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Research Papers: Gas Turbines: Structures and Dynamics

Comparison Between Numerical and Experimental Dynamic Coefficients of a Hybrid Aerostatic Bearing

[+] Author and Article Information
Mohamed Amine Hassini

Institut Pprime,
UPR3346 CNRS,
Université de Poitiers,
Poitiers, 86962 France;
Centre National d’Études Spatiales,
CNES-DLA, Paris, 86962France
e-mail: mohamed.amine.hassini@univ-poitiers.fr

Mihai Arghir

Institut Pprime,
UPR3346 CNRS,
Université de Poitiers,
Poitiers, 86962France
e-mail: mihai.arghir@univ-poitiers.fr

Manuel Frocot

Snecma Division Moteurs Spatiaux,
Vernon, 27208France
e-mail: manuel.frocot@snecma.fr

Contributed by the International Gas Turbine Institute (IGTI) of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 20, 2012; final manuscript received July 23, 2012; published online October 11, 2012. Editor: Dilip R. Ballal.

J. Eng. Gas Turbines Power 134(12), 122506 (Oct 11, 2012) (9 pages) doi:10.1115/1.4007375 History: Received July 20, 2012; Revised July 23, 2012

Hybrid journal bearings have been considered for many years as a possible replacement for ball bearings in turbopumps used by the aerospace industry. Due to flow regimes dominated by inertia and due to the nature of the lubricant (cryogenic fluids), the prediction of the linearized dynamic coefficients in these bearings must be based on the compressible bulk-flow equations. Theoretical models based on these equations were validated for hybrid bearings working with water or for liquid or gas annular seals. Validations for hybrid compressible bearings are missing. Experimental data obtained for an air lubricated hybrid aerostatic bearing designed with shallow pockets were recently presented; the data consist of linearized dynamic coefficients obtained for rotation speeds up to 50 krpm and up to 7 bars feeding pressure. The present work introduces a consolidated numerical approach for predicting static and linearized dynamic characteristics. Theoretical predictions are based on bulk flow equations in conjunction with CFD analysis. It was found that, for a given feeding pressure, the value of the pressure downstream the orifice has a major influence on all results. Special care was then taken to describe the complex flow in the feeding system and the orifice. Three dimensional CFD was employed because the bulk-flow equations are inappropriate in this part of the bearing. The pressure downstream the orifice stemming from CFD results and the feeding pressure were next imposed in the bulk flow model and the equivalent area of the orifice was obtained from the numerical solution of the steady flow in the bearing. Since the pockets of the hybrid bearing are shallow, this equivalent area is considered as being the harmonic average of the orifice cross section area and of the cylindrical curtain area located between the orifice and the rotor. The comparisons between theoretical dynamic coefficients and experimental data validated this approach of the equivalent area of the orifice.

Copyright © 2012 by ASME
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References

Figures

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Fig. 1

Schematic view of the aerostatic hybrid bearing

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Fig. 2

Pressure distribution (Ps = 3 bars - Ω = 0 krpm)

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Fig. 3

Recess pressure in the circumferential direction (Ps = 3 bars - Ω = 50 krpm)

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Fig. 4

Sensitivity of the dynamic coefficients following a 1% variation of: the supply pressure, recess pressure, supply temperature, clearance, roughness, and axial and circumferential pressure losses. (a) Direct stiffness, (b) cross coupling stiffness, and (c) direct damping.

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Fig. 5

Flow trajectories between the supply and the recess

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Fig. 6

Pressure distribution (Ps = 7 bars - Ω = 0 rpm)

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Fig. 7

Dimensionless mass flow rate (case where Aorif/Ainh=2.38)

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Fig. 8

Dimensionless mass flow rate (case where Aorif/Ainh=0.1)

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Fig. 9

Dimensionless mass flow rate (case where Aorif/Ainh=1)

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Fig. 10

Recess pressure calculated from CFD data (Ω = 0 rpm)

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Fig. 11

Direct stiffness versus the rotational speed (PS = 3 bars)

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Fig. 12

Cross coupled stiffness versus the rotational speed (PS = 3 bars)

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Fig. 13

Direct damping coefficient versus the rotational speed (PS = 3 bars)

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Fig. 14

Direct stiffness coefficient (Ps = 6 bars)

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Fig. 15

Cross coupling stiffness (PS = 6 bars)

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Fig. 16

Direct damping coefficient (PS = 6 bars)

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Fig. 17

Mach number in the axial direction (Ω = 0 krpm)

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Fig. 18

Pressure in the axial direction (Ω = 0 krpm)

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