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Research Papers: Gas Turbines: Structures and Dynamics

Numerical Investigation on the Leakage and Static Stability Characteristics of Pocket Damper Seals at High Eccentricity Ratios

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

Institute of Turbomachinery,
School of Energy and Power Engineering,
Xi’an Jiaotong University,
Xi’an 710049, China

Jun Li

Institute of Turbomachinery,
School of Energy and 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 23, 2017; final manuscript received August 1, 2017; published online November 7, 2017. Editor: David Wisler.

J. Eng. Gas Turbines Power 140(4), 042503 (Nov 07, 2017) (16 pages) Paper No: GTP-17-1388; doi: 10.1115/1.4038081 History: Received July 23, 2017; Revised August 01, 2017

Annular gas seals for compressors and turbines are designed to operate in a nominally centered position in which the rotor and stator are at concentric condition, but due to the rotor–stator misalignment or flexible rotor deflection, many seals usually are suffering from high eccentricity. The centering force (represented by static stiffness) of an annular gas seal at eccentricity plays a pronounced effect on the rotordynamic and static stability behavior of rotating machines. The paper deals with the leakage and static stability behavior of a fully partitioned pocket damper seal (FPDS) at high eccentricity ratios. The present work introduces a novel mesh generation method for the full 360 deg mesh of annular gas seals with eccentric rotor, based on the mesh deformation technique. The leakage flow rates, static fluid-induced response forces, and static stiffness coefficients were solved for the FPDS at high eccentricity ratios, using the steady Reynolds-averaged Navier–Stokes solution approach. The calculations were performed at typical operating conditions including seven rotor eccentricity ratios up to 0.9 for four rotational speeds (0 rpm, 7000 rpm, 11,000 rpm, and 15,000 rpm) including the nonrotating condition, three pressure ratios (0.17, 0.35, and 0.50) including the choked exit flow condition, two inlet preswirl velocities (0 m/s, 60 m/s). The numerical method was validated by comparisons to the experiment data of static stiffness coefficients at choked exit flow conditions. The static direct and cross-coupling stiffness coefficients are in reasonable agreement with the experiment data. An interesting observation stemming from these numerical results is that the FPDS has a positive direct stiffness as long as it operates at subsonic exit flow conditions; no matter the eccentricity ratio and rotational speed are high or low. For the choked exit condition, the FPDS shows negative direct stiffness at low eccentricity ratio and then crosses over to positive value at the crossover eccentricity ratio (0.5–0.7) following a trend indicative of a parabola. Therefore, the negative static direct stiffness is limited to the specific operating conditions: choked exit flow condition and low eccentricity ratio less than the crossover eccentricity ratio, where the pocket damper seal (PDS) would be statically unstable.

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Figures

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Fig. 1

Twelve-bladed pocket damper seal: (a) conversational PDS and (b) fully-partitioned PDS

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Fig. 2

A eight-bladed, eight-pocket, fully partitioned pocket damper seal [11]

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Fig. 3

Computational models of the fully partitioned pocket damper seal

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Fig. 4

Flow charts for generation of eccentric seal mesh and computation of static stiffness coefficients

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Fig. 5

Computational meshes of the fully partitioned pocket damper seal (7.56 million nodes)

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Fig. 6

Leakage flow rate and fluid-induced response force versus mesh nodes number (ε = 0.5, Pout = 1.0 bar, n = 15,000 rpm, u0 = 0 m/s)

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Fig. 7

Static direct stiffness and cross-coupling stiffness coefficients versus rotational speed (ε = 0, Pout = 1.0 bar, u0 = 0 m/s)

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Fig. 8

Leakage flow rates versus rotor eccentricity ratios for various pressure ratios (n = 0 rpm)

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Fig. 9

Leakage flow rates versus rotor eccentricity ratios for various pressure ratios, rotational speeds, and preswirl velocities: (a) n = 7000 rpm, (b) n = 11,000 rpm, and (c) n = 15,000 rpm

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Fig. 10

Mach number distribution in the tip gap of the last seal blade at different pressure ratios (the flow is from left to right, ε = 0, n = 15,000 rpm, u0 = 0 m/s)

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Fig. 11

Static cross-coupling response forces versus rotor eccentricity ratios for various pressure ratios (n = 0 rpm)

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Fig. 12

Static cross-coupling response forces versus rotor eccentricity ratios for various pressure ratios, rotational speeds, and preswirl velocities: (a) n = 7000 rpm, (b) n = 11,000 rpm, and (c) n = 15,000 rpm

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Fig. 13

Static direct response forces versus rotor eccentricity ratios for various pressure ratios (n = 0 rpm)

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Fig. 14

Static direct response forces versus rotor eccentricity ratios for various pressure ratios, rotational speeds and preswirl velocities: (a) n = 7000 rpm, (b) n = 11,000 rpm, and (c) n = 15,000 rpm

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Fig. 15

Static pressure contours on the cross section through the middle of cavity 7 and phasor diagram of the response force for different rotor eccentricities (left: choked exit flow, π = 0.17; right: unchoked exit flow, π = 0.50; n = 15,000 rpm, u0 = 60 m/s): (a) ε = 0.3, (b) ε = 0.5, (c) ε = 0.9

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Fig. 16

Relative static pressure of cavity 7 distributions in the circumferential direction for different rotor eccentricities (left: choked exit flow, π = 0.17; right: unchoked exit flow, π = 0.50; n = 15,000 rpm, u0 = 60 m/s)

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Fig. 17

Pressure and Mach number variations in the sealing gap along the seal from inlet to exit for (a) a unchoked exit flow and (b) a choked exit flow (n = 15,000 rpm, u0 = 60 m/s): (a) π = 0.50, unchoked exit flow and (b) π = 0.17, choked exit flow

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Fig. 18

Static cross-coupling stiffness versus static eccentricity ratios for various pressure ratios (n = 0 rpm)

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Fig. 19

Static cross-coupling stiffness versus static eccentricity ratios for various pressure ratios, rotational speeds and preswirl velocities: (a) n = 7000 rpm, (b) n = 11,000 rpm, and (c) n = 15,000 rpm

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Fig. 20

Cross-coupling stiffness representation of follower force on a forward rotor mode

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Fig. 21

Static direct stiffness versus static eccentricity ratios for various pressure ratios (n = 0 rpm)

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Fig. 22

Static direct stiffness versus static eccentricity ratios for various pressure ratios, rotational speeds and preswirl velocities: (a) n = 7000 rpm, (b) n = 11,000 rpm, and (c) n = 15,000 rpm

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