Research Papers: Gas Turbines: Structures and Dynamics

Rotordynamic Performance of a Negative-Swirl Brake for a Tooth-on-Stator Labyrinth Seal

[+] Author and Article Information
Dara W. Childs

Turbomachinery Laboratory,
The Leland T. Jordan Chair
of Mechanical Engineering,
Texas A&M University,
College Station, TX 77843
e-mail: dchilds@tamu.edu

James E. Mclean, Jr., Min Zhang

Turbomachinery Laboratory,
Texas A&M University,
College Station, TX 77843

Stephen P. Arthur

Samsung Techwin,
Houston, TX 77079
e-mail: s.arthur@samsung.com

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received September 28, 2015; final manuscript received October 13, 2015; published online November 24, 2015. Editor: David Wisler.

J. Eng. Gas Turbines Power 138(6), 062505 (Nov 24, 2015) (8 pages) Paper No: GTP-15-1464; doi: 10.1115/1.4031877 History: Received September 28, 2015; Revised October 13, 2015

In the late 1970s, Benckert and Wachter (Technical University Stuttgart) tested labyrinth seals using air as the test media and measured direct and cross-coupled stiffness coefficients. They reported the following results: (1) fluid preswirl in the direction of shaft rotation creates destabilizing cross-coupled stiffness coefficients and (2) effective swirl brakes at the inlet to the seal can markedly reduce the cross-coupled stiffness coefficients, in many cases reducing them to zero. In recent years, “negative-swirl” swirl brakes have been employed, which attempt to reverse the circumferential direction of inlet flow, changing the sign of the cross-coupled stiffness coefficients and creating stabilizing stiffness forces. This study presents test results for a 16-tooth labyrinth seal with positive inlet preswirl (in the direction of shaft rotation) for the following inlet conditions: (1) no swirl brakes, (2) straight, conventional swirl brakes, and (3) negative-swirl swirl brakes. The negative-swirl swirl-brake designs were developed based on computational fluid dynamics (CFD) predictions. Tests were conducted at 10.2, 15.35, and 20.2 krpm with 70 bar of inlet pressure for pressure ratios of 0.3, 0.4, and 0.5. Test results include leakage and rotordynamic coefficients. In terms of leakage, the negative-swirl brake configuration leaked the least, followed by the conventional brake, followed by the no-brake design. Normalized to the negative-swirl brake configuration, the conventional-brake and no-brake configurations mass flow rate was greater, respectively, by factors of 1.04 and 1.09. The direct-stiffness coefficients are negative but small, consistent with past experience. The conventional swirl brake drops the destabilizing cross-coupled stiffness coefficients k by a factor of about 0.8 as compared to the no-brake results. The negative-swirl brake produces a change in sign of k with an appreciable magnitude; hence, the stability of forward precessing modes would be enhanced. In descending order, the direct-damping coefficients C are: no-swirl, negative-swirl, and conventional-swirl. Normalized in terms of the no-swirl case, C for the negative and conventional brake designs is, respectively, 0.7 and 0.6 smaller. The effective damping Ceff combines the effect of k and C. Ceff is large and positive for the negative-swirl configuration and near zero for the no-brake and conventional-brake designs. The present results for a negative-brake design are very encouraging for both eye-packing seals (where conventional swirl brakes have been previously employed) and division-wall and balance-piston seals, where negative shunt injection has been employed.

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Fig. 1

Swirl brakes in a centrifugal compressor [3]

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Fig. 2

Shunt injection in a labyrinth balance-piston seal [4]

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Fig. 3

Steam turbine seal with reverse-swirl vanes [6]

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Fig. 4

Computational grid

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Fig. 5

Calculated stream lines

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Fig. 6

A 16-tooth, TOS test labyrinth seal

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Fig. 7

Conventional swirl brake

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Fig. 8

Negative-swirl brake

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Fig. 9

Axial seal and swirl brake fit

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Fig. 10

Cross section of the air seal test rig [2]

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Fig. 11

Test stator assembly [2]

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Fig. 12

Cross section view of the preswirl rings and pitot tube location [2]

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Fig. 13

Preswirl ring and preswirl measurement [2]

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Fig. 14

Mass flow rate with the three swirl brake arrangements

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Fig. 15

Preswirl ratio for the no-brake configuration versus Pr for Ps = 70 bar

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Fig. 16

Real parts of Hxx, Hyy, Hyx, and Hxy at 20.2 krpm, PR = 0.5, and no swirl brakes

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Fig. 17

Imaginary parts of Hxx, Hyy, Hyx, and Hxy at 20.2 krpm, PR = 0.5, and no swirl brakes

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Fig. 18

K versus PR at 20.2 krpm

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Fig. 19

k versus PR at 20.2 krpm

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Fig. 20

c versus PR at 20.2 krpm

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Fig. 21

C versus PR at 20.2 krpm

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Fig. 22

Ceff versus PR at 20.2 krpm




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