Research Papers

Experimental Force Coefficients for Two Sealed Ends Squeeze Film Dampers (Piston Rings and O-Rings): An Assessment of Their Similarities and Differences

[+] Author and Article Information
Luis San Andrés

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

Bonjin Koo

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

Sung-Hwa Jeung

Compressor Tech. & Development,
Ingersoll Rand,
La Crosse, WI 54601
e-mail: Sean.jeung@gmail.com

1Corresponding author.

2Present address: Mechanical Engineering Department, Texas A&M University, College Station, TX 77843.

Manuscript received July 5, 2018; final manuscript received July 6, 2018; published online October 4, 2018. Editor: Jerzy T. Sawicki.

J. Eng. Gas Turbines Power 141(2), 021024 (Oct 04, 2018) (13 pages) Paper No: GTP-18-1451; doi: 10.1115/1.4040902 History: Received July 05, 2018; Revised July 06, 2018

Squeeze film dampers (SFDs) in aircraft engines effectively aid to reduce rotor motion amplitudes, in particular when traversing a critical speed, and help to alleviate rotor whirl instabilities. The current work is a long-term endeavor focused on quantifying the dynamic force performance of practical SFDs, exploring novel design damper configurations, and producing physically sound predictive SFD models validated by experimental data. Piston rings (PRs) and O-rings (ORs), commonly used as end seals in SFDs for commercial and military gas turbine engines, respectively, amplify viscous damping in a short physical length and while operating with a modicum of lubricant flow. This paper presents experimental force coefficients (damping and inertia) for two identical geometry SFDs with end seals, one configuration hosts PRs, and the other one ORs. The test rig comprises a stationary journal and bearing cartridge (BC) hosting the SFD and supported on four elastic rods to emulate a squirrel cage. The damper film land length, diameter, and clearance are L = 25.4 mm, D = 5L, and c = 0.373 mm (D/c = 340), respectively. A supply feeds ISO VG 2 oil to the film land at its middle plane through either one hole or three holes, 2.5 mm in diameter, 120 deg apart. In the PRSFD, the lubricant exits through the slit opening at the ring butted ends. The ORs suppress oil leakage; hence, lubricant evacuates through a 1 mm hole at ¼ L near one journal end. The ORs when installed add significant stiffness and damping to the test structure. The ORSFD produces 20% more damping than the PRSFD, whereas both sealed ends SFDs show similar size added mass. For oil supplied at 0.69 bar(g) through a single orifice produces larger damping, 60–80% more than when the damper operates with three oil feedholes. A computational model reproducing the test conditions delivers force coefficients in agreement with the test data. Archival literature calls for measurement of a single pressure signal to estimate SFD reaction forces. For circular centered orbits (CCOs), the dynamic pressure field, in the absence of any geometrical asymmetry or feed/discharge oil condition, “rotates” around the bearing with a speed equal to the whirl frequency. The paper presents force coefficients estimated from (a) measurements of the applied forces and ensuing displacements, and (b) the dynamic pressure recorded at a fixed angular location and “integrated” over the journal surface. The first method delivers a damping coefficient that is large even with lubricant supplied at a low oil supply pressure whereas the inertia coefficient increases steadily with feed pressure. Predictions show good agreement with the test results from measured forces and displacements, in particular the added mass. On the other hand, identified damping and inertia coefficients from dynamic pressures show a marked difference from one pressure sensor to another, and vastly disagreeing with test results from the first method or predictions. The rationale for the discrepancy relies on local distortions in the dynamic pressure fields that show zones of oil vapor cavitation at a near zero absolute pressure and/or with air ingestion producing high frequency spikes from bubble collapsing; both phenomena depend on the magnitude of the oil supply pressure. An increase in lubricant supply pressure suppresses both oil vapor cavitation and air ingestion, which produces an increase of both damping and inertia force coefficients. No prior art compares the performance of a PRSFD vis-à-vis that of an ORSFD. Supplying lubricant with a large enough pressure (flow rate) is crucial to avoid the pervasiveness of air ingestion. Last, the discussion on force coefficients obtained from two distinct methods questions the use of an oversimplifying assumption; the dynamic pressure field is not invariant in a rotating coordinate frame.

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Fig. 1

Squeeze film damper test rig: (top) photograph with shakers and static loader [6], and (bottom) schematic cross section view of film land at midplane (z = 0). Coordinate system noted.

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Fig. 2

Schematic views of lubricant flow path through a film land for sealed ends damper with (a) PRs (slit location θslit = 345 deg) and (b) ORs with discharge hole at θexit = 240 deg and z = ¼ L

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Fig. 3

Schematic view of squeeze film with journal describing a CCO with amplitude r and frequency ω. Fixed (X, Y) and rotating coordinates (r, t) systems noted.

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Fig. 4

Representation of applied forces on BC (mass) and force coefficients for squeeze film, structure, and ORs

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Fig. 7

(a) O-ring-squeeze film damper and (b) PRSFD: experimental and predicted direct damping (CXX, CYY)SFD and added mass (MXX, MYY)SFD coefficients versus orbit amplitude (r/c) at es = 0c and 0.25c. Whirl frequency range: 10–250 Hz for r/c = 0.01–0.05 and 10–100 Hz for r/c = 0.1–0.15. Three open feedholes. Oil supply pressure Ps = 69 kPa(g).

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Fig. 6

Lubricated test system, PRSFD and ORSFD: real and imaginary parts of complex stiffnesses (HXX, HYY)L versus whirl frequency (ω). CCO with amplitude r = 0.15c and frequency ω =10–100 Hz. Lubricant supplied via three feedholes at pressure Ps=69 kPa (g). PR slit at θ = 345 deg and a discharge hole for OR sealed ends damper at θ = 240 deg and z = ¼L.

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Fig. 8

(a) O-ring-squeeze film damper and (b) PRSFD supplied with oil through 3-feedholes and 1-feedhole: Experimental and predicted direct damping (CXX,CYY)SFD and added mass (MXX,MYY)SFD coefficients versus orbit amplitude (r/c) at es = 0c. Whirl frequency range: 10–250 Hz for r/c = 0.01–0.05, and 10–100 Hz for r/c = 0.1–0.15. Lubricant supply pressure Ps = 69 kPa(g).

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Fig. 5

Dry test system (without lubricant): real and imaginary parts of system complex stiffnesses (HXX, HYY) versus excitation frequency. Orbit amplitude r = 0.03c. Test data and model fits over frequency range 10–210 Hz.

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Fig. 9

Piston ring-squeeze film damper: predicted film pressure at middle land plane versus film circumferential coordinate at two instants, when journal locates at θ = 0 deg and θ = 180 deg. Operation with supply pressure Ps = 1 bar(g). Orbit radius r = 0.15c (es = 0) and frequency ω = 60 Hz. One lubricant feedhole at θ = 45 deg and PR slit at θ = 135 deg. vs = 21 mm/s. Res ∼ 16.

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Fig. 10

Test identified damping (C)F,P and added mass (M)F,P coefficients versus oil supply pressure. PRSFD performs CCOs with radius r = 0.65c and frequency ω = 40 and 60 Hz. Coefficients derived from measured forces (F) and measured pressures (P), A: θ = 225 deg and B: 315 deg, and predictions from orbit model. Squeeze film velocity vs = 61 and 91 mm/s. Res ∼ 10 and 16. PR slit at θ = 135 deg.

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Fig. 11

Piston ring-squeeze film damper: Recorded film dynamic pressure Pθ = 225 deg, Pθ = 315 deg and film thickness at θ = 225 deg and z = 0 versus time. Operation with lubricant supply pressure Ps (a) 0.69 bar(g), (b) 2.8 bar(g), and (c) 6.2 bar(g). Orbit radius r = 0.65c (es = 0) and frequency ω = 60 Hz (and 40 Hz for 2.8 bar). One lubricant feedhole at θ = 45 deg and PR slit at θ = 135 deg.

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Fig. 12

Photographs of top exit plane of PRSFD operating at frequency ω = 60 Hz and orbit radius r = 0.243 mm (r = 0.65c) and es = 0. Two supply pressure conditions: (a) Ps = 0.69 bar (Qin = 0.83 LPM) and (b) Ps = 6.21 bar (Qin = 2.88 LPM). One lubricant feedhole.



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