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TECHNICAL PAPERS

Some Effects of Start-Up and Shut-Down on Thrust Bearing Assemblies in Hydro-Generators

[+] Author and Article Information
C. M. Ettles

Tribology Group, Department of Mechanical, Aerospace and Nuclear Engineering, Rensselaer Polytechnic Institute, Troy, NY 12180-3590e-mail: ettlec@rpi.edu

J. Seyler

Voith Siemens Hydro Power Generation, Inc., 2185 North Sheridan Way, Mississauga, Ontario L5K 1A4, Canadae-mail: Jan.Seyler@vs-hydro.ca

M. Bottenschein

Voith Siemens Hydro Power Generation, GmbH & Co. KG, Alexanderstrasse 11, 89522 Heidenheim, Germanye-mail: Michael.Bottenschein@vs-Hydro.com

J. Tribol 125(4), 824-832 (Sep 25, 2003) (9 pages) doi:10.1115/1.1576428 History: Received March 12, 2002; Revised July 30, 2002; Online September 25, 2003
Copyright © 2003 by ASME
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References

Pistner,  C. A., 1966, “Some Effects of Start-Up Transient Loads on Shoe Bearings for Large Hydraulic Pump Turbines,” STLE Tribol. Trans., 39, pp. 93–98.
Glavatskikh, S. G., 2000, “Transient Thermal Effects in a Pivoted Pad Thrust Bearing,” Thinning Films & Tribological Interfaces, Proceedings of Leeds-Lyon Symposium 1999, Elsevier Science, pp. 229–240.
Laffoon,  C. M., Baudry,  R. A., and Heller,  P. R., 1947, “Performance of Vertical Water Wheel Thrust Bearings During the Starting Period,” Trans. ASME, 69, pp. 372–379.
Gustafson,  R. E., 1967, “Behavior of a Pivoted-Pad Thrust Bearing During Start-Up,” ASME J. Lubr. Technol., 89, pp. 134–142.
Baudry,  R. A., Kuhn,  E. C., and Cooper,  G. D., 1959, “Performance of Large-Waterwheel-Generator Pivoted-Pad Thrust Bearing Determined by Tests Under Normal Operating Conditions,” Trans AIEE, 45, pp. 1300–1315.
Baudry,  R. A., Kuhn,  E. C., and Wise,  W. W., 1958, “Influence of Load and Thermal Distortion on the Design of Large Thrust Bearings,” Trans. ASME, 80, pp. 807–818.
Raimondi,  A. A., 1960, “The Influence of Longitudinal and Transverse Profile on the Load Capacity of Pivoted Pad Bearings,” ASLE Trans., 3, pp. 265–276.
Ettles,  C. M., 1991, “Some Factors Affecting the Design of Spring Supported Thrust Bearings in Hydroelectric Generators,” ASME J. Tribol., 113, pp. 626–632.
Ettles, C. M., 1987. “Three Dimensional Computation of Thrust Bearings,” Proceedings of Leeds-Lyon Symposium, Elsevier, pp. 95–104.
Kawaike,  K., Okano,  K., and Furukawa,  Y., 1978, “Performance of a large Thrust Bearing with Minimized Thermal Distortion,” ASLE Trans., 22, pp. 125–134.
Ettles,  C. M., 1982, “Transient Thermo Elastic Effects in Fluid Film Bearings,” Wear, 79, pp. 53–71.
Zerbe, Glenn, 1987, “Transient Heating of a Cantilever Beam,” unpublished report, Rensselaer Polytechnic Institute, Troy, NY.
Chambers,  W. S., and Mikula,  A. M., 1987, “Operational Data for a Large Vertical Thrust Bearing in a Pumped Storage Application,” ASLE Trans., 31(1), pp. 61–65.
Ettles,  C. M., 1991, “Some Factors Affecting the Design of Spring Supported Thrust Bearings in Hydroelectric Generators,” ASME J. Tribol., 113, pp. 626–632.
Ettles,  C. M., 1991, “Three-Dimensional Thermo elastic Solutions of Thrust Bearings Using Code Marmac 1,” ASME J. Tribol., 113, pp. 405–412.
Bendarek,  K., January1991, “Oil Pressure for the Hydrostatic Lubrication of a Hydro Generator Thrust Bearing With Different Diameters of Lubrication Pockets in the Tilting Pads,” Lubr. Eng., pp. 17–20.

Figures

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Wiped pad from pump-turbine applications: (a) at Bath County hydro station, from Pistner 1; and (b) illustration of wiped pad at Sir Adam Beck power station, Ontario. This photo is supplied courtesy of George Staniewski, Ontario Power Generation
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Schematic of damage to pad in Fig. 1(b)
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Effect of crowning on load capacity for centrally pivoted pads, with L/B=1.00, from Raimondi, 7, showing load capacity “a” for a plane pad with optimum pivot position and possible points of progression “b,c” through a ratchetting process
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Examples of measured temperature or deformation during starts of two pump-turbines: (a) Temperature difference across the pad faces for hot and cold starts, B=320 mm,L=353 mm, thickness H=97 mm; (b) Radial deformation during a start, from Kawaike et al. 10B=245 mm,L=425 mm, thickness H=240 mm.
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The transient thermal deflection of a steel cantilever subject to a step increase of temperature on one face, with the back face maintained at the original temperature, from 11
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(a) Schematic of arrangement used to thermal deflection model ((a) cantilever of selected material; (b) clamped end; (c) hot water, with flow started at time zero; (d) styrofoam insulation; (e) noncontacting transducer; (f ) elastomer caulking); and (b) Correlation 12 of the cantilever model, extended to allow for a convection coefficient G on the heated face. The left scale is the nondimensional peak deflection and the right scale is nondimensional time. The points are experimental data and the curves are results from the model.
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(a,b) Model result from 11, showing changes in temperature, deflection and minimum film thickness following a step increase in load from 4.64 MPa to 9.28 MPa. The pad is 120 mm long and 30 mm thick, speed is 30 m/s. The arrows show known asymptotic values. The material values used are given in an appendix to this paper.
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Changes in temperature when the load on the bearing in Fig. 7(a,b) is halved to 1.4 MPa
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First case study: Film thickness profiles on the mean radius for a disk-supported pad of size 500 mm×500 mm. Load 3 MPa. Upper diagram, without jacking. Lower diagram, with jacking.
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First case study: Isobars of pad face temperature (°C), pressure (MPa) and film thickness (μm) for the bearing in Fig. 9 at 500 rpm, when loaded to 3 MPa: (a,b,c) With jacking shut off; and (d,e,f ) With jacking at 40 lits/min. The dashed circle in (f ) indicates the position of the support disk.
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Second case study: (a) Computed minimum film thickness during a run-down from operational speed for exceptionally thick pads of size 470 mm×470 mm. Filled symbols: standard (fast) run-down, with and without jacking. Open circles: slow run-down, with jacking; and (b) Computed recess pressures (Location (a) common start point; Location (b) end point for slow run down with jacking; and Location (c) end point for standard fast run-down with jacking)
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(a) Schematic of a basic oil lift system (1. Thrust pad, 2. Pad check valve, 3. Pad orifice, 4. Flexible line, 5. Manifold, 6. Oil bath, 7. Suction line isolation valve, 8. Pressure line isolation valve, 9. Suction line, 10. Pressure line, 11. Pressure line check valve, 12. High-pressure filter, 13. Pressure gauge, 14. Pressure relief valve, 15. Pressure relief line, 16. Hydraulic pump with A.C. motor, 17. Electrical control panel, 18. Electrical connection); and (b) Schematic of an advanced oil lift system (Additional items: 19. Pressure gauge, 20. Orifice or flow meter, 21. Pressure switch alarm for low system pressure, 22. Pressure switch to indicate normal system pressure, 23. Pressure difference switch, to indicate that flow is obtained, 24. High pressure switch alarm-indicator that relief valve setting has been reached, 25. Suction line filter, 26. Vacuum gauge on suction line).

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