Ciro Del Vecchio, Luciano Miglietta
Osservatorio Astrofisico di Arcetri
Largo Enrico Fermi 5, 50125 Firenze, ITALY
email: ciro@arcetri.astro.it and migliett@arcetri.astro.it
and
W. Davison
University of Arizona, Large Binocular Telescope Project
Steward Observatory, Tucson, AZ 85721
email: wdavison@as.arizona.edu
Proceedings of SPIE conference on Advanced Technology Optical Telescopes V, 2199, p. 773, (1994)
Abstract
1. Introduction
2. Swing Arms
3. Primary Mirror Cells
4. Results of Finite Element Analysis
5. Acknowledgements
6. References
Abstract
We describe the solutions adopted for the most important mechanical
subsystems of the Large Binocular Telescope (LBT, formerly Columbus
Project), which is now in the phase of detailed design. We report in
particular the design and the results of static and dynamic finite
element analysis of the open telescope elevation structure, of the
azimuth structure, of the cantilever swing arms supporting the
auxiliary optics, and of the primary mirror cells.
1. Introduction
The Large Binocular Telescope Project is described in general terms
by Hill and Salinari3. The paper
explores the mechanical aspects of the telescope design in greater
detail. The goals of this mechanical design were to provide a stiff
support for the optical elements of the telescope in order to reduce
vibrations and misalignments; to achieve a high resonant frequency
of the structure to allow a wide bandwidth on the drives and to control
vibrations; and to produce this high performance structure without
polluting the thermal environment of the telescope with excessive
structural mass.
The present telescope design (Fig.1) maintains the initial approach, that has shown to provide better global performances than a number of alternative structural concepts considered in the past (see Davison 1and Salinari 6 ). In this approach the elevation structure is supported by two large rolling sectors of 14 m diameter, which rotate on four radial hydraulic supports and are laterally constrained by two pairs of hydraulic supports acting on the edges of each sector. The azimuth platform is a simple frame connecting the elevation supports to the four vertical azimuth supports acting on a rail with 14 m external diameter. The azimuth platform is radially constrained by a central bearing of relatively small diameter. The main mechanical advantages of this geometry derive from the very direct transmission of the loads from the elevation platform to the azimuth rail and in the large diameters available for the drives and for the encoders.
The primary mirror cells and the swing arms supporting secondary and tertiary mirrors are directly connected to the vertical rolling sectors, the stiffest parts of the elevation structure. As discussed in Sec.2, this improves the performances of the spiders and leaves the structure completely open, allowing the access to all telescope parts by an overhead crane. Compared with previous closed structures, the open elevation structure also reduces the cross section for the wind and the thermal mass of the telescope.
Simulations of the drives have shown that pinion and gear drives or friction drives acting on 14 m diameter can provide the required tracking performance. The pinion and gear solution, with four pinions directly driven by torque motors for both azimuth and elevation axis, was selected for the high rigidity, for its reproducibility and for the simple implementation of the safety equipment. The simple and stiff transmission, in consequence of the large reduction ratio (50--70), reduces byonly about one Hz the resonant frequencies of a telescope structure with ~10 Hz lowest frequency. Commercially available strip encoders can provide the required resolution (0.01 arcsec) on 7 m radius.
2. Swing Arms
The basic mechanical requirement for the spiders is to guarantee the
stability of the supported optical elements within optical tolerances
in the frequency domain between ~ 0.01 and 10 Hz, i.e., in the
region where the wind can cause problems. Displacements at lower
frequency, due to gravity and to thermal deformations of the spider
itself and of the entire structure, can be sensed and corrected by the
Active Optics system, that takes care of centering, collimation, and
tilt errors of the primary--secondary mirrors system.
The adopted performance specification is that the lowest resonant frequency should be above 25 Hz for all the spiders. In terms of displacement this corresponds, for instance, to a total deflection moving from zenith to horizon > 0.36 mm, or to a deflection < 3.6 µ m for a mirror unit with 0.5 ton mass, 1m2 section subject to a wind pressure change of 50 Pa. Pressure changes as large as 50 Pa on short time scales should be avoidable under most observing conditions by appropriate use of the wind--shields.
As discussed by Del Vecchio et. al. , the main problem we met in a previous more conventional design using four vanes spiders was that of achieving a sufficient rigidity of the structure to provide adequate support to the spiders in all directions. In particular, this was true for the back wall of the telescope, whose only purpose was that of supporting on that side the spiders and the rails for their displacement. It became evident that, rather than supporting the spiders, the back wall was instead depending on the spiders for its stability in the front--back direction. The first telescope resonant frequencies were relatively low (~ 8 Hz for the front--back mode) and we could not obtain a first resonance above ~ 12 Hz for the spiders. The four vanes spiders themselves, on the other hand, could be designed to achieve high rigidity (>30 Hz on fixed constraints) and reasonably low obstruction.
We found that removing altogether the back wall and halving the spiders, so that they were attached only to the front telescope wall, was of great benefit for both, telescope and spiders. This because one could increase the thickness of the front wall, with a great gain in its rigidity and no increase of the telescope mass, while the halved spider could preserve the original rigidity and mass and could take advantage of the increased stiffness of the front wall.
Figures 2 and 3 show the mechanical design of the F/15 spider and of the mechanism used to rotate the unit and to pre--load it when in working or in its rest position. The total spider obstruction is ~ 2.0% at the combined focus and 1.07% at the Gregorian.
A single mechanism is used to rotate the swing arms to their working and parking positions and to pre--load all the interfaces between the rotating and fixed parts. The applied pre--load, necessary to achieve the required reproducibility of position and rigidity of the articulation, is larger than twice the maximum load induced by gravity at each interface when the spider is in the working position, while is comparable with the gravity load in the parking position. A complete prototype of the F/15 swing arm with its rotation mechanism is inconstruction for checking its practical mechanical performance versus FEA results.
3. Primary Mirror Cells
The current drawing of the two 8.4 m honeycomb mirror cells has been
developed taking into account first of all the very stiff structure of
the telescope on which the mirror cells are directly connected. In
order to satisfy the requirement of working as a vacuum shell, during
the aluminization, the mirror cells have to be stiff enough to allow
the best use of the cell frames as structural elements of the
telescope. As a related result, also the mass of the mirror cell
increases. The mirror cells are basically supported externally by two
H beams directly from the main structure on the C rolling rings plane;
other two beams connect the front and the back box--type structure.
As shown in Fig.4. the more external H beam is then supported
by struts bracing on its inferior straight--arc directly to the C ring
elements. Inside the mirror cells, the main beams are then connected
to each other by transverse stiffeners in order to realize a rigid
frame to support the force actuators grid plane, the ventilation
system devices described by Miglietta4 and the space
reference system of the mirror realized by three couples of
positioning actuators. Beside the support of the primary mirror and
the Cassegrain instrumentation, the structure of the cell has to meet
some additional requirements. They have to sustain the instrument
rotators, they have to provide accurate thermal control of the
borosilicate honeycomb mirrors, they have to make up the bottom part
of the vacuum shell whose top part is the aluminizing bell--jar, and
finally they have to provide maintenance access to the mirror support
mechanisms. Some of these requirements may be in contrast to each
other but the final mirror cell drawing has to match the best
technical agreement as deeper discuss by Miglietta et.al.5.
4. Results of Finite Element Analysis
Tab.2 reports the displacements of six nodes at important positions in the telescope modeled in 1992. All data are in mm and are absolute displacements, i.e., the difference between the node positions before and after the application of the gravity load. Node 1528 is used for both the Gregorian rotator and the primary vertex because the relative displacement is negligible for these two positions. Typical displacements moving from zenith to horizon are in the range 0.1 to 0.4 mm and are of course mainly in the Z direction.
| Subsystem | Masses [kg] |
|---|---|
| Vertical rolling sectors | 109,460 |
| Primary cells (complete) | 98,460 |
| Horizontal frame | 15,260 |
| Vertical frame | 55,100 |
| Instruments | 11,636 |
| Spiders (complete) | 15,590 |
| Beams under cells | 23,800 |
| Total Elevation structure | 329,205 |
| Azimuth platform | 106,485 |
| Total telescope | 435,690 |
| Center of Gravity of elevation structure (zenith pointing) | [mm] |
| X,Y,Z | 0,238,385 |
| Vibration mode | Frequency [Hz] |
| Lateral bending | 8.48 |
| C mode (symmetric) | 10.68 |
| C mode (anti--symmetric) | 12.08 |
| elev. angle | displacement | |||||
| X dir | Y dir | Z dir | X dir | Y dir | Z dir | |
| node # 1510 (C. of G. of M2/F15) | node # 1506 (C. of G. of M2/F5) | |||||
| 85° | 0.047 | -0.231 | -0.393 | 0.089 | -0.251 | -0.374 | 55° | -0.096 | 0.242 | -0.437 | 0.106 | 0.290 | -0.342 | 25° | -0.165 | 0.387 | -0.528 | 0.133 | 0.503 | -0.578 | 5° | -0.181 | -0.403 | -0.571 | 0.136 | 0.473 | -0.721 |
| node # 1498 (C. of G. of M3) |
node # 1528 (C. of G. of DG rotator) | |||||
| 85° | -0.035 | -0.045 | -0.388 | -0.017 | 0.007 | -0.524 | 55° | 0.020 | 0.082 | -0.375 | -0.122 | -0.114 | -0.392 | 25° | 0.065 | 0.081 | -0.313 | -0.049 | -0.263 | -0.234 | 5° | 0.079 | 0.100 | -0.275 | 0.015 | 0.215 | -0.180 |
| node # 1534 (C. of G. of BG rotator) | node # 1529 (C. of G. of FC rotator) | 85° | 0.035 | -0.093 | -0.204 | 0.004 | -0.004 | -0.050 | 55° | -0.070 | 0.011 | -0.441 | 0.003 | -0.040 | -0.115 | 25° | -0.182 | -0.079 | -0.403 | 0.001 | 0.117 | -0.206 | 5° | -0.235 | -0.037 | -0.304 | 0.000 | 0.144 | -0.328 |
| Spider Mass [kg] |
Lowest Res. Freq. [Hz] | |
|---|---|---|
| F/15 | 932 | 29.8 (27.52) |
| M3 | 694 | 29.2 (27.82) |
| F/4 | 4237 | 19.9 (18.33) |
5. Acknowledgements
We are indebted with the engineering company ADS
from Lecco, Italy, one of the most important contractor companies
of the LBT Project Office, for the ideas, concepts, and
opinions described in this paper. We would like to thank in particular
W. Gallieni, who produced all the drawings, and R. Gatti
and R. Pozzi, who helped us for most of the FEA
computations.
