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Research Papers: Gas Turbines: Structures and Dynamics

Experimental Performance of an Open Ends, Centrally Grooved, Squeeze Film Damper Operating With Large Amplitude Orbital Motions

[+] Author and Article Information
Luis San Andrés

Mast-Childs Professor
Fellow ASME
Mechanical Engineering Department,
Texas A&M University,
College Station, TX 77843
e-mail: Lsanandres@tamu.edu

Sung-Hwa Jeung

Mechanical Engineering Department,
Texas A&M University,
College Station, TX 77843
e-mail: sean.jeung@gmail.com

The simplified model assumes that the central groove, taken as an infinite source or sink of flow, is impervious to the kinetics of the journal motion.

Fluid inertia due to advection effects (spatial changes in momentum) are ignored in the current model.

A true (linearized) force coefficient, say -KXX=(FX/X)limΔX0(ΔFX/ΔX), implies changes in force due to an infinitesimally small change in displacement.

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 15, 2014; final manuscript received July 17, 2014; published online October 7, 2014. Editor: David Wisler.

J. Eng. Gas Turbines Power 137(3), 032508 (Oct 07, 2014) (9 pages) Paper No: GTP-14-1397; doi: 10.1115/1.4028376 History: Received July 15, 2014; Revised July 17, 2014

Aircraft engines customarily implement squeeze film dampers (SFDs) to dissipate mechanical energy caused by rotor vibration and to isolate the rotor from its structural frame. The paper presents experimental results for the dynamic forced performance of an open ends SFD operating with large amplitude whirl motions, centered and off-centered. The test rig comprises of an elastically supported bearing with a damper section, 127 mm in diameter, having two parallel film lands separated by a central groove. Each film land is 25.4 mm long with radial clearance c = 0.251 mm. The central groove, 12.7 mm long, has a depth of 9.5 mm (38c). An ISO VG 2 lubricant flows into the groove via three 2.5 mm orifices, 120 deg apart, and then passes through the film lands to exit at ambient condition. Two orthogonally placed shakers apply dynamic loads on the bearing to induce circular orbit motions with whirl frequency ranging from 10 Hz to 100 Hz. A static loader, 45 deg away from each shaker, pulls the bearing to a static eccentricity (es). Measurements of dynamic loads and the ensuing bearing displacements and accelerations, as well as the film and groove dynamic pressures, were obtained for eight orbit amplitudes (r = 0.08c to ∼0.71c) and under four static eccentricities (es = 0.0c to ∼0.76c). The experimental damping coefficients increase quickly as the bearing offset increases (es/c → 0.76) while remaining impervious to the amplitude of whirl orbit (r/c → 0.51). The inertia coefficients decrease rapidly as the orbit amplitude grows large, r > 0.51c, but increase with the static eccentricity. A comparison with test results obtained with an identical damper but having a smaller clearance (cs = 0.141 mm) (San Andrés, L., 2012, “Damping and Inertia Coefficients for Two Open Ends Squeeze Film Dampers With a Central Groove: Measurements and Predictions,” ASME J. Eng. Gas Turbines Power, 134(10), p. 102506), show the prior damping and inertia coefficients are larger, ∼5.0 and ∼2.2 times larger than the current ones. These magnitudes agree modestly with theoretical ratios for damping and inertia coefficients scaling as (c/cs)3= 5.7 and (c/cs) = 1.8, respectively. In spite of the large difference in depths between a groove and a film land, the magnitudes of dynamic pressures recorded at the groove are similar to those in the lands. That is, the groove profoundly affects the dynamic forced response of the test damper. A computational physics model replicates the experimental whirl motions and predicts force coefficients spanning the same range of whirl frequencies, orbit radii, and static eccentricities. The model predictions reproduce with great fidelity the experimental force coefficients. The good agreement relies on the specification of an effective groove depth derived from one experiment.

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References

San Andrés, L., 2012, “Squeeze Film Damper: Operation, Models and Technical Issues,” Modern Lubrication Theory, Notes 13, Texas A&M University, College Station, TX, available at: https://repository.tamu.edu/handle/1969.1/93197
Vance, J., Zeidan, F., and Murphy, B., 2010, “Bearings and Their Effect on Rotordynamics,” Machinery Vibration and Rotordynamics, Wiley, New York, pp. 216–238.
Della Pietra, L., and Adiletta, G., 2002, “The Squeeze Film Damper Over Four Decades of Investigations. Part I: Characteristics and Operating Features,” Shock Vib. Dig., 34(1), pp. 3–26.
Adiletta, G., and Della Pietra, L., 2002, “The Squeeze Film Damper Over Four Decades of Investigations. Part II: Rotordynamics Analysis With Rigid and Flexible Rotors,” Shock Vib. Dig, 34(2), pp. 97–126.
San Andrés, L., 1992, “Analysis of Short Squeeze Film Dampers With a Central Groove,” ASME J. Tribol., 114(4), pp. 659–664. [CrossRef]
San Andrés, L., and Vance, J. M., 1987, “Experimental Measurement of the Dynamic Pressure Distribution in a Squeeze Film Damper Executing Circular Centered Orbits,” ASLE Trans., 30(3), pp. 373–383. [CrossRef]
Zhang, J., and Roberts, J. B., 1996, “Force Coefficients for a Centrally Grooved Short Squeeze Film Damper,” ASME J. Tribol., 118(3), pp. 608–616. [CrossRef]
Qingchang, T., Ying, C., and Lyjiang, W., 1997, “Effect of a Circumferential Feeding Groove on Fluid Force in Short Squeeze Film Dampers,” Tribol. Int., 30(6), pp. 409–416. [CrossRef]
Lund, J. W., Myllerup, C. M., and Hartmann, H., 2003, “Inertia Effects in Squeeze-Film Damper Bearings Generated by Circumferential Oil Supply Groove,” ASME J. Vib. Acoust., 125(4), pp. 495–499. [CrossRef]
Kim, K. J., and Lee, C. W., 2005, “Dynamic Characteristics of Sealed Squeeze Film Damper With a Central Feeding Groove,” ASME J. Tribol., 127(1), pp. 103–111. [CrossRef]
Gehannin, J., Arghir, M., and Bonneau, O., 2010, “Complete Squeeze-Film Damper Analysis Based on the “Bulk Flow” Equations,” STLE Tribol. Trans., 53(1), pp. 84–96. [CrossRef]
San Andrés, L., 2012, “Damping and Inertia Coefficients for Two Open Ends Squeeze Film Dampers With a Central Groove: Measurements and Predictions,” ASME J. Eng. Gas Turbines Power, 134(10), p. 102506. [CrossRef]
Delgado, A., and San Andrés, L., 2010, “A Model for Improved Prediction of Force Coefficients in Grooved Squeeze Film Dampers and Oil Seal Rings,” ASME J. Tribol., 132(3), p. 032202. [CrossRef]
San Andrés, L., and Delgado, A., 2012, “A Novel Bulk-Flow Model for Improved Predictions of Force Coefficients in Grooved Oil Seals Operating Eccentrically,” ASME J. Eng. Gas Turbines Power, 134(5), p. 052509. [CrossRef]
San Andrés, L., and Seshagiri, S., 2013, “Damping and Inertia Coefficients for Two End Sealed Squeeze Film Dampers With a Central Groove: Measurements and Predictions,” ASME J. Eng. Gas Turbines Power, 135(12), p. 112503. [CrossRef]
Seshagiri, S., 2011, “Identification of Force Coefficients in Two Squeeze Film Dampers With a Central Groove,” M.S. thesis, Texas A&M University, College Station, TX.
Mahecha, P., 2011, “Experimental Dynamic Forced Performance of a Centrally Grooved, End Sealed Squeeze Film Damper,” M.S. thesis, Texas A&M University, College Station, TX.
San Andrés, L., 2014, “Force Coefficients for a Large Clearance Open Ends Squeeze Film Damper With a Central Groove: Experiments and Predictions,” Tribol. Int., 71, pp. 17–25. [CrossRef]
Jeung, S.-H., 2013, “Performance of an Open Ends Squeeze Film Damper Operating With Large Amplitude Orbital Motions: Experimental Analysis and Assessment of the Accuracy of the Linearized Force Coefficients Model,” M.S. thesis, Texas A&M University, College Station, TX.
Zhang, J., Roberts, J. B., and Ellis, J., 1994, “Experimental Behavior of a Short Cylindrical Squeeze Film Damper Executing Circular Centered Orbits,” ASME J. Tribol., 116(3), pp. 528–534. [CrossRef]
San Andrés, L., Jeung, S.-H., and Bradley, G., 2013, “Experimental Force Coefficients for an Open Ends Squeeze Film Damper Performing Large Amplitude Circular Orbital Motions, Centered and Off-Centered,” 68th STLE Annual Meeting and Exhibition, Detroit, MI, May 5–9.
Fritzen, C. P., 1985, “Identification of Mass, Damping, and Stiffness Matrices of Mechanical System,” ASME J. Vib., Acoust., 108(1), pp. 9–16. [CrossRef]
San Andrés, L., 2012, “Experimental Identification of Bearing Force Coefficients,” Modern Lubrication Theory, Notes 14, Texas A&M University, College Station, TX, available at: https://repository.tamu.edu/handle/1969.1/93197
San Andrés, L., and Vance, J., 1986, “Effects of Fluid Inertia and Turbulence on the Force Coefficients for Squeeze Film Dampers,” ASME J. Eng. Gas Turbines Power, 108(2), pp. 332–339. [CrossRef]

Figures

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Fig. 1

Schematic view of an open ends SFD with a central feed groove [1]

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Fig. 2

Cross section view of SFD test rig

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Fig. 3

Cut section view of test SFD outlining the lubricant flow path

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Fig. 4

Real part of the test system direct impedances (HXX, HYY) versus excitation frequency. Tests with circular orbits with amplitudes r/c = 0.08, 0.51, and 0.71 and centered condition (eS = 0.0c).

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Fig. 5

Imaginary part of the test system direct impedances (HXX, HYY) versus excitation frequency. Tests with circular orbits with amplitudes r/c = 0.08, 0.51, and 0.71 and centered condition (eS = 0.0c).

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Fig. 6

SFD direct dynamic force coefficients (C, K, M)SFD versus orbit amplitude (r/c) at the centered condition (es = 0.0c) and three static eccentricities (es = 0.25c, es = 0.51c, and es = 0.76c). Frequency range 10–100 Hz.

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Fig. 7

Open ends SFD: comparison of direct damping (C)SFD and added mass (M)SFD coefficients versus static eccentricity (es) for two clearances c = 141 μm [12] and c = 251 μm and orbit amplitudes r = 14 μm and 20 μm

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Fig. 8

Disposition of dynamic pressure sensors in BC. Open ends damper with 12.7 mm land lengths.

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Fig. 9

Peak-to-peak dynamic pressures versus excitation frequency at (a) top film land and (b) central groove. Tests with circular centered (eS = 0) orbits with radii r = 0.08c to r = 0.71c. Insets show dynamic pressures versus time for r/c = 0.40 and ω = 90 Hz.

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Fig. 10

Open ends SFDs: direct damping (CXX, CYY)SFD coefficients versus static eccentricity (es). Circular orbits with amplitude r = 0.08c, 0.30c, and 0.51c. Predictions with effective groove depth dη = 1.75c.

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Fig. 11

Open ends SFDs: direct added mass (MXX, MYY)SFD coefficients versus static eccentricity (es). Circular orbits with amplitude r = 0.08c, 0.30c, and 0.51c. Predictions with effective groove depth dη = 1.75c.

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