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Research Papers: Hydrodynamic Lubrication

Effect of Oil Supply Pressure on the Force Coefficients of a Squeeze Film Damper Sealed With Piston Rings

[+] Author and Article Information
Sung-Hwa Jeung

Department of Mechanical Engineering,
Texas A&M University,
College Station, TX 77843
email: Sean.Jeung@gmail.com

Luis San Andrés

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

Sean Den

Department of Mechanical Engineering,
Texas A&M University,
College Station, TX 77843
email: sean.thewind@yahoo.com

Bonjin Koo

Department of Mechanical Engineering,
Texas A&M University,
College Station, TX 77843
e-mail: bjkoo@tamu.edu

3Corresponding author.

Contributed by the Tribology Division of ASME for publication in the Journal of Tribology. Manuscript received September 29, 2018; final manuscript received March 17, 2019; published online April 16, 2019. Assoc. Editor: Noel Brunetiere.

1Present address: Senior Compressor Design Engineer, Compressor Tech. & Development Ingersoll Rand, La Crosse, WI 54601.

2Present address: Mechanical Maintenance Engineer, Formosa Plastics Corp., Point Comfort, TX 77978.

J. Tribol 141(6), 061701 (Apr 16, 2019) (11 pages) Paper No: TRIB-18-1406; doi: 10.1115/1.4043238 History: Received September 29, 2018; Accepted March 18, 2019

Squeeze film dampers (SFDs) aid to both reduce rotor dynamic displacements and to increase system stability. Dampers sealed with piston rings (PR), common in aircraft engines, are proven to boost damping generation, reduce lubricant flow demand, and prevent air ingestion. This paper presents the estimation of force coefficients in a short length SFD, PR sealed, and supplied with a light lubricant at two feed pressures, Pin-1 ∼ 0.69 barg and Pin-2 ∼ 2.76 barg, i.e., low and high. Two pairs of PRs are installed in the test SFD, one set has flow conductance CS1 = 0.56 LPM/bar, whereas the other pair has CS2 = 0.89 LPM/bar. The second set leaks more as it has a larger slit gap. Dynamic load tests show that both dampers, having seal flow conductances differing by 60%, produce damping and added mass coefficients of similar magnitude, differing by at most 20%. Other experiments quantify the effect of lubricant supply pressure, Pin-1 and Pin-2, on the dynamic film pressure and force coefficients of the PR-SFD. The damper configuration with CS1 and operating with the high Pin-2 shows ∼20% more damping and added mass coefficients compared with test results for the damper supplied with Pin-1. Film pressure measurements show that the air ingestion and oil vapor cavitation coexist for operation at the low Pin-1. Computational predictions accounting for the feed holes in the physical model agree with the experimental coefficients. On the other hand, predictions from classical formulas for an idealized damper geometry, fully sealed at its ends, largely overpredict the measured force coefficients.

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References

San Andrés, L., 2012, “Squeeze Film Dampers: Operation, Models and Technical Issues,” Modern Lubrication Theory, Notes 13, Texas A&M University Digital Libraries, http://oaktrust.library.tamu.edu/handle/1969.1/93253. Accessed June 27, 2016.
Childs, D., 2013, Turbomachinery Rotordynamics With Case Studies, Minter Spring Pubs., TX, Chap. 6.
Cooper, S., 1963, “Preliminary Investigation of Oil Films for Control of Vibration,” Proceedings of Lubrication and Wear Convention, IMechE, pp. 305–315.
Zeidan, F. Y., Vance, J. M., and San Andrés, L., 1996, “Design and Application of Squeeze Film Dampers in Rotating Machinery,” Proceedings of the 25th Turbomachinery Symposium, Texas A&M University, Houston, TX, pp. 169–188.
Kuzdal, M. J., and Hustak, J. F., 1996, “Squeeze Film Damper Bearing Experimental vs Analytical Results for Various Damper Configurations,” Proceedings of the 25th Turbomachinery Symposium, Texas A&M University, Houston, TX, pp. 57–70.
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., Jeung, S.-H., Den, S., and Savela, G., 2016, “Squeeze Film Dampers: An Experimental Appraisal of Their Dynamic Performance,” Proceedings of the 2016 Asia Turbomachinery & Pump Symposium, Marina Bay Sands, Singapore, Feb. 22–25, pp. 1–23.
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]
Jeung, S.-H., San Andrés, L., and Bradley, G., 2016, “Forced Coefficients for a Short Length, Open-Ends Squeeze Film Damper With End Grooves: Experiments and Predictions,” ASME J. Eng. Gas Turbines Power, 138(2), p. 032502.
Meng, G., San Andrés, L., and Vance, J., 1991, “Experimental Measurement of the Dynamic Pressure and Force Response of a Partially Sealed Squeeze Film Damper,” Proceedings of the 13th Biennial Conference on Mechanical Vibration and Noise, Miami, FL, Sept. 22–25, pp. 251–256.
Levesley, M., and Holmes, R., 1996, “The Effect of Oil Supply and Sealing Arrangements on the Performance of Squeeze-Film Dampers: An Experimental Study,” Proc. Inst. Mech. Eng. Part J: J. Eng. Tribol., 210(4), pp. 221–232. [CrossRef]
Defaye, C., Arghir, M., and Bonneau, O., 2006, “Experimental Study of the Radial and Tangential Forces in a Whirling Squeeze Film Damper,” STLE Tribol. Trans., 49(2), pp. 271–278. [CrossRef]
De Santiago, O., and San Andrés, L., 1999, “Imbalance Response and Damping Force Coefficients of a Rotor Supported on End Sealed Integral Squeeze Film Dampers,” ASME Paper No. 99-GT-203.
San Andrés, L., Koo, B., and Jeung, S.-H., 2019, “Experimental Force Coefficients for Two Sealed Ends Squeeze Film Dampers (Piston Rings and O-rings): An Assessment of Their Similarities and Differences,” ASME J. Eng. Gas Turbines Power, 141(2), p. 021024. [CrossRef]
Jeung, S.-H., 2017, “Experimental Performance of an Open Ends Squeeze Film Damper and a Sealed Ends Squeeze Film Damper,” Ph.D. dissertation, Texas A&M University, College Station, TX.
San Andrés, L., 2012, “Experimental Identification of Bearing Force Coefficients,” Modern Lubrication Theory, Notes 14, Texas A&M University Digital Libraries, http://oaktrust.library.tamu.edu/handle/1969.1/93254. Accessed June 27, 2016.
Diaz, S., and San Andrés, L., 2001, “A Model for Squeeze Film Dampers Operating with Air Entrainment and Validation with Experiments,” ASME J. Tribol., 123(1), pp. 125–133. [CrossRef]
Diaz, S., and San Andrés, L., 2001, “Air Entrainment Versus Lubricant Vaporization in Squeeze Film Dampers: An Experimental Assessment of Their Fundamental Differences,” ASME J. Eng. Gas Turbines Power, 123(4), pp. 871–877. [CrossRef]
San Andrés, L., and Jeung, S.-H., 2016, “Orbit-Model Force Coefficients for Fluid Film Bearings: A Step Beyond Linearization,” ASME J. Eng. Gas Turbines Power, 138(2), p. 022502. [CrossRef]

Figures

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Fig. 3

PR orientation: (a) circumferential and (b) axial

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Fig. 4

Schematic views: (a) whirl orbit kinematics (exaggerated clearance for illustrative purpose), (b) coordinate systems for motion description, and (c) orbits with amplitude (r) at centered and off-centered conditions (es)

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Fig. 2

(a) Cross section of damper journal and bearing cartridge (L/D = 0.2), (b) photograph of piston ring, and (c) photograph of journal with piston rings installed

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Fig. 1

Photograph and top view of SFD test rig and components

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Fig. 5

Test SFD force coefficients versus orbit amplitude (r/c) at static eccentricity es/c = 0, 0.25, 0.50. Lubricant supply pressure Pin-1 = 0.69 bar. Sealed ends damper (flow conductance Cave-S1 = 0.56 LPM/bar).

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Fig. 6

Test SFD force coefficients versus orbit amplitude (r/c) at static eccentricity es/c = 0, 0.25, 0.50. Lubricant supply pressure Pin-1 = 2.76 bar. Sealed ends damper (flow conductance Cave-S1 = 0.56 LPM/bar).

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Fig. 7

Test SFD direct damping (C)SFD and added mass (M)SFD versus whirl orbit amplitude (r/c) and es/c = 0. Two magnitudes of lubricant supply pressure Pin-1 = 0.69 bar and Pin-2 = 2.76 bar. End seal flow conductance Cave-S1 = 0.56 LPM/bar.

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Fig. 8

Dynamic film pressure (P) and film thickness (h) versus time (t/T) recorded at Θ = 225 deg. Two oil supply conditions, Pin-1 = 0.69 bar and Pin-2 = 2.76 bar. Circular centered orbit with radius r = 0.15 mm and whirl frequency ω = 90 Hz ( = 86 mm/s).

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Fig. 9

Test SFD direct damping (C)SFD and added mass (M)SFD versus static eccentricity (es/c) and whirl orbit amplitude r/c = 0.1. Two magnitudes of lubricant supply pressure Pin-1 = 0.69 bar and Pin-2 = 2.76 bar. End seal flow conductance Cave-S1 = 0.56 LPM/bar.

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Fig. 11

Test damping (C)SFD and added mass (M)SFD force coefficients versus whirl orbit amplitude (r/c) at es/c = 0.0. Damper with seal flow conductances Cave-S1 = 0.56 LPM/bar and Cave-S2 = 0.89 LPM/bar. Lubricant supply pressure Pin-1 = 0.69 bar. Identification frequency range 10–100 Hz.

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Fig. 12

Sealed ends damper: Experimental and predicted SFD direct damping (C) and added mass (M) coefficients versus amplitude (r/c) for circular centered (eS = 0) orbits. Lubricant supply pressure Pin-1 = 0.69 barg and end seal with CS1 = 0.56 LPM/bar. Note reference values corresponding to a perfectly sealed ends SFD w/o feed holes.

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Fig. 10

Photographs of top oil collector showing lubricant exit condition. Lubricant supply pressure: (a) Pin-1 = 0.69 bar and (b) Pin-2 = 2.76 bar. Whirl motion with r = 0.45 c and ω = 80 Hz (squeeze speed = 58 mm/s). Graphs, top: elapsed time t = 0 s; middle: t = 5 s; bottom: t = 15 s.

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Fig. 13

Flow diagram with hydraulic resistances representing a sealed ends SFD

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Fig. 14

Oil flow rate at inlet (Qin) and through bottom section (QB), versus feed hole pressure (PO) for open and sealed ends SFD (c = 0.254 mm). PO estimated based on Eq. (A1). Flow conductance labeled.

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