Lio, M. D. Cecco, R.  Dolesi, M. Lapolla, W. Weber, and S. Vitale.  “Lisa pathfinder test mass injection in geodesic motion:  status of the on ground testing”. Classical Quantum Gravity, vol. 26, no. 9, pp. 094011- 1–094011- 11,  2009.   11. C. Zanoni,  D. Bortoluzzi, J. Conklin, I. Köker , C.Marirrodriga, P. Nellen, and S. Vitale . “Testing the  injection of the LISA -pathfinder test mass into geodesic conditions ”. Proceedings of the 1 5th  European Space Mechanisms and Tribology Symposium , 2013, pp. 1– 10.   12. D. Bortoluzzi, M. Benedetti, L. Baglivo, M. De Cecco, S. Vitale. “Measurement of momentum transfer  due to adhesive forces: On -ground testing of in- space body injection into geodesic motion ”. Rev. Sci.  Instrum. 82, 125107 (2011).   13. D. Bortoluzzi, P. A. M äusli, R. Antonello, and P. M. Nellen. “Modeling and identification of an electromechanical system: The LISA grabbing positioning and release mechanism case” . Adv. Space Res.,  vol. 47, no.  3, pp. 453– 465, 2011.   14. M. Benedetti, D. Bortoluzzi, and S. Vitale.  “A momentum transfer measurement technique between  contacting free- falling bodies in the presence of adhesion” . J. Appl. Mech., vol. 75, no. 1, p. 011016,  2008.   15. M. D. Cecco, D. Bortoluzzi, L. Bagli vo, M. Benedetti, and M. D. Lio.  “Measurement of the momentum  transferred between contacting bodies  during the L isa test -mass release phase uncertainty   estimation” . Meas. Sci. Technol., vol. 20, no. 5, pp. 055101- 1–055101- 15, 2009.   16. K. Rabenorosoa,  C. Clévy , P. Lutz , M. Gauthier , P. Rougeot, Micro & Nano Letters, 2009, Vol. 4, Iss.  3, pp. 148– 154.  17. A. R. Savkoor and G. A. D. Briggs  , Proceedings of the Royal Society of London. Series A,  Mathematical and Physical Sciences, Vol. 356, No. 1684 (Aug. 15, 1977), pp. 103 -114.  18. D. Bortoluzzi, J. W. Conklin, and C. Zanoni . “Prediction of the LISA Pathfinder  release mechanism  in-flight performance ”. Adv. Space Res., vol. 51, no. 7, pp. 1145– 1156, 2013.   19. C. Zanoni and D. Bortoluzz i. "Experimental -Analytical Qualification of a Piezoelectric Mechanism for  a Critical Space Application". IEEE/ASME Transactions on Mechatronics, Vol. 20, No. 1, February 2015.
29 Mechanisms Design for the S5P- TROPOMI Instrument  

Erik Tabak *, Jan de Vreugd **, Dorus de Lange** and Frits van der Knaap** 

Abstract  

TNO developed two mechanisms as a module of the Sentinel 5 Precursor (S5P)  TROPOspheric  Monitoring Instrument ( TROPOMI ). The TROPOMI  is an advanced absorption spectrometer for E arth  observation, developed in T he Netherlands under contract to NSO and ESA for  the ESA Copernicus  Space Component Programme. The TROPOMI will allow continuation of  the 15- year satellite data sets  started with GOME, OMI and SCIAMACHY. This paper will address all design  aspects of the mechanism  starting with design considerations and ending with the results of a successful qualification campaign .  The design challenges of these mechanisms are: a  tight design space, the overall mass, the bearing  loads due to both a relatively high carousel mass and vibration launch load (~30g quasi static  loading  which drives the bearing loads ), the relatively high  required angular reproducibility,  and the thermal  constraints (operational temperature range).  

Introduction  

TNO has extensive heritage in designing mechanisms for optical space instruments. In this particular  case it concerns the two mechanisms within t he instrument needed to fulfill different functionality. One of  the mechanisms, the Folding Mirror Mechanism (FMM), requires high positional accuracy and reproducibility on one angular position for the mirror , whereas the second mechanism, the Diffuser  Mechanism (DIFM), has to provide six distinctive angular positions for the diffuser carousel with sufficient  accuracy and reproducibility. The goal has been to use the same bearing and motor type for the two mechanisms. For reliability reasons a redundant winding motor concept is chosen as baseline for both mechanism s. 

Functional description of the Calibration Unit (CU)  

The two mechanisms of the Calibration Unit ( CU) cooperate to create several optical paths enabling  calibration and nadir view . Figure 1 show s the different rotational positions of both mechanisms in section  view. The F olding Mirror Mechanism ( the smaller mechanism at the right top in the figure) only has one  relevant position. It should be able to rotate to enable the slit in the correct posit ion.   The Diffuser Mechanism ( DifM) has to be able to rotate the optics in six  different positions:   • Solar calibration: this is done in two positions, mutually rotated with 180°   • White Light Source Calibration (90°  rotated w ith respect to the solar calibration)   • Laser diode calibration (180°  rotated with respect to the White Light Source calibration position)   • The LED calibration position is rotated with respect to the solar calibration with 45° ) 

The FMM has two positions:   • Reflecting light of any of the six  positions of the DifM   • Passing through Earth radiance  

                                                  *   InnoSpace, Rotterdam, The Netherlands   ** TNO, Delft, The Netherlands  

Proceedings of the 43rd Aerospace Mechanisms Symposium, NASA Ames Research Center, May 4-6, 2016
31 An overview of the driving requirements for both mechanisms is shown in Table 1.   Table 1.  Overview of the driving requirements    Requirem ent Value  Remarks   Reproducibility DifM    RY ± 0.42 ° The diffuser carousel shall have a reproducibility  of 0.42 degree (zero to peak = 2σ ) 

Reproducibility FMM    RY ± 0.1° The folding mirror shall have a reproducibility of  0.1 degree (zero to peak = 2σ ).  Testing  - The folding mirror and the diffuser mechanisms  shall be designed to allow complete testing in restricted cleanliness and environmental conditions on ground individually and at spacecraft level.   Number of full  rotations in space  30000  The folding mirror and the diffuser mechanisms shall be able to operate for 30000 cycles.   Design survival  temperature  -50°C to +45°C  incl. qualification margin  Design operational  temperature range  +7°C to +33°C  incl. qualification margin   Rotating Mass  (carousel)  3.2 kg incl. 20% design contingency. Applies to the  DifM (worst case)   Eigenfrequency  > 500 Hz  First eigenfrequency of the mechanisms. It  concerns here a design goal   Positions  DifM: 6 positions   (4x90°, 2x45°)   FMM: 2 positions    Oscillation stroke and  frequency  ± 5 deg @ 0.5Hz  During a monthly calibration, 3 periods of 10  minutes of oscillations are required   Rotational Speed  90° in 10 s    Position accuracy  RY ± 0.5  deg A worst case estimation   Parasitic translations  over stroke  ΔX, ΔY, ΔZ< 0.1 mm  A worst case estimation   Parasitic rotations  over stroke  ΔRX, ΔRZ< 0.1 deg  A worst case estimation   Number of full  rotations for on ground  testing  3000    QS Design Load  30 g
32 Bearing  Design Considerations  

Due t o the need for relatively large rotations in the mechanisms , the application of bearings is inevitable.  This automatically introduces (dominant) tribology in the system, which together with the application in  vacuum, is considered as a design driver. For that reason, the design of the mechanism(s) starts with the  design of the bearings.  

Initially, a desig n with two angular contact bearings  (in combination with a soft preload) would suffice.  This provides a relative low complex ity design. One end remains fixed whereas the other end is flexible or  able to slide in the axial direction (this is in order to minimize the disturbance forces due to thermomechanical loads).  However, in order to guarantee surface contact in the bearings (during launch loads),  the preload needs to be higher than the forces induced by launch.  This yields relatively high required  preloads . For the mechanisms as described in this article this would yield negative motorization margins.  Additionally, the bearing gapping would become unacceptably high. Both the topics motorization margin  and the bearing gapping are treated later on in this article.    A solution to the problem  of the high preload and large amount of gapping  is by applying duplex bearings .  A duplex bearing has  the advantage that it can be loaded to high loads in both directions while  maintaining a relative small preload. A duplex bearing is a bearing set which is prepared to have a  dedicated preload, once clamped together. The principle (as shown in Figure 2) is that the width of all the four rings (the inner and outer bearing rings of both bearing) is measured. Next, it is determined by the  bearing supplier how much is required to polish from the outer ring of one of the two bearings. Typically this is in the order of mic rons with a preload of tens of newtons. Next, the bearings are assembled and  compressed together in such a way that the outer rings are in contact with each other. The step- by-step  procedure is illustrated in Figure 2.  The drawback of applying a duplex  bearing is the over -constrained  system that it results in: the shaft is fixed in four single bearings (two too many). This will be a point of  attention when designing the supporting mechanics.  

     Several  µm is  polished from the outer ring of 1 of the 2 single bearings  The inner shaft and the outer housing is assembled  A flange is used to hard preload the bearings  The duplex bearing is preloaded  

Figure 2.  The principle of a duplex bearing  

The bearing properties are summarized in Table 2. These parameters have been established with the aid  of the bearing supplier ADR [1] in cooperation with ESTL [2].  

      several   µm
33 Table 2.  Bearing  properties   Parameter description  Value   Bearing material  Stainless steel 440C   Yields strength bearing material  4000 MPa   Ball diameter  4.762 mm   Loading Configuration  Face-to-face (F F) or X-configuration   Pitch diameter  ∼ 34 mm  Contact angle  30 ± 2 deg   Ball complement  17 [-]  Width (single bearing)  9 mm  Width (duplex bearing)  18 mm   Compliance (1/Stiffness) 0 m/N (‘hard’ preloading) . This is the setting in the  analysis software CABARET. For duplex bearings,  the axial stiffness of the bearing determines the compliance of the ‘preload spring’   Preload  50 ± 10N  Bearing Center Plane sp acing 9 mm  Shaft O.D.  25 mm   Housing I.D.  42 mm   Coefficient of friction (MoS 2 coating ,  applied on both the races and the balls  0.05…0.08  

Motor Choice    Each of the mechanisms is driven with a stepper motor of the same type, the Phytron [5] VSS 52.200.  The motor is procured as a complete unit, including internal bearings. Table 3 shows the main properties.   The redundant motor windings are wound around separate poles. This avoids failure propagation in the  windin gs. The eight  poles are clearly seen in Figure 3 (right) and the picture also shows that each coil is  wound around separate poles . A photograph of  the complete motor is shown on the left in the figure.  

   Figure 3.  Photograph of the Phytron stepper  motor (left) and a photograph of a similar motor  (right) provided by Phytron
34 Table 3.  Motor properties   Parameter description  Value  Remarks   Motor Type  VSS 52.200.1.2    Supplier  Phytron    Number of teeth  50   Number of phases  4 With the 50 teeth this results in a full step  resolution of 1.8 ° (200 steps/rev)   Number of poles  8   Windings  “cold” redundant    Holding Torque  > 0.320 [Nm]    Running Torque  > 0.220 [Nm]   

Conceptual Design Description  

The design of both mechanisms is optimized to the ability to survive launch loads and having a low  friction torque and therefore a high torque margin. Furthermore, the mechanisms are designed as  compact as possible to limit the overall mass  and volume. The principle of the design baseline is identical  for both mechanisms (see Figure 4) and can be roughly divided in the following features:   • A stepper motor drives the carousel of the mechanisms   • The carousel is symmetrically supported by bearings (one duplex at each side)   • The motor and the carousel are coupled through a flexible coupling. T his coupling is torsionally  stiff, but compliant  in all other directions   • The carousel consists of:   o For the DifM: A set of 6 Fused Silica components (diffusers)   o For the FMM: A slit and a folding mirror    The left duplex fixes 5 DoF (only  Rz is free), but due to the hinge the Rx and Ry are released. The X and  Y direction are still constrained since the rotation point of the hinge shaft is placed symmetrically within the duplex bearing. The Z direction is also fixed by the left duplex beari ng. The right duplex bearing is  fixed in X, Y by the membrane. The Z direction is left free which allows for thermal expansion differences between the carousel (Invar) and the fixed world (aluminium). Hereby, the preload of ~50 N remains approximately  the same at large temperature excursions. The Rx and Ry of the right duplex are left free  due to the membrane. Since both the carousel and the motor shaft are fixed in all except the rotation, a  flexible coupling is needed which only transfers the Rz in a stif f manner.
35   Figure 4.   Schematic overview of the concept design of the mechanism  

Detailed Design Features  

Hinge shaft   The elastic hinge is introduced in the design in order to allow for a kinematic design. Therefore, the hinge will ensure that the design is predictable and reproducible even with the presence of tolerance chains within the rotor and relative to the CU housing. The predictability is also guaranteed during the thermal  survival and operational r ange. The following aspects are of importance with respect to the hinge:   • Stresses due to launch loads ( 30g Quasi -Static load in axial and radial direction)   • Bending effects (stiffness and stresses)   • Stiffness in axial and radial direction (driven by the required > 500 Hz)   Figure 5 shows a side view with the two main parameters: the diameter of the hinge (D = 6mm) and the  width (h = 4 mm). The shaft is manufactured from Ti6Al4V , which is a close thermal match (CTE match)  with the 440C  bearings.  

Hard preloaded duplex  bearings (2x)   Phytron stepper  motor  

Flexible coupling  

Membrane to allow  for CTE  compensation in  axial direction  Carousel  

Shaft with elastic hinge   Z  X
36   Figure 5.   A side view of the elastic hinge with the two main parameters:   the diameter of the hinge (D) and the width (h)  

Flexible coupling   The flexible coupling transfers the motor torque to the carousel. Note that this coupling is designed to be  torsional ly stiff (around the center  line) and compliant in the 5 other DOFs. T his coupling is  needed to  enable a kinematic and therefore predictable and reproducible design. Note that the m otor is equipped  with its own bearings and that the motor housing is rigidly attached to the CU housing. The motor shaft  has only one DOF  free and that is the rotation around the center  line. Als o the carousel has only one DOF  free, i.e. , rotation around t he center  line. It is therefore essential that the coupling does not add additional  stiffness  in the remaining 5 DOF s which would result in over -constraining the system. The design of the  coupling is illustrated in Figure 6.  

Figure 6.   Flexible coupling used to couple the drive shaft of the motor to the rotor   rstruts  bstrut Thickness t strut  lstruts Attached to motor   Attached to carousel  Auxiliary body   D h
37 The coupling is a monolithic part (Ti6Al4V), manufactured by wire EDM from one single piece of plate.  The result is two pairs of flanges, mutually coupled by four struts. One flange pair is rigidly mounted to the motor side. Via one pair of parallel struts, the torque is transmitted through the auxiliary body. Next, through the second pair of parallel struts the torque is transmitted to the carousel. The principle of the  coupling is also known as an ‘Oldham’ coupling in literature.    The only drawback of this coupling is the fact that the auxiliary body is not fixed in axial direction. This stiffness is provided by the lateral stiffness of the struts which is designed to be as low as possible. Special attention is given to keep the first eigenfrequency of this auxiliary body sufficiently high.  

Membrane  One of the two duplex bearings of each mechanism is supported by a membrane ( Figure 7) . The reason  for this is to allow for CTE differences (axially) between the carousel and the CU housing. A membrane will result in a predictable design. An alternative would have been to use a sliding fit. However, this is less predictable and reproducible than an elastic solution with a membrane. The dimensions of the Ti6Al4V   membrane are shown in the figure.  

Figure 7.   CAD picture (left) and general dimensions of the membrane  

End stop  Contrary to the DifM, the FMM is equipped with an end stop. The r eason for this is that a higher  reproducibility  can thereby be achieved for the FMM. Using the detent torque of the stepper motor, a  positive torque is exerted against the end stop which enables a stable position with an unpowered motor.  The FMM has only two positions : one of the folding mirror (which has to be positioned reproducible) and  one for passing through the Earth’s radiance (oversize sloth). The end stop is shown in Figure 8 .    A cam disc is rigidly attached to the rotating shaft. Additionally, a key or spline is used to prevent mutual rotation. The cam will rest on a Vespel end stop. This end stop, which is clamped to the end stop holder,  is adjustable with a shim. This material combination was chosen because:   • Vespel combined with metal avoids cold welding.   • Vespel was used in the Sciamachy and OMI projects, also as an end- stop. Both passed a  (comparable) life test . 

b = 29.7 mm  a = 40.5 mm  t = 0.2 mm
38  

Figure 8.  The FMM is equipped with an end stop in order to guaranty an angular reproducibility    < 0.1°. A cam disc will rest on an adjustable Vespel end stop. Note that two stops   are incorporated.  

Bearing analyses  

The following analyses are considered to be relevant for assessing the feasibility of the bearing concept:   • Gapping and stress in the bearings due to axial and radial loads (Q uasi-Static  loads)   • Bearing friction torque due to preload (@ T=20° C)  • Bearing friction torque due to preload (@ worst case operational range)   • Bearing stiffness calculations (radial and axial)  

The bearing analyses have been carried with the assistance of the European Space Tribology Laboratory  (ESTL ) and using the software CABARET. The principle of gapping is explained in Figure 9. Both for  radial and axial gap ping, an unwanted physical gap will be present. The idea is to minimize this gap. The  maximum allowable gapping is set to 20 µm. This is based on the approach followed in the E arthCARE   MSI-VNS mission  [3] and is based on communication with ESA  (European Space Agency) and ESTL.  Due to the maturity of the CU design, the design goal for gapping is set to 10 µm.  

Figur e 10 shows the results for a varying axial load between -2000 to 2000 N. Both the stress and the  axial displacements of each of the two bearings are calculated. The axial force for the mechanism is  calculated with 941 81.9302.3 =⋅⋅==ma Fax N, where the mass m  of the carousel is 3.2 kg and a 30 g QS  load is applied.  The bearings start to gap at the point where the Hertzian stress becomes zero. The  displacement at this point is 1 µm. The displacement at 941 N is 10 µm. Therefore, the gapping at 941N  is 9 um. This is within the design goal of 10 µm. 

The maximum stress  at 941 N is 1800 MPa. With an allowable stress of 4000 MPa, the Margin of S afety  becomes MoS = [ (4000 / 1.25) / 1800 ] - 1 = 0.77 which is > 0 and therefore acceptable.  

Cam disk   End stop  (2x, Vespel)  Key/spline   End stop holder  End stop shim
39   Figure 9.   Radial and axial gapping illustrated  

Figure 10 .  The axial deflection and the Hertzian contact stress in the duplex bearing as a function  of the axial load. The preload is set to 50  N. 

For two reasons, r adial gapping is far less than axial gapping. Firstly, the radial force on the duplex is  estimated to be 941/2 = 470 N which is half the axial force. Secondly, due to the contact angle, the radial   loading is in a much more favorable direction. At 470  N, the radial gap is ~3 µm which is smaller than the  axial gap and thereby also below the design goal of 10 µ m. This shows that the duplex bearing is  considered to be very suited for radial loads.    Gapping analysis (ADR bearing, axial) 0.00200.00400.00600.00800.001000.001200.001400.001600.001800.002000.002200.002400.002600.00 -2000 -1500 -1000 -500 0 500 1000 1500 2000 Axial load [N] (radial 0N)Hertzian contact stress [MPa] -25-20-15-10-50510152025 Axial deflection [um]Stress A [MPa] Stress B [MPa] Deflection A ax [um] Deflection B ax [um]  

      Frad              

     Radial gap              

      Axial gap      Fax