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Research Papers: Gas Turbines: Structures and Dynamics

On the Predicted Performance of Oil Lubricated Thrust Collars in Integrally Geared Compressors

[+] Author and Article Information
Luis San Andrés

Mast-Childs Chair Professor
Fellow ASME
Department of Mechanical Engineering,
Texas A&M University,
College Station, TX 77843
e-mail: Lsanandres@tamu.edu

Travis A. Cable

Department of Mechanical Engineering,
Texas A&M University,
College Station, TX 77843
e-mail: cable.travis@tamu.edu

Karl Wygant, Andron Morton

Samsung Techwin,
11757 Katy Fwy #110,
Houston, TX 77079

Derivation of Eq. (2) starts from first principles and implements the kinematics of the bull gear and thrust collar surfaces to define the film thickness material derivative. Laminar flow and an incompressible fluid are valid assumptions for this type of mechanical element.

φ* and W* determined from current industrial practice.

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 21, 2014; final manuscript received August 7, 2014; published online November 18, 2014. Editor: David Wisler.

J. Eng. Gas Turbines Power 137(5), 052502 (May 01, 2015) (9 pages) Paper No: GTP-14-1418; doi: 10.1115/1.4028663 History: Received July 21, 2014; Revised August 07, 2014; Online November 18, 2014

Integrally geared compressors (IGCs) comprise single stage impellers installed on the ends of pinion shafts, all driven by a main bull gear (BG) and shaft system. When compared to single shaft multistage centrifugal compressors, the benefits of IGCs include better thermal efficiency, reduced footprint and simple foundation, dispensing with a high speed coupling, as well as better access for maintenance and overhauls. In IGCs, the compression of the process gas induces axial loads on the pinion shafts that are transmitted via thrust collars (TCs) to the main drive shaft and balanced by a single thrust bearing. The TCs, located on either side of pinion gears, slightly overlap with the BG outer diameter to form lentil-shaped lubricant-wetted regions. Archival literature on the design and optimization of TCs is scant, in spite of their widespread usage as they are comprised of simple geometry mechanical elements. This paper presents an analysis of the hydrodynamic film pressure generated in a lubricated TC due to the rotation of both TC and BGs and specified taper angles for both bodies. The model solves the Reynolds equation of hydrodynamic lubrication to predict the operating film thickness that generates a pressure field reacting to impellers' thrust loads, these forces being a function of the pinion speed and the process gas physical properties. The model also predicts performance parameters, such as power loss and axial stiffness, and damping force coefficients. A parametric study brings out the taper angles in the TC and BG that balance the transmitted load with a lesser friction factor and peak pressure, along with large axial stiffness and damping.

Copyright © 2015 by ASME
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References

Wygant, K. D., 2013, “Samsung Techwin Integrally Geared Compressor,” Samsung Techwin, Houston, TX.
Sadykov, V. A., and Shneerson, L. M., 1968, “Helical Gear Transmissions With Thrust Collars,” Russ. Eng. J. USSR, 48(1), pp. 31–34.
Fingerhut, U., Rothstein, E., and Sterz, G., 1991, “Standardized Integrally Geared Turbomachines—Tailor Made for the Process Industry,” 20th Turbomachinery Symposium, Houston, TX, Sept. 17–19, pp. 131–145.
Deitz, P., and Mupende, I., 2006, “Pressure Ridge—An Old Machine Element With a New Potential Application,” Konstruktion, 58, pp. 69–75.
Thoden, D., 2006, Elasto-Hydrodynamic Lubrication of Pressure Ridges, Mechanical Engineering Department, Clausthal University of Technology, Clausthal, Germany, 31, pp. 23–26.
Thoden, D., 2009, Deformation of Thick Annular Plates Under Eccentric Axial Loading, Mitteilungen aus dem Institut fur Maschinenwesender, TU, Clausthal, 34, pp. 5–12.
Jackson, R., and Green, I., 2001, “Study of the Tribological Behavior of a Thrust Washer Bearing,” Tribol. Trans., 44(3), pp. 504–508. [CrossRef]
Jackson, R., and Green, I., 2006, “The Behavior of Thrust Washer Bearings Considering Mixed Lubrication and Asperity Contact,” Tribol. Trans., 49(2), pp. 233–247. [CrossRef]
Jackson, R., and Green, I., 2008, “The Thermoelastic Behavior of Thrust Washer Bearings Considering Mixed Lubrication, Asperity Contact and Thermoviscous Effects,” Tribol. Trans., 51(1), pp. 19–32. [CrossRef]
Yu, T. H., and Sadeghi, F., 2002, “Thermal Effects in Thrust Washer Lubrication,” ASME J. Tribol., 124(1), pp. 166–177. [CrossRef]
Zirkelback, N., and San Andrés, L., 1999, “Effect of Frequency Excitation on Force Coefficients of Spiral Groove Gas Seals,” ASME J. Tribol., 121(4), pp. 853–861. [CrossRef]
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Figures

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Fig. 1

Cut view of an IGC [1]

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Fig. 2

Schematic view of an IGC with TCs and thrust bearing shown

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Fig. 3

Depiction of lubricated (noncontacting) zone for a TC and BG

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Fig. 4

Schematic view of the lubricated zone for a TC and BG along line θ = 0

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Fig. 5

Radial and tangential components of BG surface velocity

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Fig. 6

Contour plots of (a) hydrodynamic pressure and (b) film thickness in a TC with (ϕ¯B/ϕ¯TC) = 1.0. W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 7

Contour plots of (a) hydrodynamic pressure and (b) film thickness in a TC with (ϕ¯B/ϕ¯TC) = 1.25. W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 8

Contour plots of (a) hydrodynamic pressure and (b) film thickness in a TC with(ϕ¯B/ϕ¯TC) = 0.8. W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 9

Pressure field and fluid flow in a lubricated TC with (ϕ¯B/ϕ¯TC) = 1. W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 10

Minimum film thickness (h¯min) versus TC taper angle (ϕ¯TC). W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 11

Maximum pressure (P¯max) versus TC taper angle (ϕ¯TC). W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 12

Friction factor (f) versus TC taper angle (ϕ¯TC). W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 13

Axial stiffness (K¯z) versus TC taper angle (ϕ¯TC). W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 14

Axial damping (C¯z) versus TC taper angle (ϕ¯TC). W¯ = 1,w¯ = 10/115,R2/R1 = 7.14.

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Fig. 15

Minimum film thickness (h¯min) versus load (W¯) for three taper angle configurations. w¯ = 10/115,R2/R1 = 7.14.

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Fig. 16

Friction factor (f) versus load (W¯) for three taper angle configurations. w¯ = 10/115,R2/R1 = 7.14.

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Fig. 17

Axial stiffness (K¯z) versus load (W¯) for three taper angle configurations. w¯ = 10/115,R2/R1 = 7.14.

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Fig. 18

Axial damping (C¯z) versus load (W¯) for three taper angle configurations. w¯ = 10/115,R2/R1 = 7.14.

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