Resonance vibration control method and system

Abstract

A bearing arrangement includes a shaft, at least one contact bearing and at least one non-contact bearing and a controller. The controller is configured to control a magnitude of a restoring force applied to the shaft by the non-contact bearing in accordance with a sensed parameter such that a stiffness of the shaft is modified such that one or more resonance frequencies of the shaft are moved away from one or more external forcing frequencies.

Claims

1. An arrangement comprising: a shaft; at least one contact bearing and at least one non-contact bearing having a variable stiffness, the shaft being supported by and radially fixed by being in physical contact with the at least one contact bearing; and a controller configured to: receive signals corresponding to at least one sensed parameter of the shaft, wherein said at least one sensed parameter includes speed of the shaft, determine a resonance detuning target frequency fin based at least on the at least one sensed parameter, calculate a restoring stiffness Kiln of the at least one non-contact bearing from $f_n=\frac{1}{2\pi }\sqrt{\frac{\stackrel{\_}{K_n+K_{nc}}}{\stackrel{\_}{M_n}}},$ where K.sub.n is a modal stiffness of the shaft and the at least one contact bearing and M.sub.n is modal mass of the shaft and the at least one contact bearing, and control a restoring force applied to the shaft by changing the variable stiffness of the at least one non-contact bearing to the restoring stiffness.

2. The arrangement according to claim 1, wherein the at least one contact bearing comprises one of a journal bearing and a rolling bearing.

3. The arrangement according to claim 2, wherein the rolling bearing comprises one or more of a roller bearing, a ball bearing, a spherical bearing and a taper bearing.

4. The arrangement according to claim 1, wherein the at least one non-contact bearing comprises a magnetic bearing.

5. The arrangement according to claim 4, wherein the controller is configured to control voltage and/or current through one or more bearing magnetic windings to control the restoring force.

6. The arrangement according to claim 1, wherein the at least one non-contact bearing comprises an air bearing.

7. The arrangement according to claim 6, wherein the controller is configured to control air pressure and/or air flow to control the restoring force.

8. The arrangement according to claim 1, wherein the at least one sensed parameter comprises one or more of a shaft displacement and a shaft rotational or vibrational frequency.

9. The arrangement according to claim 1, wherein the at least one sensed parameter comprises a vibrational displacement of the shaft.

10. The arrangement according to claim 9, wherein the controller is configured to alter the restoring force to a value that results in a minimum vibrational displacement.

11. The arrangement according to claim 1, wherein the at least one non-contact bearing includes two or more non-contact bearings that are located at different positions along the shaft.

12. The arrangement according to claim 1, wherein the at least one contact bearing includes two or more contact bearings that are located at different positions along the shaft and at least one of the at least one non-contact bearing is provided at a position between two contact bearings of the two or more contact bearings.

13. A gas turbine engine comprising the bearing arrangement in accordance with claim 1.

14. The gas turbine engine according to claim 13, wherein the engine comprises at least one compressor and at least one turbine interconnected by a main engine shaft, wherein the main engine shaft comprises the shaft of the arrangement.

15. The gas turbine engine of claim 13, wherein the resonance detuning target frequency f1n is further based on an engine operational condition.

16. A method of controlling a bearing arrangement supporting a shaft, the bearing arrangement comprising at least one contact bearing arranged to support and radially fix the shaft by being in physical contact with the shaft, and at least one non-contact bearing having a variable stiffness, the method comprising: sensing at least one parameter of the shaft, wherein said at least one sensed parameter includes speed of the shaft, determining a resonance detuning target frequency f.sub.1n based at least on the at least one sensed parameter, calculating a restoring stiffness K.sub.nc of the at least one non-contact bearing from $f_{1n}=\frac{1}{2\pi }\sqrt{\frac{\stackrel{\_}{K_n+K_{nc}}}{\stackrel{\_}{M_n}}},$ where K.sub.n is a modal stiffness of the shaft and the at least one contact bearing and M.sub.n is modal mass of the shaft and the at least one contact bearing, and controlling a restoring force applied to the shaft by changing the variable stiffness of the at least one non-contact bearing to the restoring stiffness.

Description

(1) FIG. 1 is a sectional side view of a gas turbine engine;

(2) FIG. 2 is a sectional front view of a reduction gearbox of the gas turbine engine of FIG. 1;

(3) FIG. 3 is a sectional side view of a bearing arrangement of the gas turbine engine of FIG. 1;

(4) FIG. 4 is a sectional front view of a magnetic bearing of the bearing arrangement of FIG. 3;

(5) FIG. 5 is a sectional side view of part of an alternative bearing arrangement comprising an air bearing;

(6) FIG. 6 is a flow diagram of a first method of controlling the bearing arrangement of any of FIGS. 3 to 5;

(7) FIG. 7 is a graph showing vibration relative to shaft rotational speed before stiffness is modified; and

(8) FIG. 8 is a graph showing vibration relative to shaft rotational speed after stiffness is modified.

(9) With reference to FIG. 1, a gas turbine engine is generally indicated at 10, having a principal and rotational axis 11. The engine 10 comprises, in axial flow series, an air intake 12, a propulsive fan 13, a gearbox 19, a compressor 14, combustion equipment 15, a high-pressure turbine 16, a low-pressure turbine 17 and an exhaust nozzle 18. A nacelle 20 generally surrounds the engine 10 and defines the intake 12 and a bypass passage 21.

(10) The gas turbine engine 10 works in the conventional manner so that air entering the intake 12 is accelerated by the fan 13 to produce two air flows: a first air flow into the intermediate pressure compressor 15 and a second air flow which passes through a bypass duct 22 to provide propulsive thrust. The compressor 14 compresses the air flow directed into it before delivering that air to combustion equipment 15 where it is mixed with fuel and the mixture combusted. The resultant hot combustion products then expand through, and thereby drive the high 16 and low-pressure 17 turbines before being exhausted through the nozzle 18 to provide additional propulsive thrust. The high 16 and low 17 pressure turbines drive respectively the compressor 14 and the fan 13 via the gearbox 14. A high pressure shaft 22 extends between the high pressure turbine 16 and the compressor 14, while a low pressure shaft 23 extends between the low pressure turbine 17 and an input of the gearbox 14.

(11) The gearbox 14 is a reduction gearbox in that it gears down the rate of rotation of the fan 13 by the low pressure turbine 17. The gearbox 14 is an epicyclic planetary gearbox having a static ring gear 24, rotating and orbiting planet gears 25 supported by a planet carrier 26 via respective planet pins 27, and a rotating sun gear 28. The sun gear 28 is coupled to the low pressure shaft 23, and so acts as an input to the gearbox 14, while the planet carrier 26 is coupled to a fan shaft 29, and so acts as an output from the gearbox 14.

(12) One or more of the shafts 22, 23, 29 and gearbox components, such as the planet pins 27, carrier 26 and sun gear 28, are supported by a bearing arrangement 30.

(13) Referring to FIG. 3, a bearing arrangement 30 is shown. In this embodiment, the bearing arrangement is described in relation to the low pressure shaft 23, though it will be understood that the low pressure shaft could be substituted for one or more other shafts or rotatable components of the engine, or indeed for a rotatable component of a machine other than a gas turbine engine.

(14) The arrangement 30 comprises first and second contact bearings 31, 32 which are axially spaced from one another. The contact bearings 31, 32 are in the form of cylindrical roller bearings, though it will be understood that other suitable types of contact bearings could be employed, such as ball bearings. It will also be understood that further contact bearings may be provided, for example, to react axial loads. In some instances, only a single contact bearing may be employed.

(15) Each of the bearings 31, 32 is mounted to static structure 33 of the engine 10, such as housings etc, to support the shaft 23 relative to the remainder of the engine 10, while allowing for rotation of the shaft.

(16) The bearing arrangement 30 further comprises one or more non-contact bearings. In this embodiment, three non-contact bearings are provided, which are each in the form of magnetic bearings 34, 35, 36.

(17) One of the magnetic bearings 36 is shown in more detail in FIG. 4. The magnetic bearing 36 comprises an annular stator 37, which is typically formed of a soft magnetic material, and may comprise a plurality of laminations. A plurality of salient teeth 38 project inwardly from the stator 37. Stator windings 39 are wound round each stator, and are coupled to an electrical power supply such as a generator or battery (not shown). The stator 37 and teeth 38 surround the shaft 23, which comprises a rotor lamination 40 therearound. The rotor lamination 40 rotates with the shaft 23, and comprises a magnetic material such as iron.

(18) In use, magnetic fields produced by the windings 39 in view of electrical current within the windings 39 interact with the rotor lamination 40 magnetic field to create a restoring force F between the shaft 23 and the stator 37. The force F acts to support the shaft 23 against the force of gravity and other accelerations, such that an air gap 41 is maintained in use. Consequently, the shaft 23 is free to rotate, and is at least partly supported by the non-contact bearing 36, in addition to the contact bearings 31, 32. However, in general, the restoring force of the non-contact bearing is generally relatively small in comparison to the reaction force that the contact bearings 31, 32 are capable of generating.

(19) The magnetic bearing further comprises one or more sensors 42, 43, which are configured to sense one or more sensed parameters of the bearing system 30. Optionally, these sensors 42, 43 include a shaft speed sensor 42, a shaft vibrational magnitude sensor 43, and a shaft temperature sensor. Examples of suitable speed sensors include Hall effect sensors, optical sensors (such as a phonic wheel), etc. as would be available to the skilled person. Similarly, vibrational sensors configured to sense the magnitude of the air gap 41 would be available to the skilled person.

(20) The windings 39 and one or more sensors 42, 43 are electrically coupled to a controller 44. The controller 44 is configured to control the magnitude of the restoring force F generated by the magnetic bearing 36 in accordance with parameters determined by the sensors 43 such that a stiffness of the shaft 23 is modified such that, in turn, one or more resonance frequencies f.sub.n of the shaft 23 are moved away from one or more external forcing frequencies.

(21) In use, the shaft 23 will typically be subject to vibration. This vibration is a result of driving forces, which will have one or more frequencies and harmonics. The bearing system 30 (and the engine 10 as a whole) similarly has a natural frequency and one or more harmonics, as is familiar to the skilled person. Where one or more driving frequencies match the system natural frequencies a resonance occurs. In this situation, vibrational magnitudes can increase rapidly as the driving and response frequencies more closely match. This large increase in vibration can cause engine damage, and must be prevented. One component of the driving frequencies is related to shaft rotational speeds, since shaft imbalances (either due to mass imbalance, or directional aerodynamic forces) are a common and large contributor to the vibrational input. However, it will be recognised that other components may also be present, such as aerodynamic and maneuvering loads.

(22) FIG. 7 shows a graph relating driving frequencies and vibrational magnitude for a given driving force. As can be seen, the vibrational magnitude of the shaft 23 is highly dependent on the frequency of the driving vibration, with peaks coinciding with driving frequencies of 11 and 17 Hz. Consequently, where operating conditions (shaft rotational speed, aerodynamic force etc.) are such that the driving frequency coincides with one of the peaks (i.e. a resonant frequency of the system), large vibrations will occur. These vibrational magnitudes can conventionally be reduced by damping. Damping absorbs some of the vibrational energy, and converts this into heat. However, in order to absorb large amounts of vibrational energy, large damping forces are required.

(23) Furthermore, while damping may reduce the peak vibrational magnitude, it does not in general change the natural frequency of the system—rather, damping “flattens out” the curve, such that forcing at frequencies lying adjacent the natural frequency result in increased vibrations (since the Q factor of the system is reduced). Consequently, keep-out zones may be increased by damping.

(24) The bearing arrangement 30 seeks to avoid this problem by dynamically altering shaft stiffness during operation of the engine 10, by altering non-contact bearing 34, 35, 36 restoring forces in accordance with one or more sensed parameters which are indicative of one or more driving frequency. Consequently, the present system alters the natural frequencies of the system, so that reduced damping is required.

(25) Referring to FIG. 6, the method operates as follows.

(26) The control method uses a form of model based control, with an inner feedback loop to ensure safe operation.

(27) In a first step, a predetermined residual peak vibration level J is determined, based on maximum vibrations that can be tolerated in service. This will be determined on a case-by-case basis for a given engine design type, depending on service life requirements, materials, driveline architecture etc. This may be determined by experimentation or simulation.

(28) An engine operational model is then created, to determine a vibrational response (i.e. vibrational magnitude expected at the resonance peaks) of the engine as a function of input frequency (Hz), which is in turn related to a certain engine operational condition within the engine operational envelope (e.g. a current shaft speed within the range of possible shaft speeds). One or more peaks are determined within the model, which are representative of one or more natural frequency f.sub.n. It is also known which engine operational conditions determine resonances with one or more of the natural frequencies f.sub.n. A resonance detuning target frequency Δf can then be determined for a given shaft rotational speed, which results in a residual vibration level below the predetermined level J at that speed, achieved by a predetermined natural frequency f.sub.n shift.

(29) A required non-contact load modal (theoretical) demand (i.e. a restoring force magnitude) can then be calculated for a given engine operational condition, which is calculated to move the natural frequency fn away from the engine generated driving force by the resonance detuning target frequency shift Δf, such that the residual vibration level is below the predetermined level J, as a result of having brought the engine out of the resonance.

(30) This required non-contact load modal demand is calculated on the basis of a natural frequency model, such as the following:

(31) $f_n=\frac{1}{2\pi }\sqrt{\frac{\stackrel{\_}{K_n+K_{\mathrm{nc}}}}{\stackrel{\_}{M_n}}}$
where fn is the system natural frequency, kn is the modal stiffness of the shaft 13 and the contact bearings 31,32, knc is the variable stiffness of one or more non-contact bearings at n locations (i.e. the required non-contact load modal demand), and M.sub.n is the modal mass of the shaft 13 and the contact bearings 31,32.

(32) From the above model, a required theoretical non-contact load modal demand n can be calculated such that the natural frequency f.sub.n is moved from the engine operational condition by at least the resonance detuning target frequency shift Δf.

(33) Optionally, an adaptive correction factor calculated upon the actual operation conditions may be applied. This correction factor is applied to account for engine-to-engine variation, and differences between the assumed engine speeds, clearances, temperatures and vibrations of the model.

(34) In determining the correction factor one or more parameters of the bearing arrangement are sensed as inputs. Typically, these parameters include one or more shaft speeds and vibration. Where the parameters include shaft speeds, the sensed shaft speeds typically include at least the main shaft which is supported by the bearing arrangement, which in embodiment case is the low pressure shaft 23, which is sensed by speed sensor 42. This shaft speed is frequently referred to in the art as “n1”. Further main shaft speeds n2, n3 may also be sensed. Where the engine has two main shafts, n2 represents the high pressure shaft 22 speed. Where the engine comprises three main shafts, n2 represents an intermediate pressure shaft speed, while n3 represents a high pressure shaft speed.

(35) In the present embodiment, further parameters are also sensed in order to target the detuning of the resonance conditions. The parameters include measured temperatures at one or more locations by temperature sensors.

(36) Vibrational parameters are also sensed by the vibration sensors 43. Inputs from the sensors 43 may be processed to determine measured peak vibration amplitudes (i.e. displacement amplitudes in, for example, a radial direction, from a mean position), and also one or more peak frequencies. The peak frequencies may be determined from the vibration data from a Fourier Transform analysis or similar analysis in the frequency domain, as would be understood by the skilled person.

(37) Each of these parameters is then input to determine the correction factor to superimpose to the theoretical restoring non-contact force demand to produce an actual demand. For example, where the measured vibrations are still above the predetermined limit J, the non-contact restoring force demand is adjusted to reduce the vibration.

(38) This modal load demand is then translated into a winding electrical power using a closed feedback loop. For example, the displacement 6 of the shaft 23 from the mean position is measured by the vibration/displacement sensors 43. The restoring force F provided by the non-contact bearings 34, 35, 36 is known from a look-up table or from sensitivity coefficients (i.e. a partial derivative of restoring force upon displacement, which relates electrical winding current (which is controlled by the controller 44) to restoring force F). The resultant non-contact stiffness can then be calculated from the following relation:

(39) $\partial K_{\mathrm{nc}}=\frac{\partial F}{\partial \delta }.fwdarw.\stackrel{\_}{K_{\mathrm{nc}}}\frac{\stackrel{\_}{F}}{\stackrel{\_}{\delta }}$

(40) The calculated actual is then adjusted to correspond to the demanded by a conventional control loop, such as a PID (proportional, integral, differential) control scheme.

(41) As a result of the positioning of the non-contact bearings 34, 35, 36 between two contact bearings, controlling the restoring force F will have the effect of adjusting the stiffness of the shaft 23.

(42) FIG. 8 shows a graph of vibration relative to shaft speed once the above method has been applied to adjust the natural frequency of the system.

(43) As can be seen, the natural frequency (shown by the peak in the graph) is now moved from the driving force frequency by the target frequency difference Δf.sub.n. Similarly, the magnitude of the vibration has been reduced to a level below the target frequency.

(44) It will be understood that different control schemes could be utilised to provide for control of the non-contact bearings 34, 35, 36. For instance, a fixed schedule could be defined, which relates shaft speeds to non-contact bearing restoring forces. Alternatively, shaft vibration could be measured, and the restoring force adjusted where the vibration level exceeds a predetermined level.

(45) In general however, the control system acts over a time scale of greater than one revolution, and/or acts with equal force in all radial directions. In other words, the restoring force is approximately isotropic in a radial or axial plane. In contrast, damping systems seek to provide anisotropic forces, to directly oppose movement of the shaft, by acting in an opposite direction to a displacement.

(46) The bearing arrangement 30 allows for further control, including multiple degrees of freedom to control stiffness. As will be understood by the skilled person, shaft stiffness is dependent on support radial stiffness, and shaft tension.

(47) Bearing 36 is configured to provide a controllable radial restorative force to the shaft, and thereby control support radial stiffness. This can be achieved by arranging for the magnetic flux from the stator teeth 38 to be directed inwardly toward the shaft rotor laminations, as shown in FIG. 4.

(48) On the other hand, bearings 34 and 35 are configured to provide controllable axial restorative force to the shaft in opposite directions, and thereby control shaft axial tension. This can be achieved by arranging for the magnetic flux from the stator teeth of the bearings 34, 35 to be directed axially toward a shaft rotor lamination which projects radially from the shaft 23.

(49) The disclosed arrangement and control methodology can be applied to different non-contact bearing types. FIG. 5 shows a non-contact bearing in the form of an air bearing 45, which could be substituted for one or more of the magnetic bearings 34, 35, 36.

(50) The air bearing 45 comprises a bearing body 46, which is spaced from the shaft 23, and defines a cavity 47 within. The cavity 47 is open to the shaft 23, to define an air gap. The cavity 47 is filled with pressurised air from an air source such as the compressor 14 through an air inlet 48. The pressure and/or flow rate of air is controlled by a valve 49. Air outlets 50 adjacent the shaft 23 allow air to escape. Pressure within the cavity 47 and the outlets 50 provide a restorative force, which centres the shaft 23 without any of the bearing 45 components coming into physical contact with the shaft 23.

(51) The air bearing 45 can be controlled in accordance with the control methodology of either of FIG. 6 or 7 to achieve a similar effect.

(52) Advantageously, the invention provides for control of system resonances, to reduce machine vibration in use. In practical terms, this may result in the reduction or elimination of “keep out zones”, which may result in increased operational flexibility.

(53) Other gas turbine engines to which the present disclosure may be applied may have alternative configurations. By way of example such engines may have an alternative number of interconnecting shafts (e.g. three) and/or an alternative number of compressors and/or turbines. Further the engine may comprise a gearbox provided in the drive train from a turbine to a compressor and/or fan.

(54) It will be understood that the invention is not limited to the embodiments above-described and various modifications and improvements can be made without departing from the concepts described herein. Except where mutually exclusive, any of the features may be employed separately or in combination with any other features and the disclosure extends to and includes all combinations and sub-combinations of one or more features described herein.

(55) For example, the system could comprise one or more air bearing and one or more magnetic bearing in combination. Similarly, it will be understood that the system could be applied to machines other than gas turbine engines.