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TECHNICAL PAPERS: Gas Turbines: Structures and Dynamics

Rotordynamic Performance of Flexure Pivot Hydrostatic Gas Bearings for Oil-Free Turbomachinery

[+] Author and Article Information
Xuehua Zhu

 SKF China Ltd., Shanghai, ChinaSusan.Zhu@skf.com

Luis San Andrés

 Texas A&M University, College Station, TX 77843-3123Lsanandres@mengr.tamu.edu

J. Eng. Gas Turbines Power 129(4), 1020-1027 (Jan 02, 2007) (8 pages) doi:10.1115/1.2720518 History: Received April 14, 2006; Revised January 02, 2007

Micro-turbomachinery demands gas bearings to ensure compactness, light weight, and extreme temperature operation. Gas bearings with large stiffness and damping, and preferably of low cost, will enable successful commercial applications. Presently, tests conducted on a small rotor supported on flexure pivot hydrostatic pad gas bearings (FPTPBs) demonstrate stable rotordynamic responses up to 100,000rpm (limit of the drive motor). Test rotor responses show the feed pressure raises the system critical speed (increase in bearing direct stiffness) while the viscous damping ratio decreases. Predictions correlate favorably with experimentally identified (synchronous) direct stiffness bearing force coefficients. Identified experimental gas bearing synchronous damping coefficients are 50% or less of the predicted magnitudes, though remaining relatively constant as the rotor speed increases. Tests without feed pressure show the rotor becomes unstable at 81krpm with a whirl frequency ratio of 20%. FPTPBs are mechanically complex and more expensive than cylindrical plain bearings. However, their enhanced stability characteristics and predictable rotordynamic performance makes them desirable for the envisioned oil-free applications in high speed micro-turbomachinery.

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Copyright © 2007 by American Society of Mechanical Engineers
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Figures

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Figure 1

Schematic cross sectional view of test rig (unit: cm)

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Figure 2

Flexure pivot hydrostatic pad gas bearing. Photograph of bearing and details of bearing geometry.

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Figure 3

Synchronous peak-peak rotor amplitudes versus rotor speed for remnant imbalance and two imbalance conditions exciting combined cylindrical/conical modes. Imbalances U1=1.45μm, U2=1.45μm. Supply pressure 3.77bar (absolute).

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Figure 4

Transmitted amplitudes of bearing load versus rotor speed for remnant imbalance and two mass imbalances exciting combined cylindrical/conical modes. Left bearing vertical force sensor. Imbalances U1=1.45μm, U2=1.45μm. Supply pressure 3.77bar (absolute).

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Figure 5

Waterfall plot of rotor amplitudes with conical imbalance excitation U2. At left bearing vertical plane (LV). Test at 3.77bar (absolute) supply pressure.

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Figure 6

Baseline synchronous rotor peak-peak amplitudes versus rotor speed for three supply pressures (absolute)

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Figure 7

Baseline transmitted bearing force amplitudes versus rotor speed for three supply pressures. Left bearing vertical load sensor.

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Figure 8

Waterfall plot of rotor amplitudes at right bearing vertical plane (RV). Speed run up test without external pressurization. Speed range 20krpm to 90krpm.

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Figure 9

Estimated experimental gas bearing direct stiffness and damping coefficients (KXX,CXX) versus rotor speed for three supply pressures. Synchronous speed force coefficients, X-vertical.

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Figure 10

Comparison of predicted and experimentally identified (2.39bar supply pressure) synchronous direct stiffness and damping force coefficients for gas bearing. Predictions do not include external pressurization.

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Figure 11

Predicted journal eccentricities (e∕C) and power loss versus rotor speed for tilting pad gas bearing without external pressurization

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Figure 12

Predicted synchronous stiffness and damping force coefficients versus rotor speed for tilting pad gas bearing without external pressurization

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Figure 13

Effect of excitation frequency on predicted gas bearing force coefficients. Rotor speed=40krpm. No external pressurization.

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