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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

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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