Research Papers: Hydrodynamic Lubrication

Optimized Design of a Large Reversible Thrust Bearing

[+] Author and Article Information
Chris M. Ettles

Bearing Sciences, Inc.,
202 Van Wies Point Road,
Glenmont, NY 12077
e-mail: ettlec@rpi.edu

Guillermo D. López

IMPSA Hydro,
Godoy Cruz,
Mendoza M5503AHY, Argentina
e-mail: guillermo.lopez@impsa.com

Hugo Borgna

IMPSA Hydro,
Godoy Cruz,
Mendoza M5503AHY, Argentina
e-mail: hugo.r.borgna@power.alstom.com

1Present address: Alstom, Birr, 5242, Switzerland.

Contributed by the Tribology Division of ASME for publication in the JOURNAL OF TRIBOLOGY. Manuscript received May 13, 2015; final manuscript received February 16, 2016; published online June 15, 2016. Assoc. Editor: George K. Nikas.

J. Tribol 138(4), 041701 (Jun 15, 2016) (8 pages) Paper No: TRIB-15-1155; doi: 10.1115/1.4032824 History: Received May 13, 2015; Revised February 16, 2016

The thrust bearing duty in a pump-turbine generator can be quite arduous, since the pad support system must be symmetrical about the center of the pad, yet the oil-film must converge adequately for either direction of rotation. Special care must be taken with large machines since the thermal and elastic deformation of the pads will increase nonlinearly with size B of the pad, for example, as B2 when thermal deformation is considered. However from first principles, the thickness of the oil film will increase with only the square root of size B½. Poorly shaped films can develop when a design standard is scaled-up to larger sizes. Three options for the thrust bearing design of a particular pump-turbine were considered: (a) “semihard” supports for the pads such as a spring-disk insert, (b) “piston-type” supports in the back of the pads, which are machined to form shallow pistons that fit into recesses, allowing the pads to be supported hydrostatically, and (c) a symmetric arrangement of coil springs. In this instance, an upper limit of thrust bearing temperature was specified. Penalties would incur if this were exceeded. It is shown using a design code (GENMAT) that the best performance is achieved with a spring support (option c), arranged to give a convex film shape in the direction of sliding, and a slightly concave film in the radial direction. This is achieved by limiting the extent of the spring pack in the circumferential direction so that there are unsupported “overhangs” at the lead and trail edges. The radial concavity is arranged by having the spring pack extend edge-to-edge in the radial direction. The bearing has performed very well since commissioning. The original machining patterns are untouched after thousands of reversals under load. The pads appear as new.

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Fig. 1

The classical converging wedge of extent B, width L, showing example isobars for pressure and flow qs escaping from the sides

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Fig. 2

The effect of wedge convergence on the efficiency W* of a converging wedge, for various aspect ratios L/B

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Fig. 3

Hydrostatically supported pads

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Fig. 4

Top: area of the pad that is hydrostatically supported. Bottom: the effect of excessive thermal crowning of the pad and rotor, which occur in the opposite sense and produce an inefficient pressure profile, compared to the generally parabolic form shown dotted, which is preferred.

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Fig. 5

The transient thermal deflection of a cantilever subject to a step increase of temperature on one face, with the back face maintained at the original temperature, from Ref. [7]

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Fig. 6

General form of the elastic and thermal deformation for a disk-supported pad

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Fig. 7

Contours of temperature, pressure, and film thickness for the close-to-optimum disk-supported bearing for this application. The film shape is good within the disk area but poor outside it.

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Fig. 8

(a) Definition of the crowning ratio β = δ/hmin where δ = center-to-edge deflection (the crown) hmin = minimum film along the section considered. (b) The successive removal of columns of springs at the lead and trail edges.

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Fig. 9

Contours for the optimum case in Table 1, for 45% overhang (22.5% at lead edge and 22.5% at trail edge)

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Fig. 10

The development of a turbulent boundary layer in the groove between pads. The cooler layer above the capture streamline a–a′ will reverse and cause a secondary boundary layer on the pad. This has a cooling effect.

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Fig. 11

Spring distribution in the final design allowing a substantial overhang at the lead and trail edges. The springs are in close hexagonal packing with a gap at the mean radius. This gives an increased concavity in the radial direction.

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Fig. 12

Contours for the spring assembly in Fig. 11, with and without the jacking ring at 3.98 MPa. The actual operating load and calculated maximum load are lower. (Note: the step change in film thickness in the recess is not included in the film thickness isobars.)



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