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Research Papers: Other (Seals, Manufacturing)

Identification of Force Coefficients in a Squeeze Film Damper With a Mechanical Seal: Large Contact Force

[+] Author and Article Information
Adolfo Delgado1

Structural Mechanics and Dynamics Laboratory, GE Global Research Center, Niskayuna, NY 12309delgadoa@ge.com

Luis San Andrés

Department of Mechanical Engineering, Texas A&M University, College Station, TX 77843lsanandres@tamu.edu

Recall that in the experiments, the excitation forces are the actual inputs, and the displacements (x,y) and modeled nonlinear inputs (u,v) represent the outputs of the system.

1

Work conducted as a Research Assistant while at Texas A&M University.

J. Tribol 132(3), 032201 (Jun 15, 2010) (7 pages) doi:10.1115/1.4001458 History: Received June 21, 2009; Revised March 14, 2010; Published June 15, 2010; Online June 15, 2010

Squeeze film dampers (SFDs) aid to reduce excessive vibration levels due to rotor imbalance and to raise stability thresholds in rotor-bearing systems. SFDs commonly include end seals to increase their damping capability with a lesser lubricant flow. Seals also aid to reduce the occurrence of air ingestion/entrapment that severely reduces the damper forced performance. However, most conventional end seals do not completely eliminate lubricant side leakage, which limits their effectiveness to prevent air ingestion. A novel end seal arrangement incorporates a spring loaded, contacting mechanical seal that effectively prevents lubricant side leakage and air ingestion. The mechanically sealed damper is intended for use in power engines for unmanned aircraft vehicles. The test damper journal is 2.54 cm in length and 12.7 cm in diameter, with a radial clearance of 0.127 mm. Prior literature reports dynamic load tests on the seal-SFD and measurements of orbital motions to characterize the mechanical parameters of both the mechanical seal and squeeze film damper section. The test data to date include damper operation for a single contact load (90 N) closing the mechanical seal. Presently, measurements of damper dynamic load performance are conducted with a larger contact force (260 N). A nonlinear parameter identification method in the frequency domain determines simultaneously the squeeze film damping and inertia coefficients and the seal dry-friction force. The test results show that the system equivalent viscous damping coefficients are twice as large as those obtained earlier with the smaller contact force. On the other hand, as expected, the squeeze film damper coefficients are nearly identical for both test configurations. Predicted squeeze film damping coefficients, from an improved model that includes the flow in the damper feed and discharge grooves, correlate well with the test data for small and moderate orbit radii. The experimental fluid added mass coefficients are in par with the actual mass of the bearing housing and accurately predicted.

FIGURES IN THIS ARTICLE
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Copyright © 2010 by American Society of Mechanical Engineers
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References

Figures

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Figure 1

SFD with end mechanical contact seal: assembly cross section view

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Figure 2

Test rig for dynamic force measurements and flow visualization in squeeze film dampers (5)

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Figure 3

Assembly cut view of squeeze film land and mechanical contact seal at the bottom. The inset shows contact surfaces and normal (assembly) force FN.

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Figure 4

Applied dynamic loads (FY versus FX) and ensuing bearing displacement orbits (y(t) versus x(t)). Three load sets at excitation frequency ω=60 Hz. Damper clearance circle c=0.127 mm noted in the bottom graph.

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Figure 5

Schematic views of the equivalent representation of the SFD with mechanical seal

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Figure 6

Elements of a four-input/two-output representation of the (nonlinear) mechanical seal-SFD test system. D: time derivative operator (8).

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Figure 7

Dry-friction force Fd identified from circular centered orbits using nonlinear identification method in Ref. 8. Dry-friction from prior configuration with normal force 90 N also shown, Ref. 7.

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Figure 8

Test system dynamic stiffnesses versus excitation frequency. Symbol ●: real part of transfer functions Lxx and Lyy. Line: curve fit (Ks−ω2Ms). Circular centered orbits of radius 50 μm(Ksx=853 kN/m, Ksy=885 kN/m).

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Figure 9

Test system quadrature stiffnesses (ωC) versus excitation frequency. Symbol ●: imaginary part of transfer functions Lxx and Lyy. Line: CURVE fit (ωCSFD). Circular centered orbits of radius 50 μm. Circular centered orbits of amplitude x,y:50 μm.

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Figure 10

Test system damping (Cs−yy) and squeeze film damping (CSFDyy) coefficients versus excitation frequency for orbit radii 25 μm, 38 μm, and 50 μm. Tests with two contact forces FN=260 N and 90 N (Fd=97 N and 34 N) (7).

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Figure 11

Squeeze film damping coefficient CSFDyy versus orbit amplitude. Experiments with contact forces FN=90 N and 260 N. Predictions from Ref. 10.

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Figure 12

Bottom of the damper journal. Contact surface indicated.

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Figure 13

Top view of the ring carrier installed in bearing assembly with detailed view of worn surface. Contact surface noted.

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