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Research Papers: Gas Turbines: Structures and Dynamics

Application of Computational Fluid Dynamics Simulation to Squeeze Film Damper Analysis

[+] Author and Article Information
Gil Jun Lee

Department of Mechanical and
Materials Engineering,
College of Engineering and Applied Science,
University of Cincinnati,
584D Rhodes Hall, 2600 Clifton Avenue,
Cincinnati, OH 45221
e-mail: leeg4@mail.uc.edu

Jay Kim

Fellow ASME
Department of Mechanical and
Materials Engineering,
College of Engineering and Applied Science,
University of Cincinnati,
589 Rhodes Hall, 2600 Clifton Avenue,
Cincinnati, OH 45221
e-mail: jay.kim@uc.edu

Tod Steen

Mem. ASME
GE Aviation,
1 Neumann Way,
Cincinnati, OH 45215
e-mail: tod.steen@ge.com

1Corresponding author.

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 18, 2016; final manuscript received March 29, 2017; published online May 16, 2017. Assoc. Editor: Alexandrina Untaroiu.

J. Eng. Gas Turbines Power 139(10), 102501 (May 16, 2017) (11 pages) Paper No: GTP-16-1347; doi: 10.1115/1.4036511 History: Received July 18, 2016; Revised March 29, 2017

Squeeze film dampers (SFDs) are used in high-speed turbomachinery to provide external damping to the system. Computational fluid dynamics (CFD) simulation is a highly effective tool to predict the performance of SFDs and obtain design guidance. It is shown that a moving reference frame (MRF) can be adopted for CFD simulation, which saves computational time significantly. MRF-based CFD analysis is validated, then utilized to design oil plenums of SFDs. Effects of the piston ring clearances, the oil groove, and oil supply ports are studied based on CFD and theoretical solutions. It is shown that oil plenum geometries can significantly affect the performance of the SFD especially when the SFD has a small clearance. The equivalent clearance is proposed as a new concept that enables quick estimation of the effect of oil plenum geometries on the SFD performance. Some design practices that have been adopted in industry are revisited to check their validity. Based on simulation results, a set of general design guidelines is proposed.

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References

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Xing, C. , Braun, M. J. , and Li, H. , 2009, “ A Three-Dimensional Navier–Stokes-Based Numerical Model for Squeeze Film Dampers—Part 1: Effects of Gaseous Cavitation on Pressure Distribution and Damping Coefficient Without Consideration of Inertia,” Tribol. Trans., 52(2), pp. 680–694. [CrossRef]
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Figures

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Fig. 1

Schematic of a typical squeeze film damper

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Fig. 2

(a) Geometry of the SFD in a synchronous circular whirling orbit and the coordinate system and (b) the motion of the outer surface of SFD. (e: eccentricity, ω: whirl speed, ωs: angular velocity of a shaft, ωR/s: angular velocity of an outer race of a bearing, O: center of a housing, O’: center of a shaft, rh: radius of a housing, rb: radius of a bearing, X–Y: inertia coordinate, x–y: moving coordinate rotating about O.)

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Fig. 8

Three-dimensional plots of pressure field on the rotor of SFD. (a) h/C = 0 (no groove) and (b) h/C = 65.3. h and C are the depth of the center groove and the clearance of SFD, respectively.

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Fig. 7

Schematic of a squeeze film damper with a center groove

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Fig. 6

Circumferential pressure distributions of an SFD obtained from Reynolds equation and obtained from CFD with five different Reynolds numbers (Re)

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Fig. 5

The ratio of the results from Reynolds long bearing equation to those from CFD simulation. (Ft, Reynolds: tangential force obtained from Reynolds equation, Ft,CFD: tangential force obtained from CFD, ΔPReynolds: the difference between maximum and minimum pressure obtained from Reynolds equation, ΔPCFD: the difference between maximum and minimum pressure obtained from CFD.)

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Fig. 4

Circumferential pressure distribution for long SFD with Re = 1 × 10−11

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Fig. 3

Schematic of a squeeze film damper without grooves

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Fig. 9

Short bearing approximation for a SFD with a center groove

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Fig. 10

(a) Three-dimensional plot of the pressure field of an SFD and a JB with a center groove (h/C = 50, h: the depth of the center groove, C: clearance, Re = 1 × 10−11), (b) axial pressure distribution at 90 deg of circumferential location, and (c) axial pressure distribution at 270 deg of circumferential location

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Fig. 11

Circumferential pressure distribution on the rotor at the midplane (L = 0) with various groove depths. (h: depth of the groove, w: width of the groove, C: clearance of SFD, L: length of SFD.)

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Fig. 14

Pressure distribution of the SFD with different axial clearances (Ca) of the piston ring grooves (a) in circumferential direction at the midplane (L = 0) and (b) in axial direction at the circumferential location of the maximum pressure. C, L, and Lh are the clearance, the length of an SFD, and the height of a piston ring, respectively.

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Fig. 15

Pressure distribution of the SFD with different radial clearance (Cr) of piston ring grooves (a) in circumferential direction at the midplane (L = 0) and (b) in axial direction at the circumferential location of maximum pressure. Lw, C, and L are the width of the piston ring, clearance of the SFD, and length of the SFD, respectively.

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Fig. 16

Pressure distribution of the SFD with different widths of a piston ring (Lw) in circumferential direction at the midplane (L = 0)

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Fig. 18

(a) Circumferential pressure distribution of the SFD with a center groove and corresponding equivalent clearances (Ceq1 and Ceq2, no groove) at the midplane (L = 0), and (a) hw/CL = 3.50 and (b) hw/CL = 14.0. h, w, L, and C are groove depth, groove width, the length of SFD, and the clearance of SFD, respectively.

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Fig. 19

Forces of SFD with a center groove and corresponding equivalent clearances (Ceq1 and Ceq2, no groove) and Reynolds long bearing solution with respect to nondimensional groove depth (hw/CL). (a) Normalized radial force and (b) Normalized damping coefficient. h, w, L, and C are groove depth, groove width, the length of SFD, and the clearance of SFD, respectively.

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Fig. 20

(a) Schematic of the SFD with oil inlets and outlets and (b) its 3D CFD model

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Fig. 21

(a) Circumferential pressure distribution at the midplane (L = 0) with different ratios of the area (Ra) between inlets (Ai) and outlets (Ao) and (b) corresponding normalized mass flow rate and damping coefficient

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Fig. 12

Radial and tangential forces of SFD with respect to groove depth (h). C, L, and w are the clearance of SFD, the length of SFD, and groove width.

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Fig. 13

Schematic of the SFD with piston ring seals. L is the length of the SFD, C is the clearance, Lw and Lh are the width of a piston ring and the height of a piston ring, Ca and Cr are axial and radial clearance of the piston ring groove, respectively.

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Fig. 22

Circumferential pressure distribution at the midplane (L = 0) of an SFD with the volume of inlets and without oil flow. di is a diameter of an inlet.

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Fig. 23

(a) Circumferential pressure distribution at the midplane (L = 0) force with different area of outlets (Ao) and (b) corresponding normalized mass flow rate and damping coefficient

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Fig. 24

(a) Circumferential pressure distribution at the midplane with different supply pressure (Ps) and (b) corresponding normalized damping coefficient and mass flow rate

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Fig. 17

Cross section of the general SFD with a groove

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