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Research Papers

An Efficient Quasi-Three-Dimensional Model of Tilting Pad Journal Bearing for Turbomachinery Applications

[+] Author and Article Information
Andrea Rindi

Professor
Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: andrea.rindi@unifi.it

Stefano Rossin

General Electric Oil & Gas,
Florence 50100, Italy
e-mail: stefano.rossin@ge.com

R. Conti

Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: roberto.conti@unifi.it

A. Frilli

Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: amedeo.frilli@unifi.it

E. Galardi

Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: emanuele.galardi@unifi.it

E. Meli

Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: enrico.meli@unifi.it

D. Nocciolini

Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: daniele.nocciolini@unifi.it

L. Pugi

Department of Industrial Engineering,
University of Florence,
Florence 50100, Italy
e-mail: luca.pugi@unifi.it

Contributed by the Technical Committee on Vibration and Sound of ASME for publication in the JOURNAL OF VIBRATION AND ACOUSTICS. Manuscript received November 10, 2014; final manuscript received June 23, 2015; published online October 6, 2015. Assoc. Editor: Patrick S. Keogh.

J. Vib. Acoust 137(6), 061013 (Oct 06, 2015) (17 pages) Paper No: VIB-14-1434; doi: 10.1115/1.4031408 History: Received November 10, 2014; Revised June 23, 2015

The constant increase of turbomachinery rotational speed has brought the design and the use of journal bearings to their very limits: tilting pad journal bearings (TPJBs) have been introduced for high-speed/high-load applications due to their intrinsic stability properties and can be used both in transient and steady-state operations obtaining superior performances. An accurate analysis of the TPJBs behavior is essential for a successful design and operation of the system; however, it is necessary to reach a compromise between the accuracy of the results provided by the TPJB model and its computational cost. This research paper exposes the development of an innovative and efficient quasi-3D TPJB modeling approach that allows the simultaneous analysis of the system rotordynamics and the supply plant behavior; the majority of existing models describe these aspects separately but their complex interaction must be taken into account to obtain a more accurate characterization of the system. Furthermore, the proposed model is characterized by a high numerical efficiency and modularity, allowing for complex transient simulations of the complete plant and for the representation of different kind of bearings. The TPJB model has been developed and experimentally validated in collaboration with an industrial partner which provided the technical data of the system and the results of experimental tests.

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References

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Figures

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Fig. 2

Numerical components of the proposed algorithm

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Fig. 3

General architecture of the whole model

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Fig. 4

TPJB structure, control volume, and lubricant supply plant

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Fig. 5

Geometric and kinematic representation of the oil film

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Fig. 6

Control volume, boundaries, and flow rates

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Fig. 7

Kinematic viscosity for industrial oils in accordance with ISO 3448:1992 [37]

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Fig. 9

Scheme of the generic 3D beam element

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Fig. 10

Cross section of the beam element

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Fig. 11

Lumped parameters sump model

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Fig. 12

Numerical flowchart of the solution algorithm

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Fig. 13

Simple scheme of the experimental rotor test rig

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Fig. 14

Scheme of the test procedure

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Fig. 15

Rotor vibration amplitudes provided by the developed model xAj, yAj compared to experimental data xA,measj, yA,measj

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Fig. 16

Rotor vibration phases provided by the developed model ϕxj, ϕyj compared to experimental data ϕx,measj, ϕy,measj

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Fig. 17

Comparison between the experimental pressure field (as a function of the bearing angular coordinate) presented in Ref. [50] and the results obtained with the proposed model (3000 rpm, applied load equal to 180 kN)

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Fig. 18

Comparison between the rotor vibrations (balanced rotor, x-axis for the DE bearing) obtained with the proposed model and a conventional bearing model developed using linearized coefficients

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Fig. 19

Comparison between the experimental rotor orbit (xA1 and yA1) and those obtained with the proposed model (DE bearing location, 2500 rpm, unbalanced rotor)

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Fig. 20

Comparison between the experimental rotor orbit (xA1 and yA1) and those obtained with the proposed model (DE bearing location, 4500 rpm, unbalanced rotor)

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Fig. 21

Pads rotations γpadi,1 (DE bearing location, 2500 rpm, unbalanced rotor)

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Fig. 22

Pads rotations γpadi,1 (DE bearing location, 4000 rpm, unbalanced rotor)

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Fig. 23

Sumps pressures pi,1 (DE bearing location, 2500 rpm, unbalanced rotor)

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Fig. 24

Sumps pressures pi,1 (DE bearing location, 4000 rpm, unbalanced rotor)

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Fig. 25

Three-dimensional representation of the oil film pressure field p (a) and velocity field v¯ (b) on the surfaces of the five pads Spad (DE bearing location, 2500 rpm, unbalanced rotor)

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Fig. 26

Three-dimensional representation of the oil film pressure field p (a) and velocity field v¯ (b) on the surfaces of the five pads Spad (DE bearing location, 4000 rpm, unbalanced rotor)

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