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Research Papers: Gas Turbines: Structures and Dynamics

Design and Performance Prediction of Hybrid Air Foil Thrust Bearings

[+] Author and Article Information
Donghyun Lee

Mechanical and Aerospace Engineering, University of Texas at Arlington, 500 West 1st Street, Arlington, TX 76019

Daejong Kim

Mechanical and Aerospace Engineering, University of Texas at Arlington, 500 West 1st Street, Arlington, TX 76019daejongkim@uta.edu

J. Eng. Gas Turbines Power 133(4), 042501 (Nov 18, 2010) (13 pages) doi:10.1115/1.4002249 History: Received November 04, 2009; Revised June 21, 2010; Published November 18, 2010; Online November 18, 2010

Air foil bearings (AFBs) have been recognized as the most promising for oil-free turbomachinery. However, the applications of AFBs to the relatively large turbomachinery have many technical challenges due to limited load capacity and wear during start/stops. A hybrid air foil bearing (HAFB), which combines the benefits of AFB and hydrostatic air bearing, was introduced earlier by the authors, and the experimental studies showed much larger load capacity at low speeds and much lesser friction torque during start/stop than hydrodynamic counterpart. The benefit of HAFB was recognized through the experimental studies, and the concept of hybrid operation was further developed to thrust air foil bearings. This paper presents novel design features of the hybrid air foil thrust bearing (HAFTB) with radially arranged bump foils and preformed Rayleigh step contour, and presents simulated static and dynamic characteristics of the HAFTB. A 2D thin plate equation in cylindrical coordinate was solved with the finite difference method for the prediction of the top foil deflection. Parametric studies were performed to evaluate the effect of various design parameters on the static and dynamic performances of HAFTB. At low speeds, a design with orifice located at the center of land region showed the highest load capacity, while a design with orifice located near the leading edge of land region showed the highest load capacity at high speeds. Direct and coupled bearing coefficients were also calculated for various operating conditions. The direct stiffness increases with supply pressure but the direct damping decreases with supply pressure. In addition, typical hardening effect of gas film accompanying increase of stiffness and decrease of damping was predicted in high frequency excitations.

Copyright © 2011 by American Society of Mechanical Engineers
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Figures

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Figure 1

Schematic description of HAFTB: (a) overview of the radially arranged bump foil and (b) details of the bump foil; region with arrow indicates partially widened bump foil slot to accommodate orifice tubes

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Figure 2

Novel design feature of radially arranged bump foils

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Figure 3

Thrust plate with recess and radial slots

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Figure 4

Bump foils sitting on recessed area on thrust plate

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Figure 5

Coordinate system and variables describing the pad motions

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Figure 6

Details around the leading edge of top foil with spot weld

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Figure 7

Considered orifice locations βorifice=11 deg, 33 deg, 38.5 deg, and 44 deg for cases 1–4

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Figure 8

Grid scheme for 2D plate model of top foil (n,m)=(42,36): (a) unloaded thrust bearing, (b) loaded thrust bearing, and (c) top foil deflection of loaded thrust bearing using 2D plate model

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Figure 9

Pressure distributions and top foil deflection of HAFTB with orifice location at 33 deg from the leading edge (Ps=4, Γs=0.83 at 40,000 rpm): (a) thrust disk eccentricity and (b) dimensionless minimum film thickness

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Figure 10

Thrust disk eccentricities and dimensionless minimum film thickness for various orifice locations (Ps=4, Γs=0.83): (a) case 1 βorifice=11 deg, (b) case 2 βorifice=33 deg, (c) case 3 βorifice=38.5 deg, and (d) case 4 βorifice=44 deg

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Figure 11

Pressure distributions along θ and thrust disk eccentricities for various orifice locations: (a) stiffness coefficients related with translational motion and (b) stiffness coefficients related with rotational motion

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Figure 12

Stiffness coefficients versus rotating speed (Ps=4, Γs=0.83): (a) damping coefficients related with translational motion and (b) damping coefficients related with rotational motion

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Figure 13

Damping coefficients versus rotating speed (Ps=4, Γs=0.83): (a) direct translational stiffness coefficient kZZ, (b) direct rotational stiffness coefficient kξξ, and (c) cross-coupled rotational stiffness coefficient kψξ

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Figure 14

Stiffness coefficients versus rotating speed for different supply pressures (Γs=0.83): (a) direct translational damping coefficient cZZ and (b) direct rotational damping coefficient cξξ

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Figure 15

Damping coefficients versus rotating speed for different supply pressures (Γs=0.83)

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Figure 16

Thrust disk eccentricities versus rotating speed for different supply pressures (Γs=0.83): (a) direct translational stiffness coefficient kZZ, (b) direct rotational stiffness coefficient kξξ, and (c) cross-coupled rotational stiffness coefficient kψξ

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Figure 17

Stiffness coefficients versus feed parameter for different supply pressures at 50,000 rpm: (a) direct translational damping coefficient cZZ and (b) direct rotational damping coefficient cξξ

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Figure 18

Damping coefficients versus feed parameter for different supply pressures at 50,000 rpm: (a) direct translational stiffness coefficient kZZ, (b) rotational stiffness coefficient kξξ, and (c) cross-coupled rotational stiffness coefficient kψξ

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Figure 19

Stiffness coefficients versus excitation frequency ratio for different supply pressures at 50,000 rpm (Γs=0.83): (a) direct translational damping coefficient (cZZ) and (b) rotational damping coefficient cξξ

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Figure 20

Damping coefficients versus feed parameter for different supply pressures at 50,000 rpm (Γs=0.83)

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