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Gas Turbines: Structures and Dynamics

Damping and Inertia Coefficients for Two Open Ends Squeeze Film Dampers With a Central Groove: Measurements and Predictions

[+] Author and Article Information
Luis San Andrés

Mast-Childs Professor Mechanical Engineering Department,  Texas A&M University, College Station, TX 77843LSanAndres@tamu.edu

The assertion has various exemplary exceptions (see Refs. [4] and [5]).

A shallow groove is two to five times the film clearance, while a deep groove may be two orders of magnitude in depth compared to the land film clearance.

By definition, SFDs do not have stiffness coefficients, i.e., reaction forces due to changes in static displacement. SFDs develop forces in reaction to journal motions (velocity and acceleration).

The test data shows force coefficients much larger than classical model predictions; thus, the need for a more accurate development. Please see Refs. [5] and [22] for a comprehensive review of the literature, the foundation of the model, and comparisons to archival test data.

J. Eng. Gas Turbines Power 134(10), 102506 (Aug 22, 2012) (9 pages) doi:10.1115/1.4007058 History: Received June 20, 2012; Revised June 20, 2012; Published August 22, 2012; Online August 22, 2012

Aircraft engine rotors are particularly sensitive to rotor imbalance and sudden maneuver loads, since they are always supported on rolling element bearings with little damping. Most engines incorporate squeeze film dampers (SFDs) as means to dissipate mechanical energy from rotor vibrations and to ensure system stability. The paper quantifies experimentally the forced performance of a SFD comprising two parallel film lands separated by a deep central groove. Tests are conducted on two open ends SFDs, both with diameter D = 127 mm and nominal radial clearance c = 0.127 mm. One damper has film lands with length L = 12.7 mm (short length), while the other has 25.4 mm land lengths. The central groove has width L and depth 3/4 L. A light viscosity lubricant flows into the central groove via three orifices, 120 deg apart and then through the film lands to finally exit to ambient. In operation, a static loader pulls the bearing to various eccentric positions and electromagnetic shakers excite the test system with periodic loads to generate whirl orbits of specific amplitudes. A frequency domain method identifies the SFD damping and inertia force coefficients. The long damper generates six times more damping and about three times more added mass than the short length damper. The damping coefficients are sensitive to the static eccentricity (up to ∼ 0.5 c), while showing lesser dependency on the amplitude of whirl motion (up to 0.2 c). On the other hand, inertia coefficients increase mildly with static eccentricity and decrease as the amplitude of whirl motion increases. Cross-coupled force coefficients are insignificant for all imposed operating conditions on either damper. Large dynamic pressures recorded in the central groove demonstrate the groove does not isolate the adjacent squeeze film lands, but contributes to the amplification of the film lands’ reaction forces. Predictions from a novel SFD model that includes flow interactions in the central groove and feed orifices agree well with the test force coefficients for both dampers. The test data and predictions advance current knowledge and demonstrate that SFD-forced performance is tied to the lubricant feed arrangement.

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Copyright © 2012 by American Society of Mechanical Engineers
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References

Figures

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Figure 1

Depiction of a typical squeeze film damper with a central feed groove [3]

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Figure 2

Photograph of SFD test rig and coordinate axes

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Figure 3

Cross-section view of SFD test rig and lubricant flow path through damper film lands

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Figure 4

Schematic views of induced BC whirl motions, centered and off-centered: (a) rectilinear displacements; (b) circular orbits; (c) elliptic orbits, 2:1 amplitude ratio; (d) elliptic orbits, 5:1 amplitude ratio. The dotted line represents the clearance circle.

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Figure 5

Short SFD: direct damping coefficients (C¯XX,C¯YY)SFD versus amplitude (r) of circular orbit. Tests at centered condition (eS  = 0) and two static eccentricities, eS  = 0.29cB and 0.44cB .

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Figure 6

Short SFD: direct added mass coefficients (M¯XX,M¯YY)SFD versus amplitude (r) of circular orbit. Tests conditions stated in Fig. 5.

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Figure 7

Long SFD: direct damping coefficients (C¯XX,C¯YY)SFD versus static eccentricity (eS ). Test data for circular orbits: ΔX = ΔY = 0.09cA , 0.18cA and for elliptical orbits: ΔXY = 2:1 and 5:1 with ΔX = 0.009cA , 0.18 cA .

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Figure 8

Long SFD: direct inertia coefficients (M¯XX,M¯YY)SFD versus static eccentricity (eS ). Test conditions stated in Fig. 7.

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Figure 9

Disposition of dynamic pressure sensors in bearing cartridge. Long damper configuration.

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Figure 10

Long open ends SFD: dynamic pressures in film lands and central groove versus time. Centered bearing (es  = 0), circular orbit r = 0.1cA with frequency 130 Hz. Groove static pressure PG  = 0.72 bar.

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Figure 11

Long open ends SFD: dynamic pressures in film lands and central groove versus time. Centered bearing (eS  = 0), circular orbit r = 0.1cA with frequency 250 Hz. Groove static pressure PG  = 0.72 bar.

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Figure 12

Long open ends SFD: peak-peak dynamic pressures in film lands and central groove versus whirl frequency. Centered bearing es  = 0, circular orbit r = 0.1cA . Groove static pressure PG  = 0.72 bar.

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Figure 13

Short and long open ends SFDs: experimental and predicted damping coefficients (C¯XX,C¯YY)SFD versus static eccentricity (eS /c)

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Figure 14

Short and long open ends SFDs: experimental and predicted inertia coefficients (M¯XX,M¯YY)SFD versus static eccentricity (eS /c)

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