Vibration Isolation of Precision Machine Tools and Instruments

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Vibration Isolation of Precision Machine Tools and Instruments D. B. DeBra, Dept. of Aeronautics and Astronautics, Stanford University, Stanford, CA/USA

Abstract Successful precision engineering is the balance of robustness of the machine and how benign the environment can be made through isolation to minimize the strains caused by vibration that compromise a machine's accuracy. This paper discusses principally the process of isolating a machine from the disturbances which come from the ground, air and utilities which serve it. Principles are reviewed and the intrinsic dependence on frequency established. Requirements are discussed and the hardware realizations currently available are reviewed. A number of examples of systems with unusually demanding requirements are given to illustrate the diverse nature of the solutions in practice. Some new developing areas are identified as the various topics are presented. Key Words: vibration isolation, active isolation, 1

Introduction:

Precision machines have some dimensions which are critical to their function. When accelerated a machine strains and these functionally critical dimensions change as, for example, when ground vibrations shake a lathe and there is an undesirable change in the relative position of a work piece and the tool. The way in which the machine tool distorts under acceleration depends upon how it is constructed, how it is mounted, and the amount and direction of acceleration that it experiences. In Figure 1 a lathe is shown in a nominal configuration. Under vertical acceleration with the mounts at the end the machine distorts in a way that causes the tool to enter the work piece. However, with a horizontal acceleration the direction of the distortion may be the opposite. These effects are influenced by the design of the machine and where the mounting points are located. We will call the machine robust when the materials and design minimizes the distortion of critical dimensions for a given acceleration. The principle behind both passive and active isolation is to decouple or attenuate the ground motion andlor motions of the utilities and acoustic excitation so the machine experiences a sufficiently low level of inertial acceleration that the relative vibrations are accedable.

Fig. 2

ab

PRINCIPLE OF ISOLATION

k, permits distortion of the machine when m is accelerated. k, is added to allow the ground to move without requiring the BASE to move as much.

//////////I/,/ INERTIAL SPACE (I.S.)

propagated to the machine is related to the natural frequency of the machine on its foundation. In this case we wish to make the support natural frequency as low as possible. See Figure 3.

w x

BASE

Fig. 1

k 4

ACCELERATION

HOW VIBRATIONS MIGHT EFFECT A LATHE

We use the description precision machines and instruments in the title. In a generic sense we could mean any machine or instrument the precision of which is compromised without appropriate isolation. We'll be more specific in discussing examples. The machine is elastic. Its distortion is due to its distributed properties. We describe these in terms of modal behavior, but it is useful for understanding the principles and for the ideas in synthesis and design to think of lumped elements with discrete masses and discrete springs. The principles behind isolation and attenuation of the effect of vibration can be seen if we think of the machine as being made up of two masses, spring coupled, and mounted on another spring to its , foundations. See Figure 2. Since vibration causes inertial loading, the forces tending to distort the machine are proportional to the mass of the distorting part. Since the stiffness determines the amplitude of the response the ratio k,Jm or the natural frequency of the elastic behavior of the machine is the proper figure of merit. The higher the natural frequency the less distortion takes place for a given vibration environment. (We should mention that the bandwidth of control systems for example in controlling a carriage plays the same role when the actuator is a force device e.g. linear motor but may be stiffer if it can't be back driven e.g. a motor driven lead screw.) In a similar manner the amount of ground seismic disturbance that gets

Annals of the CIRP Vol. 41/2/1992

Fig. 3 VIBRATION TRANSMISSION is reduced by decreasing the support natural frequency. There is no attenuation below this frequency. Jacques Pettavel was kind enough to send me an example of the isolation of an entire room full of instruments at SIP showing this technique is time honored and not limited to single instruments. "At SIP. in 1942, the plant was, at that time, in Geneva. The dividing room was built in 2 parts, each having a concrete block, mass 130,000 kg resting on 32 springs placed in 8 sets of 4. The dividing machines were installed on the blocks and the natural frequency was about 2 Hz. No damping system was foreseen. The floor was supported by the surrounding ground, having, of course, no contact with the floating mass." Thus the ground motions above 2 Hz were attenuated.

71 1

Fig. 6 Photographie de la maquette de I'installation compl&te

Fig. 4 DAMPING improves transient settling. but transmits ground motion more strongly. Damping improves settling (natural). Direct damping to ground hurts isolation.

,

,

In an isolation system the mass if excited would continue to oscillate unless damping is added. However if damping is added directly between the mass and the ground the transmission of ground motion falls off with frequency instead of frequency squared as it does with a simple spring. See Figure 4. The important thing is that there be at least one path to the system which passes through just a single spring. There are several ways of doing this. The one on the right of Figure 5 was implemented at BlPM as shown in Figure 6 and Figure 7 (Carre 1966).

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quite well since most of the energy is stored in the gravity field which has no hysteresis or passive dissipation. If mechanical springs are used the stress should be low to minimize these effects. The elasticity of a spring needs to be created with a relatively small mass if the internal resonances are to be separated significantly from the isolation frequency. This implies high stress in mechanical springs and may be incompatible with linearity requirements. However with air springs the internal organ pipe resonance may occur as high as a kilohertz while the frequency of the isolated mass is typically of the order of 2 Hz. More on their implementation later. In the design of isolation systems one needs to provide an analytical basis for the sizing of the components that are used. Different models are appropriate for different portions of the design activity. In the synthesis phases relatively simple spring mass models many times are adequate to get a design started. However, extensive modal analysis may be needed to understand which modes get excited by ground motion as well as what their frequencies and mode shapes are. Some modes may not affect the critical dimension. That is. the relative motion of the tool and work piece in our lathe example may not be affected by the elastic deformation of some parts of the machine. Today finite element analysis allows us to make rather sophisticated models and include all six degrees of freedom which are typically coupled. Design for modal decoupling can be helpful in achieving effective isolation. Damping, hysteresis, and nonlinearities. to mention a few of the real world considerations, may influence both the effectiveness of the isolation,and the way in which the machine responds.

All are damped

AII have -2 slope to high frequency or more Fig 5 DAMPING WITHOUT DIRECT COUPLING TO GROUND

In Figure 7 the implementation of this technique is shown with springs mounted serially but with a plate in the middle. The plate holds one side of the damper which is connected to the isolated mass. As shown in Figure 5 there is a path from ground to the mass through a simple spring and thus one preserves the frequency squared attenuation. There is some advantage in optimizing the damping by making the spring in parallel with the damper eight times as compliant as the other but this complicates the design and may not be realizable. The implementation of springs does not always result in the ideal characteristics implied above. Real springs have mass and internal resonances (sometimes referred to as 'slinky modes' analogous to the toy) which allow ground motion to be propagated without significant attenuation at those resonant frequencies. These frequencies can be surprisingly close to the frequency of the isolated system. Furthermore, in some extremely exacting isolation requirements (for example in gravity wave antennas), the hysteresis and internal strain relief, or creaking, that occurs in many mechanical springs must be avoided. Lossiness in the spring results in an increase in the thermal noise, up-conversion of frequency content (generation of harmonics) due to nonlinearities and thus in these cases one looks for ways of applying isolation which has a "spring" which is lossless. For horizontal motion, pendulums work

712

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0

0.1

a2

0.3 m

Fig. 7 Montage anti-vibratoire

In the twenty-fifth anniversary issue of Sound and Vibration, Jan 1992, editor Mowry comments in his editorial on "Evolution" that "The introduction of low-cost microprocessors in the 80s enabled a virtual explosion of new opportunities--experimental modal analysis, computer-aided everything, monitoring of rotating machinery, and environmental stress screening. As we work our way into the 90s we can see increased application of computer workstations to the design and manufacture of products. It has been an exciting period!" It was not always thus. In the same issue George Lange reminisced about the good old days that many of us remember: "This may be difficult to believe but in 1966 everyone did not write computer programsl" The power which is available is awesome for the truly significant analyses that need to be done to check design.

Many times when a very high degree of isolation is needed at a specific frequency it is not necessary to select a very low natural frequency for the isolation system in order to allow the frequency squared attenuation to be effective. By using a series of springs and masses (viz. an isolation stack) the rate of attenuation with frequency can be increased. A spring to the first mass reduces its acceleration and thus it compresses the second spring less. Thus the attenuation can be made as high as frequency raised to the two nth power where n is the number of masses in the stack. With these extra masses come a whole series of elastic modes in the isolation system, and in general all six degrees of freedom of each mass must be considered. Ground motions at low frequencies are transmitted, even amplified somewhat. If nonlinearities exist near the isolated mass, harmonic content may in fact produce disturbances at the sensitive frequency much larger than those which are propagated directly from the outside. An example will be given later for isolators of this type with a stack of masses for a gravity wave antenna.

Experience suggests that it is difficult to work with isolation systems which have natural frequency less than 1 to 2 Hz. Without some estimate of the isolation requirements it is not clear that one can make a good choice of how to proceed. Many times the knowledge based on experience that the machine will be disturbed by vibration is enough to obtain a commercially available isolation system and install it hoping that it will be satisfactory. The isolation natural frequency should be damped to provide good dynamic performance and a friendly feel for the operator. One needs to have the structural modes damped in a machine to avoid amplification at resonance. For many small instrumentation systems a commercially available structure is used for the foundation on which to place for example optical components. These are available with good internal damping. Newport offers a honeycomb platform with substantially better damping than a granite surface plate. A body's weight requires support on the earth, but not necessarily in space. In free fall an instrument can be made the reference for a spacecraft to follow making it a drag free satellite (Space Dept. 1974). In any event the isolators active or passive can be quite modest since the forces required for low frequency positioning is very small (Garriott 1985). Isolation from other parts of the environment: 2

In order to design a vibration isolation system one must describe the requirements in terms of the amount of attenuation needed as a function of frequency. There are two parts of this problem, namely estimating ground motion and how much vibration the machine can tolerate. Below modal resonance frequencies the strain is proportional to the modal frequency squared but the relative strain of critical dimensions depends on many other factors too. For example, ground motion couples to or drives each mode by different amounts. The mode shape influences how much of its motion contributes to the strain of critical dimensions of the machine. When these factors are taken into account one can make an estimate of the effect'of the environment on the machine. This is .not usually done very precisely even for a static instrument, but it is particularly difficult for machines that change configuration continually. The ground seismic disturbance is even less well estimated. These are made up of natural seismic disturbances and those from cultural sources. One of the reasons is the surveys that are needed are difficult and time consuming and they vary widely at different locations. Thus an isolation system that is being designed for a family of machines which may be installed in many locations has no one ground motion representative of what it will experience. Therefore some nominal estimate is made and the design proceeds. Figure 8 shows how one would go about a systematic approach if one had the luxury of having a good analysis of the machine and good knowledge of the ground motion. Once established one can check relatively simple models of the isolation system to see whether adequate attenuation can be attained, for example with a frequency squared attenuation above the isolation natural frequency and an acceptably high modal natural frequency.

z = k s / m d ; zb = (k/md MSSD

4

So far, for simplicity, we have just talked about the propagation of ground motion through a suspension system to a sensitive machine. However many machines require utilities such as electricity, signal wires, air, vacuum, and sometimes hydraulics and/or cooling fluid. Each of these can provide a path for vibrations to short circuit the ground isolation system. In addition, acoustic coupling from other noise sources such as fans in air conditioners, neighboring machine tools. etc. can create disturbances on the machine which may be as significant as the ground motions. Commercially available devices through which to pass cables interrupt the vibrations which may be transmitted by them. See Figure 9 for a cable management system by Newport as an example. The principles on how you would isolate liquid transmission lines are available in, for example Viersma 1989. Acoustic isolation can be accomplished passively with an enclosure (Beranek 1954)particularly ones which are evacuated as we shall see with some examples. Active cancellation of acoustic energy will be discussed under active control. Since the acoustic energy is broadband in most cases it m?v excite modal frequencies more readily than ground motions.

Model Cable Management System

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Mounting and dimensions

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-

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Fig. 9 CABLE MANAGEMENT SYSTEM

''. \\ Fig. 8 ISOLATION REQUIREMENTS. For example for required relative vibration ,J= z-zb < 10 nm at w = 10 Hz. for k/m = 105/secZ ( 50 Hz) z, < 0.3 um

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There are significant differences in the requirements even among machine tools (e.g. see Rivin 1979A). When science experiments are included the range of attenuation required is very large. The effect of size from 100 tonne class machines to the tiny size of scanning tunneling microscopes increase the range of isolation requirements further and introduces a very wide range of lowest modal frequency.

3

Air springs and automatic levelling:

An air spring is sketched in Figure 10. The chamber of air is pressurized to support the machine. The machine has a piston which compresses the air when the ground moves up and down or the mass is disturbed and moves. The stiffness of the air spring is proportional to its pressure which must support the mass. So if it weren't for the external atmospheric pressure the natural frequency would be constant. While the frequency changes a little due to the external pressure, the insensitivity of the isolator frequency to changing load is another one of its advantages. Damping is introduced in this type of system by allowing the air in this spring to pass through some type of resistance into another chamber. The energy that is dissipated in the resistance provides damping of the natural behavior of the machine on the air spring. Since it does not introduce damping directly to the ground it preserves the frequency squared attenuation above the isolation resonance. For effective damping independent of the amplitude of the relative motion, a linear restrictor is needed and the damping air

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chamber needs to be as much as eight times the volume of the compliance chamber. To obtain linear damping one needs to keep the Reynolds number adequately low so there is laminar flow. FORCED

NATURAL

FORCED

NATURAL

84 in. DIA DIAMOND TURNING MACHINE (DTMB)

Fig. 10 AIR SPRINGS Higher velocities produce turbulence and thus nonlinearity and in the case of an orifice the square root dependence on pressure differential makes the damping completely ineffective at very large or very small amplitudes. Laminar flow can be achieved through porous media, capillaries. or the two dimensional equivalent of- a capillary, namely parallel plates with small separation the order of 100 micrometers. An example is shown on Figure 11 where flow across the boundary between two chambers must pass radially through a series of passageways the thickness of which are established by washers between a stack of plates. (DeBra 1982,4) Commercial air springs are available from, for example, Barry, Pneumatics Systems and Newport. The piston and air chamber are closed with a roll seal which provides very little dissipation for vertical isolation. However the horizontal stiffness of air cushions of this type are higher and relatively undamped. Typically for horizontal isolation the unit is mounted on a pendulum with separate damping. Air bags are commercially available from organizations like Firestone. The thicker material increases the role the rubber material plays in the damping transmission and lateral stiffness of these devices. However, they are convenient' and less expensive. An external tank and laminar flow damper can be added.

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Fig 12 DTM3 HEIGHT CONTROL A screw on the foundatdon can be adjusted to set the nornlnal helght of each of three valves

Dampers can be designed in many configurations. See DeBra 1984. In the installation shown a rectangular configuration for the laminar flow restrictor was chosen because of available fabrication equipment. The valves are designed like a bleeding regulator and have no dead band. When the damping is properly optimized the feedback for levelling can be made quite high allowing an automatic levelling system to respond to changing loads on a machine with speeds that are comparable to the isolation natural frequency. See Figure 13.

A

01

Fig. 11 A LAMINAR FLOW RESTRICTOR Kinematic design suggests three isolation devices should be used. However many times more are needed because commercially available devices are limited in size and for large machines (greater than 100 tonnes e.g. DTM 3. Bryan 1979) ten or eleven of the largest size may be required. Figure 12 shows Don Carter adjusting a valve for automatic height control. This additional feature allows the pneumatic systems to provide an automatic levelling capability as well as the vibration isolation. Adequate damping of the pneumatic spring is required when working with an automatic levelling system or the amount of gain used stably in feedback becomes quite limited. Even though more than three isolators may be used only three valves are needed if height, pitch, and roll are to be controlled. That is, each valve may feed make up air to more than one of the air cushions.

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Fig. 13 MPERIMENTAL RESPONSE OF DTM3 8ASE WITH LAMINAR-FLOW RESTRICTORS AND HIGH-GAIN LEVELING.

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Polymer Isolators:

Many manufacturers make polymer isolators. The spring element and damping are one and the same-namely a polymer material. Companies like EAR make very small spring dampers. One finds them, for example, in magnetic hard disc isolation systems. Lord & Barry and others offer a wide range of isolators as catalogue items. Furthermore matting is available of rubber, hair, and other materials which provide good damping. Elastic materials like Viton are widely available as " 0rings and show good damping at room temperature for custom designs. Though polymer isolators may be somewhat harder to evaluate they may provide adequate isolation and at their lower cost are many times

worth a try before going to a pneumatic isolator. In some applications where temperatures are not controlled the damping of the polymer may undergo significant changes. The glassy transition temperature is the temperature at which the largest hysteretic losses can be obtained. The transition from stiff to soh takes place in a range of as little as ten Celsius. The hysteresis is high only in this temperature range where it may reach 20%. For most precision machines, however, the operating environment temperature is controlled far more closely than ten Celsius. Thus polymers may be an appropriate material for both commercially available isolators and for special designs.

to maintain it at the proper height. Since a six degree of freedom system is being supported six of these actuators are necessary and can be introduced in a variety of ways. Figure 16 shows one method of using all six to provide forces in a vertical direction and hence contribute to the support of the weight. By skewing them in pairs away from their symmetrical position the magnetic force has a horizontal component. In a matched pair with the same current level these provide just a vertical force as the two horizontal components cancel. However if the current in one is increased and the other is reduced the vertical force remains the same and a horizontal force can be generated. By using three pairs of these as shown in Figure 17 all six degrees of freedom can be controlled.

=3 Need a reference to damp against Carriage - no room MOORE MEASURING MACHINE Base - too much relative motion (LLNL cl982) Add an inertia - you need enough1

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Lateral Force-Displacement Relation carriage

mirror

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Lateral Displacement (mm)

/.,

1

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Fig. 16 LATERAL STIFFNESS AND LATERAL ACTUATION

carriage mirror damper Fig. 14 LASER INTERFEROMETER MIRROR DAMPER Take as an example a laser interferometer mirror that was retrofitted to a Moore measuring machine. See Figure 14. Because of space considerations the mirror was cantilevered vertically off the carriage but no damping was included. The output of the interferometer could not be read because of the vibration of the mirror which caused the laser interferometer to scan back and forth through its last two or three significant figures at too high a rate to be observed. Since there wasn't room to introduce any damping between the mirror and the carriage an inertial reference was added. A separate mass comparable to the mass of the mirror was mounted with a potting of urethane. The plastic used was chosen by trial and error since it is frequently difficult to get practical advice. Sheet material is available and could be cut and glued to perform the same function. In this case the "isolation" spring was already in place. It was the cantilever supporting the mirror. The problem was the natural behavior of this suspension rather than its broad band transmission characteristics. Gene Rivin sent me an early paper reviewing mountings (Rivin 1965) with the comment "not much has changed". This excellent review covers rubber-metalbonded mountings and pads and carpets of polymers, felt, cork and combinations of materials as well as springs of metal and air. 5

Active Suspension:

By active suspension we mean the use of feedback to influence the support of a machine tool. The entire tool or instrument may be supported by an actuator for example by magnetic or electrostatic forces which may be applied alone or in parallel with a passive isolation system. Inertial Space

00 Actuator pair I

Sensor for vertical motion Sensor for horizontal motion

.. Fig. 17 CONFIGURATION OF SIX ACTUATORS AND SENSORS Vibration isolation in space applications, for example of a telescope, has led to similar configurations, but here typically the magnetic actuators are differential. Since there is no bias force to support the weight one creates the bias with two solenoids on either side of the armature which is attached to the isolated telescope. By arranging six of these to give appropriate forces and torques a variety of isolation configurations have been developed (Hibble 1988). Some instruments have been electrostatically supported on the ground and in space (e.9. Everitt. 1979, and Van Patten 1982). At least one commercially available general purpose magnetic system exists: Magfloat by Hopewell 8. Co.. Hudson, N.H. In a fully active isolation one can adjust the natural frequency through electronics and, to a degree, can shape the frequency response of the vibration transmission arbitrarily. One must exercise care, however, since above the control bandwidth the magnetic field intensity is influenced by changes in the gap and a "spring" coupling will exist. In order to avoid this problem one can use a Hall effect detector and servo the magnetic flux to remain constant independent of gap changes to a very high bandwidth. The force depends only on flux and not the gap. Thus, the desired flux is commanded by the low frequency height control loop, but the flux resists change due to gap changes from seismic motion of the ground. A pitfall that needs to be watched is that the Hall effect transducer noise creates fluctuations in the magnetic flux which then become disturbing forces on the suspended mass. This occurs over the whole band width of the flux loop.

A problem with active isolation systems is that a failure essentially puts the apparatus in hard contact with the,ground. When the system fails it f i l s in such a way there is no isolation remaining.

Fig. 15 MAGNETIC ACTUATOR TO SUPPORT MASS, m Take the case in which, for example, magnetic forces are used to levitate the machine. (Hedrick 1975),(Jayawant 1982).(Lee 1988). A schematic is shown in Figure 15. If the magnetic structure is biased with a permanent magnet then the feedback control need only provide the current needed to modify the magnetic support force. Figure 15 shows a mass levitated. Typically measurements are made of the location of the mass with respect to ground. The errors with respect to its proper nominal height are fed back to the electromagnet as a control system

A more interesting approach offered by, for example, Newport and Barry, is to modify a passive isolation system by using feedback around the passive isolator and modifying its characteristics actively. Since the seismic disturbances cause accelerations in the machine the preferred feedback is to use an accelerometer for the error signal feeding it back to a force actuation system. Now in the event of an active control system failure the passive isolation system serves as a backup. The system is fail-operate even though with degraded performance. While it is not vibration isolation per se. table levelling is a part of the combined suspension required for some precision machines. For example in the pneumatic isolators an automatic levelling system is included to keep the machine level as mass moves for example along

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a carriageway. A very precise table levelling system was developed for attenuatingthe effect of low frequency fluctuations in a foundation. Even isolated foundations experience tilting motions during the day of the order of an arc second. Foundations near an outside wall are influenced by ground water and/or seasonal temperature variations and may experience tilting of tens of arcseconds. For the small diurnal variations a two actuator servo system was developed for a sensitive experiment in which an air cushion vehicle needed to be kept at a constant angle with respect to gravity. Bubble levels were used and fed back to a solonoid with a built in spring. This system provided isolation above 60 Hz but principally attenuated the tilting motions which were measured by the bubble levels at frequencies below 50 mHz.(Van Patten 1968) 6

Isolation Stack ,

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Active Control:

6.1 Acoustic Isolation Active acoustic isolation has been discussed for many years and is now becoming a commercial reality. (Chaplin 1983). Chaplin gave a number of examples including cancellation within ventilation ducts. engine noise, noise from propellers in machines, etc. Companies now offer machines commercially. For example Activox and Digisonics which use a sensing microphone called the error microphone which is then used to drive a loudspeaker to cancel the acoustic radiation incident on the sensing device. See Figure 18. Even at its current state of the art free field reductions by a factor of two or three in amplitude are claimed. Error Microphone

P Quiet Space

Microphone Fig. 18 OPEN FIELD ACOUSTIC CANCELLATION

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6.2 Structural Damping Feedback can be applied to damping the motion of a structure (Cowley 1969). Wdh discrete force inputs it has been referred to as modal suppression by Fuller and Silcox who point out that 'Yuture work will center on extending these disturbanoes in broadfrequency bands, more complex structures, and improved modelling. It is known that optimizing the position of transducers is as important as increasing the number of control s0urces.U Crawley (1988) has been a leader in this field and his laboratory results are promising. peizoelectric materials can provide a good mechanical impedance match to structural materials so the active damping does not need to compromise the structural stiffness above the feedback bandwidth.

6.3 Isolation We have seen an example of automatic leveling with the pneumatic isolators. Frolov (1990) provides the theory and by his examples some insight into the use of hydraulic as well as pneumatic actuators for controlling with a combination of height and accelerometer feedback. Magnetic and electrostatic support just discussed are further examples. With electronic signal processing it is possible to maintain uniform natural frequencies so the benefits discussed by Rivin (19798) can be realized. Examples: Gravity Wave Antenna An aluminum bar of approximately five tons is shown in Figure 19 (Michelson 1987). (Similar requirements exist for laser interferometer gravity antennas.) (Saulson 1984). This bar gets strained if a gravity wave passes. It is referred to as a gravity wave antenna. Gravity waves are exceedingly feeble and have never been detected but their existence is predicted by theory. It is important in the foundations of gravitation to verify their existence. lt is expected that several gravity waves a year would be detected ifthe strain sensitivity of such a bar could be made lower than 10". Clearly much larger strains of the aluminum bar would occur continually due to vibrations and other disturbances if it were not isolated from ground motion, acoustics. and the propagation of any other vibration through the wires used to measure its strain. The longitudinal natural frequency of the bar is 850 Hz. This is the mode that is excited by the gravity waves and hence it is the frequency requiring isolation. By using stacks of masses and springs very high attenuation rates as a function of frequency can be achieved. In addition an air mount is added at the first interface between the ground and the suspension system. The bar is located in an evacuated chamber to prevent acoustic coupling. See Figure 19. 7 7.1

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Fig. 19 GRAVITY WAVE ANTENNA (Michelson 1987) Multi-level isolation stacks provide a high rate of attenuation above 100 Hz

In order for the bar to have adequate sensitivity it must be cooled to minimize thermal noise. The current bar operates at 4 K and a modification presently under way will bring that temperature to 40 mK. The two orders of magnitude increase in sensitivity possible with this lower temperature has required a reexaminationof the vibration isolation to increase the attenuation from 200 d6 to 240 d6. Mcloughlin (1990) developed an approach to choosing locationsfor damping these modes of the multi-mass spring system .with an eddy current damper. The isolation stack design by Tom Aldcroft very cleverly shapes the spring between each of the masses so that the natural frequencies in translation and rotation for all six degrees of freedom are approximately the same. See Figure 20. The frequencies are at around a hundred Hz. Thus, for example, rotational frequencies can't propagate down the stack and then through coupling create translational disturbances of the bar. Mcloughlin found eddy current damping is the only effective and linear damper available at these low temperatures. McLoughlin designed and built such a damper and optimized the location for where one or more of these dampers would be located in the vibration stack to be most effective.

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Fig. 20 ISOUTION-STACK SPRINGS (Mcloughlin 1990) (a) Spring concept (b) Interfacing joins H-shaped masses alternating 90 deg in orientation and providing approximatley equal natural frequencies for all six degrees of freedom. He used a performance index which included the strain of the springs ( < % l o 5 of yield strength to avoid nonlinear up-conversion), heat dissipation (< 10 pW and the further away from the bar the better) and the deflection of the bar faces (< l o 8 m most of which comes from low frequencies) with appropriate weights based on the required performance. His simulation results (requiring nearly 500 states) show (Figure 21) an isolation of 272 dB at the resonant frequency of the new bar at 1100 Hz.

7.2 The Space Test of The Equivalence Principle (STEP) In STEP (Worden 1991) two coaxial proof masses are suspended magnetically by magnetic repulsion of a superconductor (Worden 1982). The suspension is used both in space and for ground testing. To provide the isolation from ground motion in all six degrees of freedom a combined isolation scheme was developed. This was a combination of the equivalent of a pendulum, namely a spherical air bearing of very large radius ("30m) for horizontal isolation, and the later addition of air cushions. In combining the technologies it was not properly anticipated

passive system of 1 Hz in both the horizontal and vertical directions and a second active vibration isolation system that will operate within the vacuum shell of the machine to attenuatethe low frequency components of the external excitation CI. 10 Hz) and attenuate the forced vibrations that are induced by the carriage motions. Two stages of acoustic isolation are an air-tight environmental acoustic shell and the UHV vacuum chamber. CORE STRUCTURE TEMPERATURE ROL SHELL

frcqucncy (Hz)

ACTIVE SOLATION

VACUUM

I

SHE

1.0

0.5

frequency (Hz)

Fig. 21 VIBRATION TRANSMISSION (McLoughlin 1990) Isolation is acceptable at critical antenna longitudinal mode even with 1/4 and 4 times the optimal value for a 2DOF damper near the top of the stack. (Note le21 scale range)

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Fig. 22 OVERVIEW OF VIBRATION ISOLATION FOR

MOLECULAR MEASURING MACHINE 8

that there are two possible ways of doing this. First, by using a pendulum on which to mount the air bags and then on top of those the experiment. Or, as was done, mounting the air bearing (or pendulum) on top of the air bags. As a historical note to insure such an oversight is not repeated one should be careful about such combinations and consider the possibility of instability. Because the air bags are so soft slight compression in one makes the air bearing tilt and the effective pivot point is moved laterally a significant amount. As the experiment moves to the side it causes the air cushion to compress further. It is not practical to increase the separation of the pads enough or to make an automatic levelling system with a sufficient bandwidth to overcome this instability. Hence the appropriate approach would be to make the pendulum support the air cushions and the experiment as opposed to stacking the suspension in the opposite order. 7.3 Diamond Turning Machines (DTM 3) This large diamond turning machine shown in Figure 12 weighs in excess of 100 tons. It required 10 of the largest air springs avaiiable. Clustered into groups of 3-5 isolators, each group is controlled by one height adjusting valve. The vertical natural frequency is 1.5 Hz critically damped. The horizontal motion has a natural frequency of 2.5 Hz and before the slideways were mounted had a Q of 40. The most significant damping comes from the mass center of the machine being above the plane of the isolators. Horizontal motion causes rocking in pitch and roll which couples into vertical compression of the isolators which is well damped. As the slideways were added, their mass of 10-15 tons each moved the mass center higher increasing the coupling and therefore the damping of the horizontal natural behavior. (DeBra 1982). Riiin, (1979b), has pointed out that when the sensitive direction on a machine is horizontal and since vertical ground vibration is frequently about twice the horizontal amplitude coupling can increase the disturbance to the sensitive direction. Thus for isolators with isotropic damping one may wish to minimize coupling but for air springs the coupling is an important way to achieve horizontal damping. We attempted to do a survey of the vibration on the isolated base. Only the largest of the commercially available seismometers have enough sensitivity to provide meaningfuldata. Weighing the order of 10 kg each it is clear these are suited only for large machines. Instrumentation problems prevented the completion of this work in the time available. To obtain unambiguous data on the attenuation obtained with the isolation system one should measure all six degrees of freedom of the ground and machine motion simultaneously. These data are usually not available so one settles for single measurements as an indication of performance rather than a quantitative measure of the vibration transmission. Molecular Measuring Machine (MMM) Clayton Teague has been developing the MMM (Teague 1989) which should provide nanometer accuracy in a work space 50 mm square. Since natural frequencies scale linearly with the reciprocal of size, this relatively large machinecompared with other instruments using scanning tunneling microscopes requires special attention to isolation and the proper choice of structural materials (high specific stiffness) The spherical geometry of the design has been developed to minimize the distortion for a given acceleration without compromisingthermal design. The approach is shown in Figure 22. The environmental isolation system provides temperature control and acoustic and vibration isolation. Two stages of vibration isolation are employed : a commercial 7.4

.....................................

Conclusions

There is a rich variety of technology for vibration isolation. Conventional mounts have changed only slowly in recent years with some modest improvements due to'new materials. The principles are clear but not easily applied in the very complex dynamic environment in which they are expected to operate. Fortunately the tools are available today to provide simultaneous evaluation of many modes and to support an experimental program of isolation development with adequate analysis. However, simple models still serve an important purpose in providing the creative synthesis aspects and preliminary sizing for isolation systems. Active techniques for isolation, modal suppression, and attenuation of free field acoustic waves are increasing in popularity as their effectiveness increases with improvement in technology and experience. 9 Acknowledgements With great appreciation for your time and effort. my thanks to ClRP colleagues Jim Bryan, Jacques Pettavel, Gene Rivin, Clayton Teague, and Eric Thwaite for your help and inputs to this review and to many others who kindly replied to my requests for manufacturer's data, references, etc. The research was partially supported by the NSF through Grant No. DDM-8914232. REFERENCES Beranek. L. L. 1954, Acoustics, Chapt. 11, "Noise Control", McGrawHill, New York. NY. 1954 Bryan, J.B., 1979, "Design and Construction of an Ultra Precision 84 Inch Diamond Turning Machine", Precision Engineering, vol.1, no. 1, Jan 1979 Cannon, R.H.. Jr., 1967, Dvnamics of Phvsical Svstems, McGraw-Hill, 1967 Carre. P.. 1966, "Installation et utilisation du comparateur phdoelectrique et interferentiel", Metrologia, vol. 2, no. 1, 1966 Chaplin. 8.. 1983, "Anti-noise - the Essex Breakthrough". CME January, 1983 Cowley, A., Boyle. A., 1969, "Active Dampers for Machine Tools", Annals of the C.I. R. P., Vol. XVlll No. 2, pp. 213-222, Geneva, Sept 210, 1969 Crawley. E. F.. et al., 1988, "Development of Piezoelectric Technology for Applications in Control of Intelligent Structures", Proceedings of the American Control Conference, pp 1890-1896, June 15-17, 1988 Crede, C. E.. 1951. Vibration and Shock Isolation, John Wiley & Sons, Inc., New York, NY, 1952 Damping '91, San Diego, CA, Proceedings 1991, Sponsored by Wright Laboratory, FDL. AFSC, Feb 13-15. 1991 DeBra. D.B.. 1984. "Design of Laminar Flow Restrictors for Damping Pneumatic Vibration Isolators", ClRP Annals, Vol. 33/1, 1984

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