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Research Papers: Hydrodynamic Lubrication

Study of the Influence of Heat Convection Coefficient on Predicted Performance of a Large Tilting-Pad Thrust Bearing

[+] Author and Article Information
M. Wodtke

Faculty of Mechanical Engineering,
Gdańsk University of Technology,
80-233 Gdańsk Narutowicza 11/12, Poland
e-mail: mwodtke@pg.gda.pl

M. Fillon

Institut Pprime,
CNRS-Université de Poitiers-ENSMA,
UPR 3346, Dépt. Génie Mécanique et Systèmes Complexes SP2MI,
Bd Marie et Pierre Curie, BP 30179 F86962 Futuroscope Chasseneuil
Cedex, France
e-mail: michel.fillon@univ-poitiers.fr

A. Schubert

ALSTOM Hydro (Switzerland) Ltd,
Zentralstrasse 40, 5242 Birr, Switzerland
e-mail: andreas.schubert@power.alstom.com

M. Wasilczuk

Faculty of Mechanical Engineering,
Gdańsk University of Technology,
80-233 Gdańsk, Narutowicza 11/12, Poland
e-mail: mwasilcz@pg.gda.pl

Contributed by the Tribology Division of ASME for publication in the JOURNAL OF TRIBOLOGY. Manuscript received July 30, 2012; final manuscript received November 11, 2012; published online December 26, 2012. Assoc. Editor: Daniel Nélias.

J. Tribol 135(2), 021702 (Dec 26, 2012) (11 pages) Paper No: TRIB-12-1123; doi: 10.1115/1.4023086 History: Received July 30, 2012; Revised November 11, 2012

Part of the heat generated by the shearing of the lubricating film during operation of a hydrodynamic bearing is transferred to the bearing components. In the case of the pad, which is usually fully submerged in the lubricating oil, heat is further transferred at the pad free walls to the oil by convection. This mechanism causes a thermal gradient in a pad and, consequently, its thermal deflection. In large hydrodynamic thrust bearings, thermal deflection of the pads is an important phenomenon influencing bearing performance. For such bearings, pad distortion can reach the level of hydrodynamic film thickness and can significantly change the bearing's properties. In this paper, the study of the influence of the heat convection coefficient on the predicted performance of a large hydrodynamic thrust bearing is presented. Two sets of convection coefficients at the pad free surfaces are investigated with the use of thermo-elasto-hydrodynamic (TEHD) calculations. An analysis is carried out for the Itaipu hydro turbine thrust bearing with the outer diameter equal to 5.2 m, which is one of the biggest hydro power plants in the world. The results of the theoretical predictions are compared to the measured data collected during bearing operation.

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References

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Figures

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Fig. 1

Itaipu thrust bearing system with main dimensions

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Fig. 2

Sensors arrangement in the Itaipu thrust bearing: (a) temperature measurement points placed in the thrust bearing pad, and (b) runner sensors: temperature measurement points (Tr), oil gap sensors (hr), and pressure sensors (pr) positions.

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Fig. 3

Pad boundary conditions: support arrangement and assumed heat transfer coefficient distribution at the free pad surfaces for the “hcvar” case of thermal boundary conditions

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Fig. 4

FEM model of the Itaipu thrust bearing pad used for the evaluation of the pad sliding surface thermo-elastic deflection: (a) support boundary condition, and (b) pressure profile (MPa) applied to the pad sliding surface

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Fig. 5

Heat convection coefficient hc (W/m2K), calculated with the use of the FSI technique for the Itaipu pad (including support)

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Fig. 6

Thermal boundary condition applied to the FEM model of the Itaipu thrust bearing pad. (a) Example of the temperature (°C) profile at the pad sliding surface, and (b) the heat convection coefficient distribution (W/m2K) applied on the inlet, inner radius, and bottom pad walls in the “hcvar” case of the TEHD analysis.

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Fig. 7

Calculated temperatures (°C) for the pad sliding surface and pad bottom and the temperature gradient across the pad thickness (°C)

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Fig. 8

Calculated axial deflection of the pad sliding surface (μm), gap thickness (μm), and the oil film pressure (MPa)

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Fig. 9

Comparison of the measured and calculated pad temperatures for the tangential (a) (R = Rin, Rmean, and Rout; see Fig. 2), and radial (b) (θ = 0.5θo, 0.75θo, and θo) cross sections

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Fig. 10

(a) Comparison of the measured and calculated pressures, and (b) oil gap profiles for three pressure (pr1 to pr3) and distance (hr1 to hr3) radial sensor locations

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