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Research Papers: Gas Turbines: Structures and Dynamics

Development of a High Speed Gas Bearing Test Rig to Measure Rotordynamic Force Coefficients

[+] Author and Article Information
J. Jeffrey Moore, Andrew Lerche, Timothy Allison, David L. Ransom

 Southwest Research Institute® , P.O. Box 28510, San Antonio, TX 78228-0510

Daniel Lubell

 Capstone Turbine Corporation, 21211 Nordhoff Street, Chatsworth, CA 91311

J. Eng. Gas Turbines Power 133(10), 102504 (May 06, 2011) (9 pages) doi:10.1115/1.4002865 History: Received May 10, 2010; Revised August 09, 2010; Published May 06, 2011; Online May 06, 2011

The use of gas bearings has increased over the past several decades to include microturbines, air cycle machines, and hermetically sealed compressors and turbines. Gas bearings have many advantages over traditional bearings, such as rolling element or oil lubricated fluid film bearings, including longer life, ability to use the process fluid, no contamination of the process with lubricants, accommodating high shaft speeds, and operation over a wide range of temperatures. Unlike fluid film bearings that utilize oil, gas lubricated bearings generate very little damping from the gas itself. Therefore, successful bearing designs such as foil bearings utilize damping features on the bearing to improve the damping generated. Similar to oil bearings, gas bearing designers strive to develop gas bearings with good rotordynamic stability. Gas bearings are challenging to design, requiring a fully coupled thermo-elastic, hydrodynamic analysis including complex nonlinear mechanisms such as Coulomb friction. There is a surprisingly low amount of rotordynamic force coefficient measurement in the literature despite the need to verify the model predictions and the stability of the bearing. This paper describes the development and testing of a 60,000 rpm gas bearing test rig and presents measured stiffness and damping coefficients for a 57 mm foil type bearing. The design of the rig overcomes many challenges in making this measurement by developing a patented, high-frequency, high-amplitude shaker system, resulting in excitation over most of the subsynchronous range.

Copyright © 2011 by American Society of Mechanical Engineers
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References

Figures

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Figure 1

Cross section of test rig

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Figure 2

Photo of assembled test rig

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Figure 3

Drawing of a typical cantilever beam type foil bearing

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Figure 4

Photo of test bearing

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Figure 5

Vertical bending mode at 602 Hz

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Figure 6

Housing horizontal bending mode at 943 Hz

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Figure 7

Housing vertical bending mode at 1123 Hz

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Figure 8

Frequency response plot-drive end proximity probe mounting location

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Figure 9

Frequency response plot-nondrive end proximity probe mounting location

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Figure 10

Graphical representation of mass-elastic model

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Figure 11

First damped mode shape

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Figure 12

Second damped mode shape

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Figure 13

Third damped mode shape

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Figure 14

Fourth damped mode shape

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Figure 15

Test shaft nonsynchronous response near nondrive end proximity probe

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Figure 16

Test shaft nonsynchronous response near drive end proximity probe

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Figure 17

Bearing housing free body diagram

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Figure 18

Sample radial vibration during asynchronous excitation, outer shaft speed=60,000 rpm

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Figure 19

Linear curve-fit of impedance data in X-direction

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Figure 20

Linear curve-fit of impedance data in Y-direction

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