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Research Papers: Gas Turbines: Structures and Dynamics

Flexure Pivot Tilting Pad Hybrid Gas Bearings: Operation With Worn Clearances and Two Load-Pad Configurations

[+] Author and Article Information
Luis San Andrés

Mechanical Engineering Department, Texas A&M University, College Station, TX 77843-3123lsanandres@mengr.tamu.edu

Keun Ryu

Mechanical Engineering Department, Texas A&M University, College Station, TX 77843-3123keun@tamu.edu

J. Eng. Gas Turbines Power 130(4), 042506 (Apr 29, 2008) (10 pages) doi:10.1115/1.2800346 History: Received May 10, 2007; Revised June 06, 2007; Published April 29, 2008

Gas film bearings enable the successful deployment of high-speed microturbomachinery. Foil bearings are in use; however, cost and lack of calibrated predictive tools prevent their widespread application. Other types of bearing configurations, simpler to manufacture and fully engineered, are favored by commercial turbomachinery manufacturers. Externally pressurized tilting pad bearings offer a sound solution for stable rotor support. This paper reports measurements of the rotordynamic response of a rigid rotor, 0.825kg and 28.6mm in diameter, supported on flexure pivot tilting pad hybrid gas bearings. The tests are performed for various imbalances, increasing supply pressures, and under load-on-pad (LOP) and load-between-pad (LBP) configurations. Presently, the initial condition of the test bearings shows sustained wear and dissimilar pad clearances after extensive testing reported earlier, (Zhu, X., and San Andrés, L., 2007, “Rotordynamic Performance of Flexure Pivot Hydrostatic Gas Bearings for Oil-Free Turbomachinery  ,” ASME J. Eng. Gas Turbines Power, 129, pp. 1020–1027). In the current measurements, there are no noticeable differences in rotor responses for both LOP and LBP configurations due to the light-weight rotor, i.e., small static load acting on each bearing. External pressurization into the bearings increases their direct stiffnesses and reduces their damping, while raising the system critical speeds with a notable reduction in modal damping ratios. The rotor supported on the worn bearings shows an 10% drop in first critical speeds and roughly similar modal damping than when tested with pristine bearings. Pressurization into the bearings leads to large times for rotor deceleration, thus demonstrating the little viscous drag typical of gas bearings. Rotor deceleration tests with manually controlled supply pressures eliminate the passage through critical speeds, thus paving a path for rotordynamic performance without large amplitude motions over extended regions of shaft speed. The rotordynamic analysis shows critical speeds and peak amplitudes of motion agreeing very well with the measurements. The synchronous rotor responses for increasing imbalances demonstrate the test system linearity. Superior stability and predictable performance of pressurized flexure pivot gas bearings can further their implementation in high performance oil-free microturbomachinery. More importantly, the measurements show the reliable performance of the worn bearings even when operating with enlarged and uneven clearances.

Copyright © 2008 by American Society of Mechanical Engineers
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Figures

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Figure 1

Schematic cross sectional view of test rig

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Figure 2

Dimensions of flexure pivot tilting pad hydrostatic bearing (units: mm)

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Figure 3

Estimated radial clearances and axial shape of pads in test bearings

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Figure 4

Amplitudes of rotor synchronous response versus speed. Baseline imbalance. 5.08bar feed pressure. Test setup for static experiments. Side view.

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Figure 5

Phase difference (left∕right) of recorded imbalance responses versus speed. Baseline imbalance, LOP configuration, 5.08bar feed pressure.

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Figure 6

Amplitude ratio (left∕right) of recorded imbalance responses versus speed. Baseline imbalance, LOP configuration, 5.08bar feed pressure.

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Figure 7

Effect of increasing supply pressure on test rotor synchronous response. Baseline imbalance.

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Figure 8

Effect of supply pressure on time extent for coast-down rotor response. Baseline imbalance.

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Figure 9

Amplitudes of rotor synchronous response versus speed. Imbalance displacement U(B2): 0.49μm (out of phase), LOP configuration, 5.08bar feed pressure.

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Figure 10

Manual changes in supply pressure as rotor speed coasted down

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Figure 11

Amplitudes of rotor synchronous response versus speed for controlled feed pressures. LOP configuration, baseline imbalance. Measurements at right bearing side, vertical direction (RV).

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Figure 12

Measured and predicted mass flow rates versus supply pressure for test bearings. LOP configuration.

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Figure 13

Structural model of test rotor

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Figure 14

Damped natural frequency map of test rotor-bearing system. 5.08bar feed pressure. LOP configuration.

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Figure 15

Predicted damping ratios for rotor-bearing system versus rotor speed. 5.08bar feed pressure. LOP configuration.

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Figure 16

Predicted and measured imbalance response for three mass imbalance conditions. Tests at 5.08bar feed pressure. LOP configuration. Displacements at left bearing, vertical direction (LV). Baseline response subtracted.

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Figure 17

Predicted and measured normalized imbalance responses. Tests at 5.08bar feed pressure. LOP configuration. Displacements at left bearing, vertical direction (LV). Baseline response subtracted.

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Figure 18

Predicted static journal eccentricity (e∕C) for increasing feed pressures. Right bearing, LOP and LBP configurations.

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Figure 19

Predicted attitude angle for increasing feed pressures. Right bearing, LOP and LBP configurations.

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Figure 20

Predicted synchronous direct bearing force coefficients versus speed. LB and RB. LOP configuration. 5.08bars (absolute) and no feed pressure.

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Figure 21

Comparison of direct bearing force coefficients for LOP and LBP configurations. Right bearing without feed pressure. Synchronous speed coefficients.

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