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Research Papers

Analysis of Counter-Rotating Roller Bearing in Different Mounting Configurations

[+] Author and Article Information
Wenjun Gao

Univ Lyon, INSA-Lyon, CNRS,
UMR5259 LaMCoS,
Villeurbanne F-69621, France
e-mail: gaowenjun@nwpu.edu.cn

Daniel Nelias

Professor
Univ Lyon, INSA-Lyon, CNRS,
UMR5259 LaMCoS,
Villeurbanne F-69621, France
e-mail: daniel.nelias@insa-lyon.fr

Zhenxia Liu

Professor
School of Power and Energy,
NPU,
Xi'an 710072, China
e-mail: zxliu@nwpu.edu.cn

1Present address: School of Power and Energy, NPU, Xi'an 710072, China.

Manuscript received January 16, 2018; final manuscript received March 9, 2019; published online April 3, 2019. Assoc. Editor: Philip Bonello.

J. Eng. Gas Turbines Power 141(8), 081008 (Apr 03, 2019) (8 pages) Paper No: GTP-18-1023; doi: 10.1115/1.4043216 History: Received January 16, 2018; Revised March 09, 2019

Advanced engine configuration studies have shown large advantages for an engine with counter-rotating spools with intershaft counter-rotating roller bearings. Mounted on two counter-rotating differential-speed hollow rotors, the bearing internal kinetic behavior, dynamic behavior, and then thermal behavior change greatly, causing a severe challenge to engine designers using traditional analysis methods. A special quasi-dynamic model for counter-rotating roller bearing is proposed, considering rings deformation and windage effects, to analyze the bearing mechanical and thermal behavior in different mounting configurations. Roller sliding and bearing heat generation are calculated and compared with experimental data to verify the model capabilities. It shows that the configuration that connects the inner ring to the high-speed rotor has life cycle advantage with more uniform load distribution, smaller roller/ring clearance, and lower cage speed. This leads to less drag loss due to the rotation of the rollers and cage assembly. The decrease of the total power loss is a key element to minimize the quantity of oil required to lubricate the roller bearing.

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Figures

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Fig. 1

The natural (top) and inverted (bottom) configurations

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Fig. 2

Frontal area (in gray) used for the estimation of the drag force acting against the translation of the roller

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Fig. 3

Roller in the cage pocket

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Fig. 4

Boundary layer flow on a rotating disk [11]

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Fig. 5

Geometrical relation linked to the relative displacement between the inner and outer ring centers

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Fig. 6

Forces and torques acting on the roller

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Fig. 7

Forces and torques acting on the cage

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Fig. 8

The roller bearing configuration

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Fig. 9

Effect of centrifugal forces due to the rotation of the rings on the diametral clearance of the roller bearing

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Fig. 10

Deformation of inner ring caused by external load

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Fig. 11

Deformation of outer ring

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Fig. 12

Modification of the diametral clearance for the natural (black triangle symbol plot bottom right) and inverted designs (green dot symbol plot top left), in mm. Note the sign: for the natural design the local diametral clearance increases whereas it diminishes for the inverted one.

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Fig. 13

Comparison of static and dynamic load distributions

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Fig. 14

Roller sliding versus radial load

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Fig. 15

Roller rotation speed for the natural configuration (ωi = 7500 rpm, ωo = 13,300 rpm)

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Fig. 16

Roller rotation speed for the inverted configuration (ωi = 13,300 rpm, ωo = 7500 rpm)

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Fig. 17

Cage rotation speed versus radial load for the natural configuration (ωi = 7500 rpm, ωo = 13,300 rpm)

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Fig. 18

Cage rotation speed versus radial load for the inverted configuration (ωi = 13,300 rpm, ωo = 7500 rpm)

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Fig. 19

Predicted and experimental total power loss versus radial load

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Fig. 20

Total power loss predicted versus radial load for the natural (full lines) and inverted (dash lines) configurations

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Fig. 21

Local power loss predicted versus radial load for the natural (full lines—8600/14,300 rpm) and inverted (dash lines—14,300/8600 rpm) configurations. The figure on the right is a zoom for the lowest contributions.

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