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Research Papers: Gas Turbines: Structures and Dynamics

Numerical Investigations on the Leakage and Rotordynamic Characteristics of Pocket Damper Seals—Part I: Effects of Pressure Ratio, Rotational Speed, and Inlet Preswirl

[+] Author and Article Information
Zhigang Li, Zhenping Feng

Institute of Turbomachinery,
School of Energy & Power Engineering,
Xi'an Jiaotong University,
Xi'an 710049, China

Jun Li

Institute of Turbomachinery,
School of Energy & Power Engineering,
Xi'an Jiaotong University,
Xi'an 710049, China
Collaborative Innovation Center of
Advanced Aero-Engine,
Beijing 100191, China
e-mail: junli@mail.xjtu.edu.cn

1Corresponding author.

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 11, 2014; final manuscript received July 17, 2014; published online September 30, 2014. Editor: David Wisler.

J. Eng. Gas Turbines Power 137(3), 032503 (Sep 30, 2014) (15 pages) Paper No: GTP-14-1373; doi: 10.1115/1.4028373 History: Received July 11, 2014; Revised July 17, 2014

Effects of pressure ratio, rotational speed and inlet preswirl on the leakage and rotordynamic characteristics of a eight-bladed fully partitioned pocket damper seal (FPDS) were numerically investigated using proposed three-dimensional (3D) transient computational fluid dynamics (CFD) methods based on the multifrequency elliptical whirling orbit model. The accuracy and availability of the multifrequency elliptical whirling orbit model and the transient CFD numerical methods were demonstrated with the experimental data of frequency-dependent rotordynamic coefficients of the FPDS at two rotational speeds with high preswirl conditions. The frequency-dependent rotordynamic coefficients of the FPDS at three pressure ratios (three inlet pressures and three outlet pressures), three rotational speeds, three inlet preswirls were computed. The numerical results show that changes in outlet pressure have only weak effects on most rotordynamic coefficients. The direct damping and effective damping slightly increase in magnitude with decreasing outlet pressure at the frequency range of 20–200 Hz. The effect of inlet pressure is most prominent, and increasing inlet pressure for the FPDS results in a significant increase in the magnitudes of all rotordynamic coefficients. The magnitudes of the seal response force and effective damping are proportional to pressure drop through the seal. Increasing rotational speed and increasing inlet preswirl velocity both result in a significant decrease in the effective damping term due to the obvious increase in the magnitude of the destabilizing cross-coupling stiffness with increasing rotational speed or increasing preswirl velocity. The crossover frequency of effective damping significantly increases and the peak magnitude of effective damping decreases with increasing rotational speed or increasing preswirl velocity. The destabilizing cross-coupling stiffness is mainly caused by the circumferential swirl velocity generating from high rotational speed and inlet preswirl. Reducing swirl velocity (such as swirl brake) can greatly enhance the stabilizing capacity of the FPDS.

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Figures

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Fig. 1

Geometry of experimental FPDS [20]

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Fig. 2

Computational model and mesh of the FPDS

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Fig. 3

Elliptical orbit whirling model for the rotor vibration with a single frequency

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Fig. 4

Multiple frequencies elliptical whirling orbit of the rotor

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Fig. 5

Dynamic monitoring data: (a) rotor motion and (b) response force (x excitation)

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Fig. 6

Rotordynamic coefficients versus vibration frequency with inlet preswirl velocity (u0 = 60 m/s): (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 7

Seal leakage flow rate versus pressure ratio: (a) with decreasing outlet pressure and (b) with increasing inlet pressure

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Fig. 8

Rotordynamic coefficients versus vibration frequency at different pressure ratios (decreasing outlet pressure): (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 9

Rotordynamic coefficients versus vibration frequency at different pressure ratios (increasing inlet pressure): (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 10

Seal leakage flow rate versus rotational speed

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Fig. 11

Rotordynamic coefficients versus vibration frequency at different rotational speeds without inlet preswirl: (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 12

Rotordynamic coefficients versus vibration frequency at different rotational speeds with inlet preswirl: (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 13

Static pressure contours on the cross section through the middle of cavity 3 and phasor diagram of the response force at different rotational speeds (u0 = 0 m/s x excitation, T=0.1 s): (a) 0 rpm, (b) 7000 rpm, and (c) 15,000 rpm

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Fig. 14

Seal leakage flow rate versus inlet preswirl velocity

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Fig. 15

Rotordynamic coefficients versus vibration frequency at different preswirl velocities (0 rpm): (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 16

Rotordynamic coefficients versus vibration frequency at different preswirl velocities (7000 rpm): (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 17

Rotordynamic coefficients versus vibration frequency at different preswirl velocities (15,000 rpm): (a) direct stiffness, (b) direct damping, (c) cross-coupling stiffness, and (d) effective damping

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Fig. 18

Static pressure contours on the cross section through the middle of cavity 3 and phasor diagram of the response force at different inlet preswirl velocities (n = 7000 rpm, x excitation, T = 0.1 s): (a) u0 = 0 m/s, (b) u0 = 30 m/s, and (c) u0 = 60 m/s

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