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Research Papers: Flows in Complex Systems

Methodology of Turbulence Parameter Correction in Water-Lubricated Thrust Bearings

[+] Author and Article Information
Xin Deng

Mem. ASME
Rotating Machinery and Controls (ROMAC) Lab,
Mechanical and Aerospace Engineering,
University of Virginia,
Charlottesville, VA 22904
e-mail: xd9fw@virginia.edu

Harrison Gates

Mem. ASME
Rotating Machinery and Controls (ROMAC) Lab,
Mechanical and Aerospace Engineering,
University of Virginia,
Charlottesville, VA 22904
e-mail: hrg9aa@virginia.edu

Roger Fittro

Mem. ASME
Rotating Machinery and Controls (ROMAC) Lab,
Mechanical and Aerospace Engineering,
University of Virginia,
Charlottesville, VA 22904
e-mail: rlf9w@virginia.edu

Houston Wood

Mem. ASME
Rotating Machinery and Controls (ROMAC) Lab,
Mechanical and Aerospace Engineering,
University of Virginia,
Charlottesville, VA 22904
e-mail: hgw9p@virginia.edu

Contributed by the Fluids Engineering Division of ASME for publication in the JOURNAL OF FLUIDS ENGINEERING. Manuscript received July 2, 2018; final manuscript received November 26, 2018; published online January 7, 2019. Assoc. Editor: Oleg Schilling.

J. Fluids Eng 141(7), 071104 (Jan 07, 2019) (9 pages) Paper No: FE-18-1452; doi: 10.1115/1.4042161 History: Received July 02, 2018; Revised November 26, 2018

Oil-lubricated bearings are widely used in high-speed rotating machines such as those found in automotive industries and aerospace. However, environmental issues and risk-averse operations are resulting in the removal of oil and the replacement of all sealed oil bearings with reliable water-lubricated bearings. The low viscosity of water increases Reynolds numbers drastically and therefore makes water-lubricated bearings prone to turbulence effects. This requires finer meshes for finite element modeling when compared to oil-lubricated bearings as the low-viscosity fluid produces a very thin lubricant film. Analyzing water-lubricated bearings can also produce convergence and accuracy issues in traditional oil-based analysis codes. Fitting the velocity profile with experiments having a nondimensional wall distance y+ in a certain range results in Ng-optimized Reichardt's constants k and δ+. The definition of y+ can be used to approximate the first layer thickness calculated for a uniform mesh. On the condition that the y+ is fixed to that of a standard oil bearing for which an oil-bearing code was validated, the number of elements across the film thickness and coefficients used in the eddy-viscosity equation can be adjusted to allow for convergence with other fluids other than that which the traditional oil-bearing code was designed for. This study proposed a new methodology to preserve the y+ value to make water-lubricated thrust bearing models valid. A method for determining the required number of cross-film elements in water-lubricated bearings was found. The results of this study could aid in improving future designs and models of water-lubricated bearings.

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Figures

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Fig. 1

Working principle of fluid film bearings: convergent wedge

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Fig. 3

Turbulent boundary layer

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Fig. 4

Adjusted δ+ versus pivot Reynolds number [25]

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Fig. 5

Comparison of minimum film thickness between the benchmark code and thrust bearing code with and without an optimized cross-film element number

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Fig. 6

Comparison of pivot film thickness between the benchmark code and thrust bearing code with and without an optimized cross-film element number

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Fig. 7

Relative difference of minimum film thickness predictions between the benchmark code and thrust bearing code with and without a modified cross-film element number

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Fig. 8

Relative difference of pivot film thickness predictions between the benchmark code and thrust bearing code with and without an optimized cross-film element number

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Fig. 9

Comparison of maximum pressure in the fluid film between the benchmark code and thrust bearing code with and without an optimized cross-film element number

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Fig. 10

Pressure contour of the highest load in thrust bearing code (a) with and (b) without an optimized cross-film element number

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Fig. 11

Comparison of power loss between the benchmark code and thrust bearing code with and without an optimized cross-film element number

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Fig. 12

Relative difference of power loss predictions between the benchmark code and thrust bearing code with and without an optimized cross-film element number

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Fig. 13

Mesh independent study with different cross-film element number for oil lubrication: (a) lowest load and (b) highest load

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Fig. 14

Mesh independent study with different cross-film element number for water lubrication: (a) lowest load and (b) highest load

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Fig. 15

Comparison of inlet film thickness between the experiment, THRUST and CFD

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Fig. 16

New approximation for δ+

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Fig. 17

Minimum film thickness sensitivity to δ+ and load

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Fig. 18

Minimum film thickness sensitivity to pivot Reynolds number and load

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