TECHNICAL PAPERS: Gas Turbines: Structures and Dynamics

Rotordynamic Coefficients Measurements Versus Predictions for a High-Speed Flexure-Pivot Tilting-Pad Bearing (Load-Between-Pad Configuration)

[+] Author and Article Information
Adnan M. Al-Ghasem, Dara W. Childs

 Texas A&M University, Turbomachinery Laboratory, College Station, TX 77840

J. Eng. Gas Turbines Power 128(4), 896-906 (Sep 06, 2005) (11 pages) doi:10.1115/1.2179467 History: Received August 30, 2005; Revised September 06, 2005

Experimental dynamic force coefficients are presented for a four pad flexure-pivot tilting-pad bearing in load-between-pad configuration for a range of rotor speeds and bearing unit loadings. Measured dynamic coefficients have been compared to theoretical predictions using an isothermal analysis for a bulk-flow Navier-Stokes (NS) model. Predictions from two models—the Reynolds equation and a bulk-flow NS equation models are compared to experimental, complex dynamic stiffness coefficients (direct and cross-coupled) and show the following results: (i) The real part of the direct dynamic-stiffness coefficients is strongly frequency dependent because of pad inertia, support flexibility, and the effect of fluid inertia. This frequency dependency can be accurately modeled for by adding a direct added-mass term to the conventional stiffness/damping matrix model. (ii) Both models underpredict the identified added-mass coefficient (32kg), but the bulk-flow NS equation predictions are modestly closer. (iii) The imaginary part of the direct dynamic-stiffness coefficient (leading to direct damping) is a largely linear function of excitation frequency, leading to a constant (frequency-independent) direct damping model. (iv) The real part of the cross-coupled dynamic-stiffness coefficients shows larger destabilizing forces than predicted by either model. The frequency dependency that is accounted for by the added mass coefficient is predicted by the models and arises (in the models) primarily because of the reduction in degrees of freedom from the initial 12 degrees (four pads times three degrees of freedom) to the two-rotor degrees of freedom. For the bearing and condition tested, pad and fluid inertia are secondary considerations out to running speed. The direct stiffness and damping coefficients increase with load, while increasing and decreasing with rotor speed, respectively. As expected, a small whirl frequency ratio (WFR) was found of about 0.15, and it decreases with increasing load and increases with increasing speed. The two model predictions for WFR are comparable and both underpredict the measured WFR values. Rotors supported by either conventional tilting-pad bearings or flexure-pivot tilting-pad (FPTP) bearings are customarily modeled by frequency-dependent stiffness and damping matrices, necessitating an iterative calculation for rotordynamic stability. For the bearing tested and the load conditions examined, the present results show that adding a constant mass matrix to the FPTP bearing model produces an accurate frequency-independent model that eliminates the need for iterative rotordynamic stability calculations. Different results may be obtained for conventional tilting-pad bearings (or this bearing at higher load conditions).

Copyright © 2006 by American Society of Mechanical Engineers
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Figure 1

Flexure-pivot tilting-pad bearings

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Figure 2

Main test section of the test rig

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Figure 3

Shaker-stinger configuration (NDE side)

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Figure 4

Bearing-stator assembly

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Figure 5

Experimental and theoretical dynamic stiffnesses versus the excitation frequency at 12,000rpm and 689.5kPa bearing unit loading

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Figure 6

Stiffness coefficients versus rotor speeds for different bearing unit loading: (a) 1.4kPa, (b) 175.9kPa, (c) 348.4kPa, (d) 513.8kPa, (e) 692.2kPa, and (f) 1038.2kPa

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Figure 7

Damping coefficients versus rotor speeds for different bearing unit loading: (a) 1.4kPa, (b) 175.9kPa, (c) 348.4kPa, (d) 513.8kPa, (e) 692.2kPa, and (f) 1038.2kPa

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Figure 8

Added-mass coefficients versus rotor speed for different bearing unit loading: (a) 1.4kPa, (b) 175.9kPa, (c) 348.4kPa, (d) 513.8kPa, (e) 692.2kPa, and (f) 1038.2kPa

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Figure 9

Experimental WFR versus for rotor speed for different bearing unit loading

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Figure 10

Experimental and theoretical (bulk-flow and Reynolds models) WFR versus rotor speed for different bearing unit loading, with fluid inertial effects




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