design and modelling of smart structures for payloads antennas

Design and Modelling of Smart Structures for Payload and Antenna (ESA SSPA project)

Frédéric BOSSENS Micromega Dynamics s.a.

Frédéric CUGNON

SAMTECH s.a.

Amit Kalyani Université Libre de Bruxelles

André PREUMONT

Université Libre de Bruxelles Abstract: This paper presents results obtained in the frame of the ESA SSPA (Smart Structures for Payload and Antenna) project. It reports on numerical developments performed to design an active damping interface to be integrated in large space-structures. Emphasis is put on the modelling tools, consisting in a non-linear finite-element model (SAMCEF Mecano) including both structural dynamics and active control. Experimental results obtained on a preliminary active damping interface developed at ULB are also presented 1 Introduction Scientific satellites are increasingly demanding in stability and pointing accuracy. These requirements for future scientific space missions have triggered extensive researches in the area of the active damping/isolation of flexible structure. The main objective of the SSPA project [2] is to design active control systems capable of improving mechanical precision of future scientific space missions. For instance, the future Darwin space interferometer requires a pointing stability of less than 40 nrad for each individual telescope. In this paper, we present active damping interface connecting two structural components. This interface provides uniform control capability and uniform stiffness in all direction using six independent control loops made each of a couple of piezoelectric sensor and actuator. The design of such interface, applied to Darwin space-interferometer is investigated using SAMCEF models. The telescope model is exported to MATLAB in order to design the controller. Once controllers are defined, a SAMCEF Mecano simulation is performed to verify the efficiency of the control under the presence of external perturbations.

9th SAMTECH Users Conference 2005 1/18

9th SAMTECH Users Conference 2005

This paper is organized as follows: section 2 presents background theory about active damping interfaces, section 3 describes Darwin space-interferometer used as target application in SSPA, section 4 presents numerical tools (SAMCEF / MATLAB) used in this study, section 5 explains the design procedure for the active damping interface, section 6 presents some relevant numerical simulations, section 7 presents some experimental results, and describes future experimental setup, finally section 8 gives some concluding remarks. 2 Theoretical background

2.1 Active strut The active system considered in this study is based on active struts, each of them being composed of a linear actuator, colinear with a force transducer.

Figure 1 : Active strut composed of a linear actuator, colinear with a force transducer

It can be shown [1] that, when using such actuator/sensor pair integrated in a structure, with an Integral Force Feedback (IFF) controller and assuming perfect actuator/sensor dynamics, energy can only be extracted from the system (Energy absorbing control). The IFF is defined by the following equation: ∫= dttfGt )()(δ (1)

2.2 Active interface The active strut discussed in the previous section can be integrated into a generic 6 d.o.f. interface connecting arbitrary substructures. Such an interface is shown in Figure 2.

controller

Figure 2 : active inte

r
actuato
rfa

sensor

Flexible tips

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ce


The approach followed in this study is to use a decentralized controller: each active strut is driven by an independent controller as described in Eq (1), with the same gain for all feedback loops.

2.3 Performance predictions Let the dynamics of a passive structure be described by 0=+KxxM && (2) And the dynamics of an active structure (including the active interface) be governed by δaKBKxxM =+&& (3) where the right hand side represents the equivalent piezoelectric loads (for numerical ease, the effect of a linear actuator with stiffness KBa B and stroke command δ can be obtained by applying opposite forces with magnitude KBa Bδ at the extremities of the actuator): δ=(δB1 B, ..., δB6B) P

TP is the vector of piezoelectric extensions, KBaB is the stiffness of

one strut and B is the influence matrix of the interface in global coordinates. The output y = (yB1 B, ..., yB6 B)P

TP consists of the six force sensor signals which are proportional to the

elastic extension of the active struts y = KBa B(q − δ) (4) where q = (qB1B, ..., qB6 B) P

TP is the vector of global leg extensions, related to the global

coordinates by q = BP

TP x (5)

The same matrix B appears in Equ.(3) and (5) because the actuators and sensors are collocated. Using a decentralized IFF with constant gain (a reduced gain g=G KBa Bis introduced for analytical convenience),

ysK

g

a

=δ (6)

The closed-loop characteristic equation is obtained by combining Equs.(3) to (6):

02 =⎥⎦

⎤⎢⎣

⎡⋅

+−+ xBKB

gsgKMs T

a (7)

In this equation, the stiffness matrix K refers to the complete structure, including the full contribution of the Stewart platform. The open-loop poles are ±jΩBi Bwhere ΩBi B are the natural frequencies of the complete structure (see Figure 4). The open-loop zeros are the asymptotic values of the eigenvalues of Equ.(7) when g ∞ ; they are solution of [Ms P

2P + K − B KBa B BP

TP ]x = 0 (8)


The corresponding stiffness matrix is K − BKBa BBP

TP, where the axial stiffness of the legs of

the Stewart platform has been removed from K. Without bending stiffness in the legs, this matrix is singular and the transmission zeros include the rigid body modes (at s = 0) of the structure where the piezo actuators have been removed. However, the flexible tips are responsible for a non-zero bending stiffness of the legs and the eigenvalues of Equ.(8) are located at ±jωBi B, at some distance from the origin along the imaginary axis (see Figure 4). Using numerical tools described in section 4, one can describe the open-loop structural dynamics in state-space representation (characterized by state-space matrices A, B, C, D). This open-loop system can be combined with the IFF controller (characterized by state-space matrices ABcB, BBcB, CBcB, DBcB), as illustrated in Figure 3.

Closed-loop system complete model: Acl Bcl, Ccl ,Dcl

MechanismDynamics

A, B, C, D Leg force sensors

External disturbances acting on the structure

Controller

Ac, Bc, Cc, Dc

Linear actuators commands

Motion of the structure at some points of interest

Figure 3 : closed-loop system model

To estimate the controller performances, the model of Figure 3 can be used to draw a root-locus of the structure: the root locus is built by plotting the eigen values of ABcl B for increasing values of g. Such a locus is represented in Figure 4 (only the upper-half of the symmetric locus is represented).


Figure 4 : root-locus of an active structure with active interface, with 2 sets of f

joints (soft and stiff) In this graph, constant damping lines (oblique) are indicated to estimate thedamping that can be introduced in the controlled structure. The root locus was dr2 sets of flexible tips, some soft ones (large loops) and some stiffer ones (smallThe bending stiffness of the flexible tips mainly influences the frequencies of thωBi B, which in turn determine the size of the loop, thus the performance of thedamping system. 3 Case study, the Darwin space telescope One of the target applications in project [2] is the Darwin space telescopestructure version. A schematic view of the spacecraft is shown in Figure 5.

controller600kg

600kg

1800 kg

1800 kg400 kg

Figure 5 : Darwin-structure space telescope, equipped with 2 active interfac

Ω B

Ω B2B

Ω3B

ωB2 B

ωB3 B

ω

2 B

ω 3 B

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It is basically composed of 4 telescopes, two large ones fixed on a rigid central module, and the 2 smaller ones located at the extremity of carbon-fibre / polymer composite booms. Such structure is designed to performed deep-space interferometric measurements. Active interfaces are located between the booms and the central module. They aim at reducing the relative motion between the telescopes, after disturbances are applied in the central part (e.g. attitude control, cryocoolers…). Figure 6 shows the natural frequencies and corresponding mode shapes obtained with a Samcef finite element model (Figure 7). Not shown on the figure is the pair of torsion modes at 0.27 and 0.29 Hz, respectively.

Fig 4 Mode This section linear mechafrom a mechDynam have and to evaluashould be loc Once sensorsdedicated eledeveloped inacceleration arotations, forcsome addition

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of Darwin-structure (finite-elem

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eeded to incorporate digital crious stages needed to obtain ribed. First, some modal anader to define control objectives observability of the systems;here strain energy is high for a

selected, they can be includePA project, two elements (SE1er to measure nodal displacinked to the structure and the reen two nodes. The measured qreedom (exactly as any other M


ent model)

ontrollers into a non-a mechatronic model

lysis using SAMCEF (modes to be damped) sensors and actuators ll modes of interest.

d in the model using D & SE3D) have been ement, velocity and elative displacements, uantities are linked to CE measure element)

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by introducing some element-related constraints and associated Lagrange multipliers. Doing so, those elements can be used as sensors and as actuator by applying prescribed conditions (force or displacement) on the additional nodes. Those sensor’s and actuator’s nodes can be connected to any control boxes or retained to create a super element. 1

Stewart 1 Service module

D

4

Stewart 22

Figure 7 : Finite element model of the Dar In the finite model of the Darwin interferometertelescopes and the service module are represented by two big telescopes (T3 and T4) and the service modulconsidered as rigid. The small telescopes (T1 and T2)two composite carbon tubes (1mm thick laminate melements), which are connected to the main structudamping interfaces. Table 1 : structural characteristics of Darwin Component Mass [Kg] Ixx [Kg m²]Small telescopes (T1 & T2) 600 900 Large telescopes (T3 & T4) 1800 4300 Service module 400 1100 The active interfaces are made of two rigid plates connleg is modelled as a succession of flexible and stiffright): flexible joint, rigid bar, displacement actuator, fo


T3

T

win space telescope

telescope, the four lumped masses and ine are fixed on the mai are localized at the exade of six +/-45° layre by two hexapod-

Iyy [Kg m²] Izz [K460 880 4500 4500 610 580

ected by six identical elements, namely (frrce sensor and flexibl

T

T

External isturbance Loads

individual ertias. The n structure tremity of

ers – shell like active

g m²]

legs. Each om left to e joint.

7/18


Figure 8 : model of an active leg

Two rigid bodies connected by a bushing element define the joints; the sensor is defined by a lumped mass and a MCE SE3D element; the actuator is modelled by a lumped mass and a bushing in parallel with a distance sensor, rotation and lateral degrees of freedom are constrained by a prismatic slider. Table 2 : stiffness properties of the legs Component stiffness

Axial Bending Shear Torsion

Joint 1E8 22 2.1E6 330 Sensor 1E7 500 1E7 500 Actuator 1E8 ∞ (rigid) ∞ (rigid) ∞ (rigid) Using this model, a super element can be created by retaining measure nodal degrees of freedom of the 12 force sensors, distance degrees of freedom of the 12 actuators, and main nodes of the model (telescopes and service module where perturbation are applied). Once the super element has been created, it can be read into Matlab/Simulink using the “Tsamcef-read-m003.m” sc Tript supplied in the SAMCEF manual. The use of this script is straight forward, and described in the m-file itself. The Mass, Damping and Stiffness matrix of the super element are read, together with a localization vector. After the matrices of the super-element have been read into MATLAB, the user can transform them into a state space model. The state-space format is the format used for controller design. When transforming the mass, damping and stiffness matrix into the T[A,B,C,D]T state space form it is important to remove the fixations from the super element. All this, can be done automatically using “ Tsamcef-stsp-m003.m” T Matlab/Simulink script supplied in this manual.

Once the controllers have been designed, their performances can be evaluated by numerical analyses simulating structural behaviour of the structure under several perturbations. Those dynamic transient simulations can be done using the state space model in MATLAB/Simulink or by introducing the controllers in a SAMCEF Mecano model. Digital controllers that have been designed with the help of MATLAB / Simulink can be exported to ANSI C-code; these ANSI C-controllers can then be compiled and linked with Mecano, so that the digital controller element has access to all nodal variables during the simulation. Because the controller and SAMCEF exchange

joint joint Displacement actuator

Force sensor


data at every sample time of the controller the total calculation time can become excessively long in the case of a digital controller with a high sample frequency. A solution for the problem is provided with the help of the continuous controller option. In this case the controller will no longer impose a time step to MECANO, but the controller and SAMCEF will only exchange data at the time steps imposed by SAMCEF. Even though the controller will not exchange at every sample time with SAMCEF it will be integrated in time using its own sample time. Furthermore it will not only supply the output value of the controller at the given exchange times, but it will also calculate the appropriate Stiffness/Conductivity, damping/capacity and mass matrix for SAMCEF. In this case the controller will act as a continuous "bushing type" element with time dependent properties. This approach is only valid if the sample time of the controller is smaller than the time step of the SAMCEF calculation and if the control box is a user defined element or a predefined controller; MATLAB Simulink control boxes defined by C-code are always digitals.

In the adopted design of the Active damping interface, the 12 independent control loops are simply integrators (see Equ.1) so they can be modelled using the SAMCEF predefined PID continuous controller. Three models have been used. First, a model with few elements has been defined using the super element described above and 12 DIGI elements; as number of unknowns is low, it provides a quick answer (about a minute). Secondly, the 12 DIGI elements are included in the detail model taking into account all possible non-linearities. Finally, a MATLAB model defined from the state space form and internal controllers has been used. All models have given almost identical results; this validate the PID control element available in SAMCEF by comparison to standard control simulation tool and verifies the small displacement assumption made for the linearized super element model. All further shown results are obtained from the Mecano model using a super element to describe the structure.

5 Active interface design As explained in section 2.3, the performances of the active interface will depend on the stiffness of the flexible joint (mainly in bending and in torsion). It also depends on the actuator stiffness, and on the geometry of the active interface. Actually, four parameters were optimized during the design process: (i) the global size of the interface, (ii) the leg inclination in the interface (see Figure 9), (iii) the actuator axial stiffness and (iv) the flexible tip design.

α

leg inclination global sizing factor

Figure 9 : geometrical parameters optimized in the active interface design


5.1 Actuator stiffness The actuator stiffness mainly influences the overall stiffness of the passive structure, therefore its natural frequencies (ΩBi B in Figure 4). Figure 10 shows how Ω Bi Bvaries with actuator stiffness. For actuators with axial stiffness > 2 N/µm (which is very common for piezo-translators), the loss in the natural frequencies is lower than 2%, (as compared to a passive structure where the interfaces are completely rigid). This issue is thus not critical as long as actuator is not too soft axially (as well as the flexible tips). For this Darwin application, it was decided that the actuator stiffness had to be higher than 4N/µm

Figure 10 : variation of passive structure natural frequency (Ω Bi B in Figure 4), in function

of the actuator axial stiffness

5.2 Overall sizing The overall sizing play similar role as the actuator stiffness: it influences mainly the open-loop poles of the system (Ω Bi B). By increasing the size of the interface (keeping the axial stiffness individual legs constant), one increases the lever arm and thus the bending and torsion stiffness of the interface. In Darwin-case, it was found out that, building an active interface with diameter higher than 25cm was not useful, because it wouldn’t increase substantially the natural frequencies ΩBi B of the system.

5.3 Flexible tips The bending and torsion stiffness of the flexible tips plays a critical role in the active system performances, since it determined the frequencies of the zeros (ωBi B), and thereby the size of the loops in the root locus. This effect is illustrated in Figure 4, showing the root-locus of an active structure for two types of flexible tips. The softer flexible tips lead to much larger loops in the root locus.

Figure 11 : Flexible tips to be used with the active interface of Darwin


5.4 Leg inclination Leg inclination influences the active damping performances in two ways:

• Placing the legs horizontal (α=0°) will maximize the torsion stiffness (and the control authority) of the active interface, and minimize its bending stiffness, while placing the legs vertical (α=90°) would produce the inverse effect. This effect is illustrated in the root-locus of Figure 12.

• It allows to chose the configuration that will produce good active damping performances for all considered modes with a single gain (this is not the case in general). This effect is illustrated in Figure 13.

Figure 12 : influence of the leg angle on the root locus of the controlled structure

5.5 Feedback gain selection Figure 13 shows how the modal damping varies in function of the IFF control gain (G in Equ.1, for α=25° (left panel) and α=50° (right panel).

bending 1

bending 2

bending 3

torsion

bending 1

bending 2

bending 3

torsion

Figure 13 : influence of the leg angle on the root locus of the controlled structure


For α=25°, a large amount of damping can be introduced into bending modes #1 and 2, for a gain G ≈ 2-3•10P

-5P but for that gain, the amount of active damping in the torsion

modes is rather small. For α=50°, this situation improves substantially: for G ≈ 2-3•10P

-5

P(see vertical line), the active damping in the bending modes #1 and 2 is still very large, and it is more than doubled for the torsion modes. Such study led to selecting an angle of 50° for the active leg’s inclination in the interface. 6 Simulation results

6.1 Active damping performances In order to evaluate the performances of the active platforms in time domain, several simulations have been done with and without active control in order to highlight the effect of the active damping system. Three loading cases have been considered: transverse force (Fx), bending (Mx) and torsion (Mz) moments applied on the service module. For each, a unitary magnitude is used with 3 different time functions: • f1(t) is a sine-sweep function with frequency linearly increasing from 0 to 1 Hertz

during the [0, 300 sec.] time interval. • f2(t) is a sinusoidal function (0.05 Hz) from time 0 to time 100 sec.; then frequency

continuously increases to reach 1 Hz at time 300 sec (sine-sweep). • f3(t) is an alternating impulse; piecewise linear function joining 0, 1and –1 in a 0.2

sec. time interval in order to excite all modes below 5 Hz. All these disturbances are applied on the central module (see Figure 7). The performances of the active system will be evaluated in terms of the strokes of the actuators, the measured force in the legs and the relative rotations between the small telescopes and the large ones fixed to the rigid central module.

-15

-10

-5

0

5

10

15

0 50 100 150 200 250 300

time (sec)

forc

e (N

)

no controlwith control

-0.08

-0.07

-0.06

-0.05

-0.04

-0.03

-0.02

-0.01

0

0.01

0 50 100 150 200 250 300

time (sec.)

disp

lace

men

t (m

m)


Figure 14 : measured force and actuating displacement (leg 1 of interface 1) resulting

from disturbance function f1 applied along x. Figure 14 shows typical behaviour of a leg during perturbation; in order to reduce the force measured in the sensor, the controller imposes some displacement in the actuator. Next figures compare the pointing accuracy (relative displacement and rotation between large and small telescopes) when structure is passive or active under previously defined perturbations; only main deformations are plotted.

-10

-8

-6

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-2

0

2

4

6

8

0 50 100 150 200 250 300

time (sec)

rela

tive

x-di

spla

cem

ent (

mm

)


-0.002

-0.0015

-0.001

-0.0005

0

0.0005

0.001

0.0015

0.002

0 50 100 150 200 250 300

time (sec)

rela

tive

y-ro

tatio

n (r

ad)


Figure 15 : Pointing accuracy under disturbance function f1 applied along x.

-0.0015

-0.001

-0.0005

0

0.0005

0.001

0.0015

0 50 100 150 200 250 300

time (sec)

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tive

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ent (

mm

)


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-400

-200

0

200

400

600

0 25 50 75 100 125 150

time (sec)

rela

tive

y-ro

tatio

n (n

rad)


Figure 16 : pointing accuracy under disturbance function f3 applied along x.

-0.0008

-0.0006

-0.0004

-0.0002

0

0.0002

0.0004

0.0006

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0 50 100 150 200 250 300

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ad)


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100

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0 25 50 75 100 125 150

time (sec)

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tive

x-ro

tatio

n (n

rad)


Figure 17 : pointing accuracy under Mx*f1 and Mx*f3

-0.0006

-0.0004

-0.0002

0

0.0002

0.0004

0.0006

0 50 100 150 200 250 300

time (sec)

rela

tive

z-ro

tatio

n (ra

d)


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-400

-200

0

200

400

600

0 25 50 75

time (sec)

rela

tive

z-ro

tatio

n (n

rad)

100


Figure 18 : pointing accuracy under Mz*f1 and Mz*f3



The damping effect is more spectacular when an impulse is applied (f3 time function) than when a forced excitation (f1 or f2 time functions) is maintained. This is because the effect of active damping is visible only close to structural natural frequencies, and f1 excite the entire range of frequency while f3 concentrates the structural response close to the natural frequencies. Furthermore, modal damping can be estimated from free responses obtained with f3. From Figure 16 (left panel), where the response is dominated by bending mode #1, one can see that with active control, the oscillation is over-critically damped (no more oscillations): ξBbending1B> 70% In Figure 16 (right panel), the residual oscillating response corresponds to bending mode #3. The amplitude decreases by a factor 2 in about 3 periods of the oscillation. This corresponds to ξBbending3 B≈ 5%. In Figure 18 (right panel), the response is dominated by the torsion modes. With active control, the vibration amplitude decreases by a factor 2 in about 1 period, which corresponds to ξBtorsion B≈ 10%. These numbers are in good agreement with the active damping predictions obtained (for a control gain of 3 10P

-5P).

Last simulation consists in verifying that the structure behaves properly when some active legs stop working. For this test, legs 1 and 2 of platform 1 and leg 2 of platform 2 are passives. Next figure shows that the pointing accuracy is lightly affected illustrating the robustness of the decentralized active damping system concept.

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12 active legs9 active legs

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0

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0 50 100 150 200 250 300

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rela

tive

x-di

spla

cem

ent (

mm

)

12 active legs9 active legs

Figure 19 : pointing accuracy with damaged platforms (Fx*f1)

6.2 Envelope definition As the Darwin mission is still at some preliminary stages, lack of data concerning external perturbation forced us to define an envelope of excitation level under which the implemented damping system works well enough to meet optical requirements (less than 40 nrad of relative pointing of coupled telescopes) under technical constraints of active components (maximal actuators stroke of 20 µm). From those requirements,


computed values and linear assumption, one can estimate maximum allowed perturbations summarized in next table. Table 3 : Envelope for uncoupled external perturbations

Max. relative pointing error around axes (nrad)

Unitary external load x-axis y-axis z-axis

Actuating maximal displacement

Maximal allowed load

Maximal load w/o control

Fx*f1(t) 5.8 1375000 3600 70.5 µm 0.029 mN 0.024 mN Fx*f2(t) 2.9 1045000 2400 52 µm 0.038 mN 0.009 mN Fx*f3(t) 10P

-4P 480 0.2 9.3 µm 83 mN 73 mN

Mx*f1(t) 405000 0.4 0.4 38.6 µm 0.098 Nmm 0.056 Nmm Mx*f2(t) 347000 0.2 0.2 19.4 µm 0.115 Nmm 0.087 Nmm Mx*f3(t) 220 10P

-5P 10P

-5P 3.7 µm 182 Nmm 130 Nmm

Mz*f1(t) 996 5520 151000 5.3 µm 0.265 Nmm 0.074 Nmm Mz*f2(t) 221 2620 145000 4.1 µm 0.276 Nmm 0.088 Nmm Mz*f3(t) 10P

-4P 0.1 487 0.005 µm 82.1 Nmm 80.8 Nmm

If ones allow a transient phase after start of perturbation for which the pointing is disturbed, a new envelope assuring acceptable behaviour of the system after 60 seconds of damping is available in next table. Table 4 : Envelope for uncoupled external perturbations (after 60 seconds)

Max. relative pointing error around axes Unitary external load x-axis y-axis z-axis

Actuating maximal displacement

Maximal allowed load

Maximal load w/o control

Fx*f1(t) 5.8 86000 740 70.5 µm 0.465 mN 0.024 mN Fx*f2(t) 2.9 630000 750 52 µm 0.064 mN 0.009 mN Fx*f3(t) 10P

-4P 2.5 0.03 9.3 µm 2150 mN 82 mN

Mx*f1(t) 50000 0.4 0.4 38.6 µm 0.800 Nmm 0.056 Nmm Mx*f2(t) 265000 0.2 0.2 19.4 µm 0.151 Nmm 0.087 Nmm Mx*f3(t) 0.5 10P

-5P 10P

-5P 3.7 µm 5405 Nmm 135 Nmm

Mz*f1(t) 996 5520 151000 5.3 µm 0.265 Nmm 0.074 Nmm Mz*f2(t) 221 2620 145000 4.1 µm 0.276 Nmm 0.088 Nmm Mz*f3(t) 10P

-4P 0.1 0.24 0.005 µm 167 Nm 0.085 Nm

7 Experiments The SSPA project was initiated after ULB had performed theoretical and experimental work concerning generic active damping interfaces [3,4]. In this section, we summarize the experimental results obtained with a truss fit on an active damping interface (Figure 20), and the development of a demonstrator in the frame of SSPA.

Figure 20: Stewart platform with piezoelectric legs as generic active damping

interface. (a) General view. (b) With the upper base plate removed. (c) Interface acting as a support of a truss.

Figure 21 : Experimental time response and FRF of the truss mounted on the active

interface. The six independent controllers have been implemented on a DSP board; the feedback gain is the same for all the loops. Figure 21 shows some typical experimental results; the time response shows the signal from one of the force sensors of the Stewart platform



when the truss is subjected to an impulse at mid height from the base, first without, and then with control. The FRFs (with and without control) are obtained between a disturbance applied to the piezoactuator in one leg and its collocated force sensor. One sees that fairly high damping ratios can be achieved for the low frequency modes (4−5Hz) but also significant damping in the high frequency modes (40−90Hz). In the frame of SSPA project, the following experimental setup is currently under development, as shown in Figure 22. This demonstrator is representative of Darwin (to some scaling factor), and will be installed on a soft-spring gravity-compensation system, to approximate the free-free boundary conditions.

50 kg

50 kg

6 kg

6 kg

6 kg

3 m

0.5 m

Figure 22 : Experimental setup developed in SSPA

8 Conclusions This paper reports on the latest developments in SAMCEF, related to the design and modeling of active structures. All features are now available to design an active structure (generation of a MATLAB Simulink-compatible model), and to test it in fully non-linear environment (actuators, sensors and controller integrated in SAMCEF Mecano model). Such developments were a necessary step in the SSPA project, aiming at designing and building active systems for damping large space structures. Darwin space interferometer was taken as target application in SSPA project, and used as numerical benchmark in this study. Simulations presented here show an excellent agreement between the theoretical predictions and numerical experiments, which assesses the reliability of the newly-developed tools. Furthermore, active damping performances predictions obtained by numerical simulation are promising, in the sense that the active system reduces efficiently the impact of external disturbances on large flexible structure. The efficiency of the active system was experimentally validated through a preliminary experiment carried out at ULB. A larger scale-experiment, representative of a large space-structure is currently under development in the frame of SSPA.

9 Acknowledgement The authors greatly acknowledge support from ESA within the frame of SSPA project (contract n°17114/03/NL/SFe), and in particular its technical representative D. Sciacovelli. 10 References [1] Vibration Control of Active Structures, An Introduction, Kluwer Academic Publishers, 2002 (Second edition), February 2002 [2] Smart Structures for Payload and Antennae, ESA/ESTEC Contract n° 17114/03/NL/SFe. [3] A. Abu Hanieh, M. Horodinca & A. Preumont, Six-degrees-of-freedom hexapods for active damping and active isolation of vibrations. Journal de Physique IV, Vol. 12, pp. 11-41, December 2002. [4] A. Abu Hanieh Active Isolation and Damping of Vibrations via Stewart Platform, Ph.D Thesis, ULB, April 2003 [5] SAMTECH, “SAMCEF” Users’ Manual, v11.0-5, December 2004.


design and modelling of smart structures for payloads antennas

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