IMPROVEMENTS IN OR RELATING TO DAMPENING THE RADIAL VIBRATIONS OF A RQTOR This invention relates to dampening the radial vibrations of a rotor; by radial vibrations we mean those vibrations that occur in a plane lying substantially perpendicular to the axis of rotation of the rotor. The invention applies particularly to high speed rotors supported by at least two fixed bearings spaced apart axially; such rotors are frequently found in turbo-machinery, for instance. Such rotors are essentially flexible and are subject to many forces inherently capable of setting them into types of radial vibration in which a point on the rotor surface describes a circular or elliptical orbital motion centred on the equilibrium position of the rotor in that radial plane. This motion or precession is in the same direction as the basic rotation of the rotor. While the machinery will generally be designed so that the amptitude of such vibration is acceptable in the majority of running conditions, there is always the danger of severe vibration developing for instance within certain ranges of speed or load. This can result in a rapid increase in the magnitude of the orbit and thus of the radial amplitude of the vibration, with the obvious risk of mechanical failure. The type of potentially harmful vibration just described can arise in response to forces of many kinds - for instance those generated in the bearings by which the rotor is supported, or by the impact of fluid upon impellers and blades, or in labyrinths or other seals, or from magnetic effects among others. The present invention relates especially to the common cases in which the occurrence of the vibrations, in response to such forces, can be said to be due to either forced response or instability. Forced response is largely a matter of unbalanced distribution especially where flexible rotors are involved. In certain cases a rotor that was initially well balanced may have been running for a long time at a running speed well above a natural frequency. In turbo- machinery the original balance is often disturbed by a combination

of blade erosion and deposits from the process fluid. For large machines not fitted with a braking mechanism the run down time is considerable thus allowing a significant time within speed ranges where resonance occurs. This problem is recognised in the petro- chemical industry by specifying that vibration measurements are obtained during run-down tests. High response is due to the small damping of the natural frequencies of the rotor system. One well known method of reducing vibration response is to increase damping. However conventional damping devices rely on using the motion of the vibrating body itself to provide the damping force e.g. oil dash pots and shock absorbers. As the motion is necessarily small, viscous fluids are normally used to provide sufficient damping forces- For rotating machinery squeeze film bearings are often used particularly in aero engines. However such devices require very accurately-made hardware, tend to use viscous fluids, and are essentially dependent on the motion of the rotor for their operation.

Instability, the other mechanism mentioned above from which vibration commonly results, is a well-known problem in high speed turbo-machinery rotors and typically occurs because the speed and/or load passes beyond the stability boundary. The problem usually manifests itself as an increase in the vibration level at a non-synchronous frequency. In most cases this frequency is a natural frequency of the system which is insufficiently damped. As this frequency is often exceeded in the acceleration to running speed its reappearance as a result of instability is normally sub-synchronous i.e. at a frequency less than that corresponding to running speed. However unlike synchronous resonance it is usually impossible to pass through successfully without either reducing load or speed. Instability occurs because a lateral motion of the shaft gives rise to a force perpendicular to the displacement. The force vector is aligned with the translational velocity of the precessing shaft and thus behaves as a negative damping force. Experimental measurements of cross-coupling forces of this general nature are well established in bearings, impellers,

blade rings and seal passages. In small annular clearances, typical of labyrinth seals, the mainly axial inlet flow develops a strong circumferential component as a result of friction from the rotating shaft. Eventually the mean tangential component is equivalent to half the surface velocity of the shaft. Some manufacturers fit anti-swirl vanes at the entrance to labyrinth seals to impose a backward whirl to the inlet flow. This delays the full development of the mean circumferential half-speed swirl and so reduces any cross-coupling that may be present in the labyrinth. Also a combination of a roughened stator with a smooth rotor has been shown to reduce mean tangential velocity. However neither of these approaches essentially alter the basic nature of the circumferential flow.

The object of the present invention is to provide a damping mechanism working according to a different principle from any of the known mechanisms already described, and offering the potential advantages among others of less need for precision components, relatively easy fitting to existing machinery, no requirements for viscous fluids and relative independence of rotor speed. According to the invention a device to dampen radial vibrations of a rotor of circular section comprises a stator presenting a bore within which the rotor is located with a substantially annular clearance; motor means to rotate the rotor; a plurality of jet orifices formed in the bore, these orifices being located at equal intervals of circumference around the bore and so that they discharge into the clearance in a direction that is opposite to that in which the rotor is rotating, and that has a substantial tangential component relative to the annular clearance, and means to supply fluid under pressure to at least some of the orifices so that it discharges from them as jets into the clearance, whereby the resulting force upon the rotor tends to dampen the tendency of the rotor to lateral vibration.

The bore may be of circular section, and the direction of discharge of each jet orifice substantially tangential relative to the bore.

The surface of that part of the rotor lying within the bore may be profiled so as to enhance the damping action of the discharged fluid upon it. The surface may be knurled or otherwise roughened, or may be profiled by being formed with a plurality of blind holes or other depressions. The rotor may include an outer sleeve in which holes or other reaction-enhancing profile features are formed, so that the profiled outer surface of the sleeve becomes effectively the outer surface of the rotor.

The jet orifices may all be similarly connected in common to the fluid supply means, and sensors responsive to the speed of the rotor may control the fluid supply means, whereby the fluid supply means operate only when that speed enters certain ranges; those ranges may be those within which, resonant radial vibration of the rotor is to be expected. Alternatively individual jet orifices may be separately connected to the fluid supply means, and monitoring means may be provided for continuously monitoring the rotor for eccentricity, and may be operable to connect and disconnect the fluid supply means to only a selection of the orifices whereby to correct the instability. The monitoring means may operate whereby an eccentricity of the rotor, detected during one revolution, results in a corrective operation of the chosen selection of orifices during a succeeding revolution.

The invention also includes a method of damping the radial and precessing vibration of a rotor as it rotates with a narrow annular clearance within a stator, by directing at lest one jet of fluid substantially tangentially into the clearance in a direction opposite to that of the rotating surface of the rotor.

The invention will now be described, by way of example, with reference to the accompanying drawings in which:- Figure 1 is a diagrammatic and exploded view of .one damping device;

Figure 2 shows the nozzle ring of Figure 1 in perspective and on an enlarged scale;

Figure 3 shows parts of a stator in transverse section;

Figure 4, 5 and 6 show alternative designs for part of the rotor;

Figure 7 shows one damping device with an associated and simple control system schematically, and Figure 8 is similarly schematic, but shows a device associated with a more complex control system.

In Figure 1 a rotor 1 comprises a shaft 2 supported at axially- spaced locations by fixed bearings 3, 4 and driven in rotation - in the direction of arrow 5 - by a motor 6. A narrow band 7 of the surface of shaft 2 is roughened in a manner to be described in greater detail with reference to Figures 4 to 6, and lies within the nozzle ring 8 of a stator assembly which is shown in exploded view. The assembly comprises end plates 9 and 10 and a middle plate 11, secured together face-to-face by means of bolts (as indicated diagrammatically at 12) passing through registering holes 13 and all, like the bearings 3, 4 and the motor 6, anchored to ground 14. Ring 8 lies within the central aperture 15 of plate 11 and, as Figure 2 shows best, comprises four segments 16 separated by clearances or orifices 17 which are of wedge shape, being narrower at their mouths 18 - that is to say, their radially- innermost ends - then they are at their inlets 19 - that is to say, their radially-outermost ends. Shaft 2 passes with easy clearance through the central apertures 20, 21 of plates 9 and 10 but rests with only slight clearance within the bore of ring 8, the discontinuous wall of which is defined by the curved inner faces 23 of segments 16. Typically the radial dimension of the clearance between band 7 of the shaft and the wall of the bore is of the order of one hundredth of the radius of shaft 2 about its nominal axis 24. By a conduit indicated in outline at 25, the inlets 19 of orifices 17 are connected to a source 26 of gas under pressure by way of a control system 27 and holes 28 formed within plate 9. As will be readily apparent, pressurised air or other fluid so supplied to the inlets 19 of the orifices 17 will emerge from the mouths 18 as jets, aimed in a direction opposite to that

of the rotation 5 of shaft 2, and substantially tangential to that shaft. Such fluid may then escape from the damping device to exhaust by way of the central apertures 20, 21 in plates 9 and 10. Figure 3 is a transverse section through the ring and rotor of an alternative construction in which the ring now comprises five segments 16, the orifices 17 of which are again equally spaced around the roughened-surface band 7 of the rotor 2 which rotates in the direction of the arrow 5. The narrow and substan¬ tially annular clearance 35 between ring and rotor is now clearly shown. Segments 16 are mounted on the end plate 10 by means of studs (not shown) which pass through holes 31 in segments 16 and engage with tapped holes 32 in plate 10. The more rounded shape of the extremities of orifices 17 in Figure 3 is more true to what practice could require, and this Figure also shows the annular space 36 between the outer faces of the segments 16 and the boundary of the central aperture 15 in plate 11 which constitutes a manifold or plenum in the fluid supply path between the conduit 25 (see Figure 1) and the five orifices 17.

The appropriate location of the ring 8 upon the shaft is dependent on the predicted modal behaviour of the shaft and in many instances, if there is a reasonable amount of overhang, it may be effective to locate the ring close to a free end of the shaft, that is to say outside the two supporting bearings 3 and 4, instead of between them as shown. Figures 4 to 6 show alternative ways of providing an appro¬ priate surface to the band 7 of the shaft 2 that lies directly within the ring 8, and on which the fluid jets emerging from the mouths 18 of the orifices 17 directly bear. In Figure 4 a knurled or diamond pattern 40 is machined on the surface of shaft 2, and in Figure 5 blind holes.41 are drilled in the shaft surface.

Figure 6 shows a hollow sleeve 42 which is perforated by a mass of holes 43 and which makes an interference fit over the shaft 2 so that the outer surface 44 of the sleeve, profiled as it is by the holes 43 formed in it, forms the effective surface of band 7 of the shaft in use.

The rotor of Figure 7 is driven by motor 6 and supported by bearings 3 and 4 as before, but now essentially comprises a shaft 45 which carries a disc-like enlarged part 46 which presents the roughened surface 7. The ring 8 is not shown but is similar in principle to the rings already described with reference to Figure 1 to 3, except that it only presents two orifices 17, arranged diametrically opposite to each other. Each of the branches of conduit 25 connecting the respective orifices 17 to the fluid source 26 contains a valve 47, responsive to the control unit 27 which is in turn responsive to a speed sensor 48, 49 which monitors the speed of rotation of shaft 45. When sensor 48, 49 indicates to the unit 27 that the speed of rotation of the shaft 45 is within certain ranges, typically those close to a predictable resonant condition, the unit 27 opens valves 47 so that high pressure air from source 26 issues as jets from orifices 17. As both jets are supplied with fluid equally, the damping action which they exert relies on the precessing shaft restricting the flow of fluid from the orifice closest to the minimum radial gap between the surface 7 and the ring surrounding it. In this way a differential force is created which opposes the precession by acting against the lateral motion of the shaft, and by so opposing tends also to dampen the radial vibration with which it is associated.

In the more complex control system shown in Figure 8 several of the features of Figure 7 are repeated and are similarly referenced. Again for simplicity only two of the jet orifices 17 are shown, although in practice more orifices - for instance, as many as the five shown in Figure 3, or more - would be used. Now however the separate valves 50, 51, in the two branches of conduit 25 are controlled by a more complex control unit 52 in the form of a micro-processor which constantly receives and samples inputs indicative of the instantaneous shaft speed w(t) from the sensor 49, and of the instantaneous positions x(t) and y(t) of the shaft 45 relative to two orthogonal axes x and y, from non-contact sensing probes 53, 54 respectively. For a shaft designed to rotate at a maximum speed of say 5000-6000 rpm a sampling rate of

say 5000 samples per second for icon rtrol unit :'52 ^< may be appropriate. : From such sampling the unit continuously updates the mean position of the shaf from the mean values of the two positional inputs and the current deviation in each of the x and y directions from the mean so computed. Control unit 52,is then programmed continually to check whether the resultant of the two deviations is greater than is permissible. If so, the unit, is programmed to cause the valve (50 or 51) of the jet orifice nearest to the eccentric position represented by that resultant to open after an appropriate delay. That delay will in principle equal (nT-tc-tv) where n is a whole number, preferably unity and certainly as small as possible, T is the time required for a complete revolution of the shaft* tc is the time which unit 52 takes to make a computation, and tv is an allowance related to the speed of operation of the necessary mechanical and pneumatic components. Thus the motion of the shaft is continuously monitored and the jet closest to the instantaneous position of the shaft is activated at the correct instant in a succeeding cycle.

The purpose of the discharge of a single tangential jet (as in the system of Figure 8) or of a plurality of jets (as in previous Figure) is to obtain a high circumferential flow of the jet fluid acting, principally by means of friction, against the forward rotation of the rotor. When the rotor is concentric with the stator, the fluid friction is a pure torque proportional to the mean flow of the fluid within the even, radial clearance between rotor and stator. However, when the rotor axis is radially displaced from its nominal position, the overall effect of the friction force of the fluid upon the rotor acts at right angles to the displacement. With the direction of motion of each jet being opposed to that of the part of the surface of the rotor on which it directly impinges, the fluid/rotor friction force acts against forward precession of the rotor and thus adds damping to the related radial vibration.

One apparatus according to the invention included a rotor of the kind shown in Figure 7 and 8, comprising of a silver steel

shaft 9.5 mm in diameter and 28 cm long with a 0.846 kg disc of diameter 76.2 mm and width 25 mm mounted at mid span. The disc corresponded to item 46 in Figures 7 and 8 except that its radially- outermost surface was smooth, not roughened. Supporting the shaft at either end were brass bushes mounted in housings with a single rubber "0"-ring between each bush and Its housing, these rings themselves of course introducing a small amount of damping into the rotor system. The motor 6 was a 76W controllable DC electric motor, connected to the shaft by a flexible coupling and capable of a maximum speed of around 6000 rpm the ring 8 was formed with four orifices 17 as in Figures 1 and 2, the wedge angle of each orifice being 10 and the rectangular exit area of the mouth 18 of each orifice being 0.5 mm by 9.5 mm. Radial clearance between ring 8 and rotor 2 was 0.457 mm giving a rotor radius/radial clearance ratio of approx 83. Air at a pressure of 1.4 bar absolute was supplied to the plenum area (36, Figure 3) between the ring and plate 11, resulting in jet velocities from the orifices 17 of the order of 240 m/sec compared with rotor surface speeds up to a maximum of say 25 m/sec. Air velocity of a lower order would have sufficed had the outer surface of the disc been roughened.

When the motor of such an apparatus was run up to maximum speed from zero, with the air supply disconnected - so that the pressure inside the manifold 36 equalled ambient pressure - it was found that the amplitude of synchronous radial vibrations of the rotor rose to a clear maximum value at around 3000 rpm, thereafter falling. With the air supply connected to raise the manifold pressure to only about 1.41 times ambient, the maximum value of the vibration amplitude occurred at the same frequency value and the general shape of the amplitude/frequency curve was as before, but the magnitude of the maximum amplitude was reduced by about 40%. While the invention has been described with reference to air as the jet fluid, other fluids - including liquids - can also be used.

112B