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Research Papers: Gas Turbines: Structures and Dynamics

An Active Auxiliary Bearing Control Strategy to Reduce the Onset of Asynchronous Periodic Contact Modes in Rotor/Magnetic Bearing Systems

[+] Author and Article Information
Iain S. Cade1

 University of Bath, Bath BA2 7AY, UKi.s.cade@bath.ac.uk

M. Necip Sahinkaya

 University of Bath, Bath BA2 7AY, UKm.n.sahinkaya@bath.ac.uk

Clifford R. Burrows

 University of Bath, Bath BA2 7AY, UKc.r.burrows@bath.ac.uk

Patrick S. Keogh

 University of Bath, Bath BA2 7AY, UKp.s.keogh@bath.ac.uk

1

Corresponding author.

J. Eng. Gas Turbines Power 132(5), 052502 (Mar 03, 2010) (9 pages) doi:10.1115/1.3204644 History: Received March 24, 2009; Revised March 26, 2009; Published March 03, 2010

To prevent rotor/stator contact in a rotor/magnetic bearing system, auxiliary bearings may be located along the shaft and at the magnetic bearings. Rotor responses after a contact event may include periodic trapped modes where repeated contact may lead to highly localized thermal stresses. This paper considers an active auxiliary bearing system with a control strategy designed to limit the trapped contact modes in a rotor/magnetic bearing system that are induced by rotor unbalance. The controller is evaluated from a system model and its responses to short duration contact events. An active auxiliary bearing model is introduced to the system where the dynamic response of the bearing is dependent on the controller. From a harmonic decomposition of rotor/bearing contact, dynamic controllers are sought, which limit the numbers of possible periodic solutions for a given rotor unbalance and operating speed. A case study is performed considering a simple two degree of freedom system with passive and active auxiliary bearings. Recovery of a rotor trapped in an asynchronous contact mode is shown with variation of the auxiliary bearing controller parameters.

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Copyright © 2010 by American Society of Mechanical Engineers
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Figures

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Figure 1

Contact mode maps for different auxiliary bearing controller configurations. The same effective stiffness, kb=1 MN/m, is used for all cases. Effective damping values are (a) 10 N s/m, (b) 50 N s/m, (c) 100 N s/m, and (d) 500 N s/m.

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Figure 2

Contact mode phase angle in the rotating reference frame for different unbalance conditions

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Figure 3

(a) Number of contact modes, M, in the range 0<ωτ<1000 rad/s for different effective damping terms. Effective stiffness kb=1 MN/m. (b) Number of contact modes, M, in the range of 0<ωτ<1000 rad/s for different effective stiffness terms. Effective damping cb=10 N s/m. (c) Minimum contact modes (white) in the range of 0<ωτ<1000 rad/s for different effective damping terms. Effective stiffness kb=1 MN/m. (d) Minimum contact modes (white) in the range of 0<ωτ<1000 rad/s for different effective stiffness terms. Effective damping cb=10 N s/m.

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Figure 4

Rotor orbit within normalized clearance circle: (a) fixed frame and (b) synchronous rotating frame. The dashed line shows the clearance circle.

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Figure 5

Rotor vibration frequency content

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Figure 6

Rotor orbit within normalized clearance circle: (a) fixed frame and (b) synchronous rotating frame. After 0.02 s, the controller was changed to induce forward whirl. The dashed line shows the clearance circle.

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Figure 7

Rotor/auxiliary bearing contact force variation. After 0.02 s, the controller was changed to induce forward whirl (unshaded region).

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Figure 8

Rotor displacement in (a) a fixed and (b) a synchronous rotating reference frame with the case 1 auxiliary bearing controller. (c) and (d) show the displacement in fixed and synchronous rotating reference frames with the case 2 auxiliary bearing controller. (e) and (f) show the displacement in fixed and synchronous rotating reference frames with the case 3 auxiliary bearing controller. The dashed line shows the clearance circle.

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Figure 9

Rotor/auxiliary bearing contact forces variation under case 1–3 controllers

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