The Large Binocular Telescope primary mirror support
control system description and current performance results
David S. Ashbya , Jonathan Kerna , John M. Hilla , Warren B. Davisonb , Brian Cuerdenb ,
Joar G. Brynnela , Chris Biddicka , Kenneth Duffekb
a Large
Binocular Telescope Observatory, University of Arizona, Tucson, AZ, 85721, USA;
b Steward Observatory, University of Arizona, Tucson AZ, 85721, USA
ABSTRACT
The Large Binocular Telescope (LBT) is built around two lightweight borosilicate honeycomb mirrors which, at
8.4 meters in diameter, are the largest operational examples of this technology. Since the mirrors are relatively
stiff, the LBT mirror support system relies on passive position control and active force control. Passive position
control is performed by six extendable hardpoints organized as a truncated hexapod, which may be positioned
as required by the active optics control loop. The hardpoints rely on their axial stiffness to maintain the mirror
position against residual external disturbances. The active force control system minimizes the force exerted by
the hardpoints on the glass. Additionally, the axial component of the nominally uniform active support forces
can be perturbed to distort the mirror as required by the active optics control loop. Because of the relatively
large CTE of borosilicate glass, the differential temperature of the mirror is critical. Thus, the force control
system must support a 16 metric ton mirror using less than 100 Watts of electrical power. The authors present
a description of the primary mirror support system as implemented at the LBT. Initial stability problems made
the mirrors nearly unusable in freezing temperatures. The authors explain the reason for this instability and
describe the solutions implemented. Data demonstrating the current performance of the primary mirror support
system are also presented.
Keywords: LBT, mirror support, borosilicate, honeycomb, load cells, telescope
1. SYSTEM OVERVIEW
The Large Binocular Telescope is built around two 8.4 meter borosilicate honeycomb mirrors on a common
mount. These mirrors are each located in an associated primary mirror cell. The overall function of the primary
mirror cells is to position each primary mirror with respect to the telescope and provide an environment that
limits distortions. Three essential elements achieve these ends: the mirror support system, the mirror positioning
system and the ventilation and thermal monitoring system.
The most complex element of the primary mirror cells is the mirror support system. The goal of the mirror
support system is to counter the mirror weight as if it were neutrally buoyant, floating in a fluid of a density
equal to that of borosilicate glass. When the support system is working properly, the mirror seems as though
it were nearly weightless. To facilitate this behavior, the axial force on each of the six hardpoints is measured
and the resulting vector is transformed into the six Cartesian force and moment components in the platform
reference frame. Six independent control loops (one for each degree of freedom) then drive the force and moment
components to zero by commanding 108 dual-channel and 52 single-channel pneumatic force actuators. The
forces applied by each actuator is transmitted to the mirror through loadspreaders which are bonded to the rear
surface of the mirror. When the active system is not operating the mirror is supported passively by a network of
elastic elements known as static supports. The static supports are essentially springs which loosely connect the
load spreaders to the cell. These connections are such that the static supports do not contact the load spreaders
Further author information: (Send correspondence to D. Ashby)
David S. Ashby: E-mail: dashby@as.arizona.edu, Telephone: +1 520 626 9238
Jonathan Kern: E-mail: jkern@as.arizona.edu, Telephone: +1 520 626 1485
John M. Hill: E-mail: jhill@as.arizona.edu, Telephone: +1 520 621 3940
Figure 1. In its most fundamental capacity, each LBT mirror cell serves to support the weight of the associated mirror.
The weight can either be supported passively by the static supports or actively by the collective force from the 268
pneumatic cylinders integrated into 160 force actuators. This diagram illustrates the active and passive force control
features of the LBT primary mirror cells. The objective of the active force control system is to drive the force detected
by the six hardpoints to zero by manipulating the individual actuator forces.
unless the mirror has moved several millimeters from its nominal position. The full mirror support system is
illustrated in Figure (1).
The second element of the primary mirror cells is the position control system. The hardpoints control the
position of the mirror and also provide the necessary system stiffness. However, if the axial hardpoint force
becomes too excessive the hardpoints are designed to “breakaway” and allow the mirror to move until the load
spreaders contact the static supports. Integrated into each of the six hardpoints is a motorized roller screw and an
axial LVDT to facilitate the positioning of the mirror in all six rigid body dimensions. Thus low frequency force
disturbances are countered by the active support system and higher frequency force disturbances are countered
by the hardpoints. The axial stiffness of the hardpoint should be between 80 and 120 Newtons per micron to
complement the stiffness of the mirror. The force actuators and hardpoint breakaway mechanisms are pneumatic.
The air supply for these mechanisms is controlled by a central safety system. In the event that an unsafe condition
is detected by the hardware or software, a cell can “panic”. This results in the compressed air bleeding from the
system and the mirror settling gently onto the static supports.
The final element of the primary mirror cells is the ventilation and thermal monitoring system. This system
controls the thermal distortions and the “telescope seeing” resulting from the mirror being a different temperature
than the outside air. Each cell contains approximately 1,400 Air Nozzles which inject temperature controlled air
into the hollow cores of the mirror. As the surface area of a honeycomb mirror is very large, with the ventilation
system operating, the mirror temperature can be steered relatively quickly. A total of 91 thermal couples are
used to monitor the temperature of each mirror, the associated mirror cell weldment and the air temperature
inside of the associated cell.
2. LOADSPREADERS
The honeycomb mirror structure is not infinitely stiff, consequently print-through (shear deflections under gravity) on the optical surface appears above each supported point. Supporting forces are applied under rib intersections of the honeycomb structure where the local strength and stiffness are greatest. The number of supported
points is determined by these local print-throughs rather than by the global deformations of the mirror. A spacing of 384 mm, or two honeycomb cells between support points reduces support print-through to an insignificant
level. This results in 426 axial support points, more than the number of actuators needed to control the shape of
the mirror. The axial forces on these points vary with the local thickness, and thus the mass they must support.
Figure 2. The purpose of the loadspreader is to distribute the forces required to support the weight of the mirror. Adhesives
are used to attach the loadspreader pucks to the rear surface of the mirror in the distribution illustrated in the figure on
the left. The figure on the right is an illustration of a 3-puck loadspreader, which is the most numerous variety. Dual and
even single-puck interfaces are also used.
Close points have similar forces so loadspreaders can be used to group support points and reduce the number of
actuators. We have worked out support patterns using single, dual, or triple support points above each actuator.
160 support actuators are necessary to correct the global figure of the mirror under a changing gravity load and
these forces are in turn distributed by loadspreaders onto the 426 pucks bringing the local print-through above
each pad below the optical error budget. These pucks are Invar/steel pads which are permanently glued to the
back of the mirror.
The loadspreaders and actuators also have to support the mirror laterally. The single and dual loadspreaders
are not designed for sheer forces so the lateral forces must all be on triple loadspreaders. The lateral force is
applied below the backplate of the mirror so there is an overhanging moment that must be corrected by axial
forces. The force on a loadspreader is thus the axial force times the sine of the elevation angle plus the lateral
load times the cosine of the elevation angle plus an axial compensation load for lateral times the cosine of the
elevation angle. The axial and lateral force vectors are combined to a single vector.
3. STATIC SUPPORTS
The passive support system consists of a network of wire rope basket springs known as static supports which are
located under the loadspreaders. When the active support system is not operating, the mirror is supported by
these static supports which deflect in a very nonlinear manner: they are stiff when initially contacted and get
soft under load. Their lateral spring constant is also similar to that of the axial direction. The goal is to have
the static supports deflect much more than the cell so there is a relatively even support force. This system is
designed to resist the gravity load of the mirror in any telescope orientation plus an additional 0.3 g load. If
air pressure or power were to fail with the mirror in operating position, the mirror experiences forces above the
normal gravity load. When acceleration exceeds the hardpoint forces, they “breakaway” and the mirror translates
until the loadspreaders contact the static supports. The static supports must be able to safely decelerate the
mirror when it drops so they are designed with a spring rate to resist a 0.3 g load, in addition to gravity in
any direction with a deflection of less than 3 mm. The static supports must also protect the mirror against
unexpected accelerations which might arise from earthquakes, telescope collisions, drive failures, etc. Normally,
Figure 3. The static supports are nothing more than wire-rope springs which loosely connect the loadspreaders to the
mirror cell structure. Their purpose is to passively support the weight of the mirror when the active support system is
turned off. Also pictured here is a “glass wedge” where one of the hardpoints interfaces to the mirror.
these static supports are close to the loadspreaders but not in contact with them. The static supports must stop
the mirror motion before the mirror support actuators or other parts of the mirror cell hit their hard stops. The
static supports are allowed 3 mm of travel to absorb 1.3 g of load. Another 3 mm may be used to account for
imprecise placement of the pads.
4. ACTUATORS
Figure 4. The actuators apply the force necessary to support the weight of the mirror. The bulk of the actuators used
in the LBT primary mirror cells are dual channel as shown here. These units are capable of generating an axial force of
3000 N and a lateral force of 2100 N while dissipating only 700 mW of electrical power. The image on the right show a
cut-away view of a ball decoupler which allows the actuator to apply force to the mirror while also permitting the mirror
to move with the hardpoints.
When the active support system is operating, the mirror is supported by 160 pneumatic force actuators.
The actuators apply axial and lateral forces to the mirror in reaction to its weight. Most of the actuators are
Figure 5. This is a block diagram of a single-channel force actuator. Force commands from the mirror cell computer are
locally regulated at the actuator. A ball decoupler allows the mirror to translate as required by the hardpoints while
a flexure limits the magnitude of unwanted coupling moments. The actuator controller is often referred to as the inner
control loop.
dual-channel, supporting the lateral force caused by the mirror weight at off-zenith angles and simultaneously
providing the required axial support. Four of the dual-channel units are oriented to provide force in the cross
lateral direction. The major components of an actuator are identified in Figure (4). The components present
on each channel of an actuator are: a load cell, a ball decoupler, a flexure, a pneumatic cylinder, two pressure
transducers and an actuator controller.
The actuator load cell measures the axial force applied by the pneumatic cylinder. This force is regulated by
the inner control loop by changing the cylinder pressure. Each load cell is capable of measuring 4500 N .
The ball decoupler allows the mirror to move in the lateral and cross lateral directions without appreciably
changing the load applied to the mirror. The ball decoupler on the diagonal axis of a two axis unit allows the
mirror to move in a tilted plane perpendicular to the diagonal axis. Since both air cylinders are free to move
axially, the mirror movement is in no way constrained by the actuator. A cut-away of the ball decoupler is shown
in Figure (4).
The actuator flexure which is on the mirror side of the ball decoupler is necessary to accommodate mirror
rotation. The flexure also accommodates slight misalignments of the actuator with respect to the mirror without
imparting potentially large moments to the mirror.
The pressure transducers are proportional analog pressure regulators. A voice coil is driven by the input
voltage which regulates flow through an orifice. Changes in this flow ultimately determine the pressure at the
output port of the device. While not linear, this is acceptable since the actuator controller linearizes the response
by controlling the measured force at the load cell. Two pressure transducers are required for each channel of an
actuator. Each transducer regulates the pressure on one side of a piston.
Finally, the actuator controller consists of a microcontroller and at least one analog servo controller. Force
commands received from the mirror cell computer are converted to analog and used to command the analog force
control loop. This controller is often referred to as the inner control loop. See Figure (5) for details.
5. HARDPOINTS
The hardpoint’s function is simply to place and retain the primary mirror in the desired position with respect to
the cell. There are six hardpoints arranged in three pairs that form a truncated hexapod as shown in Figure (6).
The six hardpoints interface to the mirror though a glass wedge which is shown in Figure (3). Extension and
contraction of the six hardpoints allow the mirror to be moved in all six rigid body dimensions. Extension and
contraction in small precise increments is accomplished with a motor, gearbox and roller screw combination.
Figure 6. The six hardpoints act against three thrust blocks locate in the bottom of each mirror cell. A pair of hardpoints
are nearly orthogonal to one another. Note the counter weights in the image on the right. These weights counter the
moment of the structure above the flexures.
An independent measurement of hardpoint displacement is performed by a axially mounted LVDT. Flexures are
incorporated that tilt when the mirror moves while the hardpoint base remains stationary.
It is necessary to measure the load on the hardpoint so that it can be used to control the actuator forces.
This is done by a load cell which measures the axial load applied to the mirror plus the weight of the extension
arm and the counterweights. The counterweights reduce the lateral load (lateral relative to the hardpoint axis)
transmitted to the mirror to essentially zero. Two sets of counterweights are needed because the lower flexures
are separated by 45 mm and the hardpoint components connected above the middle flexures are balanced about
each flexure. The heavy counterweights attached to the extension arm balance the mirror side components about
the mirror side lower flexure. The small counterweights, attached between the two lower flexures, shift the CG
back an additional 45 mm to balance the mirror side components about the cell side lower flexure. The weight
of the extension arm and counterweights that are acting along the hardpoint axis must be subtracted from the
indicated load cell force to arrive at the net axial load being supported by the hardpoint for any elevation angle.
A carbon fiber reinforced extension has been designed that reduces the mass above the breakaway by about 30%,
but this has not yet been implemented.
The hardpoints sense the net forces and moments on the mirror and the actuators apply compensating forces.
The bandwidth of this outer control loop is about 5 Hz, so it provides excellent resistance to wind in terms of
rigid body motion as well as bending. At frequencies about half the bandwidth to higher frequencies the stiffness
of the hardpoints and the mirror stiffness resist deformations. At frequencies well above the bandwidth the
simple mass of the mirror is enough to resist deformation.
The general layout of the hardpoint is shown in Figure (7). The base at the left is attached to the cell and
the rightmost flexure is bolted to a Ni-resist puck (a nickel alloy with the same CTE as the mirror) bonded to
an E6 glass wedge with a 0.5 mm thick Q3-6093 RTV bond. The glass wedge is bonded to the mirror with a
0.002/0.003 Norland 61 adhesive bond. The glass-wedge-to-mirror bond is loaded in shear by axial load in the
hardpoint so it must have high shear stiffness. The puck-to-glass-wedge bond is loaded in compression and a
low shear stiffness bond is permissible (shear compliance in the bond accommodates the small differential CTE
between the puck and the glass wedge).
5.1 Breakaway Mechanism
The hardpoint has a breakaway mechanism that extends or contracts when a specified load is reached. This is
required for glass safety because the stress in the glass must be kept small (0.7 M P a). The break-away mechanism
is a pneumatically activated system that fixes the location of the shaft that exits the mirror side of the hardpoint
Figure 7. The hardpoint is fundamentally a linear position actuator used to control the position of the mirror. The axially
mounted LVTD measures the position of the mirror with respect to the hardpoint base. Flexures allow the mirror to
translate and rotate with respect to the hardpoint base and a load cell measures the axial force applied to the mirror.
Figure 8. Illustrated here is a more detailed drawing of the hardpoint breakaway mechanism. The mechanism allows the
hardpoint to remain mechanically stiff as long as the applied force remains small. When the compression or tension force
becomes excessive, the hardpoint extends or retracts as necessary to control the force.
body relative to the hardpoint body at loads below about 300 lbs, but allows extension or contraction at higher
loads. There are two pistons having equal net area of 84.6 cm2 sliding on the shaft. Pneumatic pressure is
introduced between the two pistons driving piston 1 to the seat on the shaft, and piston 2 to the adjustable stop
at the left end of the shaft in Figure (8). The net force on each piston with a pneumatic pressure of 0.26 M P a
is 1431 N. The end stop is adjusted to leave a minimal gap and no interference. The springs push on piston 2
driving the shaft to the right until piston 1 contacts the main housing cap seat. Compressive breakaway occurs
when the force between piston 1 and the seat on the main housing (not the shaft) drops to zero. This force is:
F1h = Fspring − Fcomp .
(1)
Where: F1h is the force between piston 1 and the main housing, Fspring is the force of the springs pushing on
piston 2 and Fcomp is the compressive force (a positive number).
After compressive breakaway, the shaft moves through the minimal gap clearance between piston 2 and the
main housing. Once piston 2 contacts the housing, the shaft is again constrained axially until the compressive
load increases to the pneumatic pressure times the effective area of piston 2 at which point the shaft slides
through piston 2 until piston 1 contacts piston 2 or the shaft contacts the main housing. Which contact occurs
depends on the motorized axial extension of the hardpoint since the breakaway shaft may contact the stationary
end of the roller screw shaft when the main housing has been contracted. If the clearance between piston 2 and
the main housing is negative, the shaft will have axial play equal to the amount of interference. This condition
is not allowed. Tension breakaway occurs when the force between piston 1 and the shaft seat drops to zero. This
force is:
F1s = P A − Fspring − Ftension
(2)
Where: F1s is the force between piston 1 and the shaft, P A is the pressure force on piston 2 and Ftension is
the compressive force (a positive number). If the shaft continues to extend after a tension breakaway the spring
force relaxes away to nothing after a few mm of travel leaving a net tensile load equal to P A.
LBT hardpoints suffer from low initial breakaway forces. This is because of the relatively large mass of the
extension and counterweights that must be supported by the breakaway mechanism. The axial compression
component of the weight is highest when the hardpoint is oriented vertically. In some cases, as the telescope
approaches horizon pointing, some of these gravitational forces appear as tension at the breakaway mechanism.
Thus either the compression or tension breakaway forces may be undesirably decreased depending on the telescope
orientation.
We plan to reduce the weight of the shaft extension and counterweights thereby increasing the force that
can be applied to the mirror without compromising safety (the mirror force is the breakaway force ± the axial
component of the shaft extension weight, which varies with elevation angle).
6. CHALLENGES AND RESOLUTIONS
Initial experiences with the first integrated mirror cell at the Seward Observatory Mirror Laboratory proved
positive. However, after installation, the system began to demonstrate evidence of instability. The instability
manifested itself as a force oscillation observable by the hardpoint load cells at a frequency between 1.2 and
1.8 Hz. The oscillation was most commonly triggered when the ambient temperature fell below about 5◦ C
and the hardpoints were moved. The symptoms would become progressively worse as the temperature fell. At
temperatures below -2◦ C the mirror cell was rendered unusable. The operating temperature range of the LBT
primary mirror cells is -20 to +25◦ C. The instability characteristics were:
a. The outer control loop gains impacted the amplitude of the oscillation but not the frequency.
b. The oscillation was most commonly triggered by hardpoint motion. However, it was possible to trigger the
oscillation by enabling or disabling actuators near the edge of the mirror.
c. The oscillation was most sensitive to motion in the axial direction.
d. If the temperature was sufficiently low, the oscillation could be triggered with the outer control loop
disabled.
Though it was clear that the actuators were responsible for the instability, recreating the effect using an
individual actuator in a laboratory environment proved to be very difficult. An individual actuator showed no
evidence of instability at low temperatures when acting against a rigid restraint. Compliance was then added
to the restraint but no signs of instability were observed, regardless of the temperature. Finally, a movable
restrain was constructed that could be cooled to sub-freezing temperatures. Only then could the instability be
demonstrated in a controlled environment.
Identifying the offending component also proved to be difficult. The possible candidates were eventually
reduced to the following:
a. The Krytox grease used in the cylinders and the ball decouplers grew very stiff as the temperatures fell
below freezing.
b. The amount of play in the ball decouplers shrank as the temperature fell due to the use of materials with
different thermal properties.
c. The shaft bushings used in either end of the cylinders grew tight at low temperatures. The tolerances of
these bushings were also loose which produced variable results depending on the unit tested.
d. The Viton seals used to isolate the volumes on either side of the piston grew stiff at very low temperatures.
The seals showed no sign of leakage and the manufacture speced the seals for a much lower temperature
than required for this application.
In order to understand which of the above changes could cause the instability, a detailed Simulink model of
the servo was then developed. This model suggested that compliance in the actuator and/or in the constraint
results in an acute sensitivity to mechanical friction. As the friction is increased, stability margins decreased.
In fact, the measured friction is far too high to achieve the desired actuator bandwidth of 5 Hz in the presence
of the measured axial actuator compliance. However, if the friction is decoupled using an elastomeric element,
mechanical bandwidth can be restored. Seal compliance allows the shaft to move the small amounts required to
apply force changes without sliding the seals. As is demonstrated in the simulated plant in Figure (9), the seal
stiffness has a significant impact on the overall open-loop response.
Bode Diagram
Magnitude (dB)
50
0
Nominal Seal Stiffness
High Seal Stiffness
Phase (deg)
−50
0
−45
−90
−135
−180
−225
−2
10
−1
0
10
10
1
10
Frequency (Hz)
Figure 9. The Simulink model shows the impact of the seal stiffness on the open-loop actuator response. When the seal
stiffness is increased by a factor of ten, the model predicts an oscillation at 1.3 Hz. The same results would occur if the
actuator axial stiffness were to be decreased by a factor of ten.
The authors have thus theorized that as the actuator becomes cold, the elastomeric seals become stiff and the
mechanical bandwidth of the cylinder is compromised. As the mechanical bandwidth of the cylinder decreases,
the stability margins of the control loop deteriorate and the actuator eventually oscillates. It follows that since
the bushing friction is rigidly coupled to the actuator shaft, the controller is far less tolerant of the bushing
friction than it is of the seals friction. Simply measuring the magnitude of the friction does not yield a complete
picture. Possible strategies to correct this problem can include any combination of the following:
a. Reduce the sliding friction
b. Increase the axial stiffness of the actuator and its mounting
c. Decrease the stiffness of the actuator shaft to friction coupling
d. Increase the servo lead gain to compensate for friction-induced lag
e. Decrease the overall servo gain and actuator bandwidth
Of the above strategies, all but option (b.) were explored in addressing the instability. Increasing of the
actuator’s axial stiffness was found to be impractical since the stiffness is dominated by that of the load cell.
Replacement of the 536 actuator load cells with stiffer variant would have been prohibitively expensive.
The following are highlights of the changes made to the LBT primary mirror cell in order to correct the
instability. Most of the changes are intended to minimize the effect of temperature on the support system.
Specific attention was payed to friction and seal stiffness.
6.1 Seals
Earlier mirror cells developed by Steward Observatory use rolling diaphragm cylinders, which exhibit very low
friction. Sliding-seal cylinders, such as the ones used in the LBT primary mirror cell, have friction that is ten
or more times greater than that of the rolling diaphragm cylinders and low temperature testing using a rolled
diaphragm cylinder did yield sufficient low temperature performance.
Significantly reducing the sliding seal friction was found to be impractical. Several unsuccessful attempts were
made at producing a low friction seal using different materials without compromising leakage or bandwidth. It
seemed that any low friction seal developed was also stiff and thus it failed to produced satisfactory results. In
the end, the authors concluded that if the actuators are to operate with the sliding-seal cylinders, either the
seals must remain soft or the servo gain terms must be modified to accommodate the friction.
The material used in both the shaft and piston seals has been found to be critical. For example, by changing
the seal material from Viton to NBR, the actuators can be made to operate at temperatures below -20◦ C.
Unfortunately, NBR is ozone intolerant and thus is not an acceptable replacement for the Viton seals.
Various materials have been investigated, including NBR, silicone, HNBR and EPDM, but few have provided
the combination of ozone resistance, low temperature compliance and durability required for the actuators.
Of the materials tested, EPDM offered the best combination of propertied for this application. The material
is far more compliant at low temperatures than Viton and it offers far more durability than Silicone. EPDM is
also ozone resistant. After encouraging results from testing with standard o-rings, an order for custom EPDM
seals was placed with SMC, the manufacturer of the cylinders.
6.2 Lubricants
The original lubricant used in the cylinders and the ball decouplers was Krytox, a vacuum grease. This grease
becomes very thick at temperatures below freezing. After looking at several similar alternatives, the authors
settled on Braycote 601EF for the cylinders, which also a vacuum grease. SMC, the manufacturer of the
custom seals, expressed some concern about compatibility of EPDM with the Braycote 601EF grease. Their
recommendation was to test for compatibility with Braycote 815Z oil, the base oil used in the grease. After
extensive material testing the authors concluded that the materials are compatible. The Krytox used in the ball
decouplers was replaced with Nye Rheolube 951, a low-stiction polyolester-based grease with a vapor pressure of
.1 X 10-7 torr at 75◦ C.
6.3 Bushings
The Simulink model indicated that the actuators would be particularly sensitive to the bushing friction. The
authors found that the tolerance of the bushing bore was such that the clearance may be insufficient at very low
temperatures. Removal of 25 microns from the nominal diameter of the bushings ensured low friction operation
at temperatures below -20◦ C.
6.4 Inner control loop compensation
The gains of the inner control loop were also explored. The Simulink model did suggest that by modifying the
controller gains, more bandwidth could be obtained and stability could be improved in the presence of friction.
However, the authors could not entirely compensate for the mechanical changes observed in the actuators.
Figure (10) shows the measured response of the prototype actuator with the updated gain set.
Bode Diagram
Magnitude (dB)
20
0
−20
Phase (deg)
−40
0
−50
−100
−150
−200
0
1
10
10
Frequency (Hz)
Figure 10. Increasing the lead gain helps to compensate for the effects of friction. As a side effect, the nominal actuator
bandwidth has been increased from 5 Hz to 9 Hz. The data shown here is the actual measured response of the prototype
actuator. The command amplitude is 13 N .
7. CONCLUSION
The left LBT mirror cell and dummy mirror were mounted to the telescope in January of 2004. Following
a reasonably successful testing campaign, the real mirror was integrated into the cell which was mounted to
the telescope in October of 2005. As the temperature fell, the stability of the active support system began to
deteriorate until it was rendered unusable at temperatures below about -2◦ C. After studying the instability at
cold temperatures, the authors became convinced that the source of the problem was the force actuator used to
actively support the mirror.
The as-built actuators failed to operate properly over the full specified operational temperature range. The
instability was found to occur as a result of insufficient bushing clearance and the stiffening of the piston and
shaft seals in the pneumatic cylinders. The modifications to the pneumatic cylinders defined in Section (6) are
designed to extend the operating temperature range of the actuator to below -25◦ C. A Simulink model was
used to develop a set of optimized gain making the inner loop controller more tolerant of variable friction. The
resulting actuator exceeds the room-temperature performance of the original actuators and the performance is
now essentially uniform over the tested temperature range of -25◦ C to 25◦ C.
The left LBT mirror cell was reintegrated in September of 2006 and the right mirror cell followed shortly
thereafter. At this time, the left mirror cell has now successfully operated through two winters and the right has
operated through one. No temperature-related issues have surfaced since the re-engineered support system was
installed in the two mirror cells.
Though the performance of the active mirror support system is now considered excellent, there continues to
exist room for improvement in the hardpoints. Future hardpoint modifications are expected to increase stiffness,
increase the breakaway force and improve smoothness of the motion. These modifications are expected to be
made over the course of the next two years.
Additional information about the LBT can be found at the website (http://lbto.org) and in other papers
presented at this conference: “Commissioning and early operations of the Large Binocular Telescope”,1 “The
Large Binocular Telescope”,2 “Prime focus active optics with the Large Binocular Telescope”,3 “Use of field
aberrations in the alignment of the Large Binocular Telescope”4 and “An overview of instrumentation for the
Large Binocular Telescope”.5
REFERENCES
[1] Green, R. F. and Hill, J. M., “Commissioning and early operations of the large binocular telescope,” Proc.
SPIE 7016(08) (2008).
[2] Hill, J. M. and Green, R. F., “The large binocular telescope,” Proc. SPIE 7012(02) (2008).
[3] Hill, J. M. and Ragazzoni, R., “Prime focus active optics with the large binocular telescope,” Proc.
SPIE 7012(57) (2008).
[4] Rakich, A. and Hill, J. M., “Use of field aberrations in the alignment of the large binocular telescope,” Proc.
SPIE 7012(56) (2008).
[5] Wagner, M. R., “An overview of instrumentation for the large binocular telescope,” Proc. SPIE 7014(08)
(2008).