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Research Papers: Other (Seals, Manufacturing)

Numerical and Experimental Analyses of Static and Dynamic Characteristics for Partially Helically Grooved Liquid Annular Seals

[+] Author and Article Information
Keisuke Nagai

Graduate School of
Nagaoka University of Technology,
Department of Energy and
Environment Science,
Kamitomioka-machi 1603-1,
Nagaoka-shi 940-2188, Niigata, Japan
e-mail: s091057@stn.nagaokaut.ac.jp

Kazuki Koiso

Graduate School of
Nagaoka University of Technology,
Department of Mechanical Engineering,
Kamitomioka-machi 1603-1,
Nagaoka-shi 940-2188, Niigata, Japan
e-mail: 143031@stn.nagaokaut.ac.jp

Satoru Kaneko

Nagaoka University of Technology,
Department of Mechanical Engineering,
Kamitomioka-machi 1603-1,
Nagaoka-shi 940-2188, Niigata, Japan
e-mail: kaneko@mech.nagaokaut.ac.jp

Hiroo Taura

Nagaoka University of Technology,
Department of Mechanical Engineering,
Kamitomioka-machi 1603-1,
Nagaoka-shi 940-2188, Niigata, Japan
e-mail: htaura@vos.nagaokaut.ac.jp

Yusuke Watanabe

EBARA Corporation,
Honfujisawa 4-2-1,
Fujisawa-shi 251-8502, Kanagawa, Japan
e-mail: watanabe.yusuke@ebara.com

Contributed by the Tribology Division of ASME for publication in the JOURNAL OF TRIBOLOGY. Manuscript received March 8, 2018; final manuscript received June 12, 2018; published online September 17, 2018. Assoc. Editor: Alan Palazzolo.

J. Tribol 141(2), 022201 (Sep 17, 2018) (12 pages) Paper No: TRIB-18-1108; doi: 10.1115/1.4040574 History: Received March 08, 2018; Revised June 12, 2018

Numerical and experimental analyses of the static and dynamic characteristics of the liquid annular seals with axially partial helical grooves were conducted to investigate the effects of the axial length gal of a helically grooved section in a seal stator. The numerical solution and experimental procedures were applied in the same manner as in previous studies on through-helically grooved seals, wherein the grooves extend across the seal length. The numerical results qualitatively agreed with the experimental results, demonstrating the validity of the numerical analysis. The leakage flow rate Q was lower in the partially helically grooved seals than that of conventional through-helically grooved seals across a small range of rotor spinning velocities. In contrast, the reduction in Q due to the pumping effect caused by the spin of the rotor diminished with the decrease in gal. For a small concentric whirling motion of the rotor, the radial dynamic reaction force Fr and magnitude of variation in the tangential dynamic reaction force Ft with the whirling angular velocity increased with the decrease in gal, and their values approached the corresponding values for the smooth-surface seal. Under the same rotor whirling velocity, the Ft for the partially helically grooved seals became lower than that for the smooth-surface seal (similar to the case for the through-helically grooved seal), although decreasing gal tended to increase Ft. These results suggest that partially helically grooved seals can improve the efficiency and stability margin of the pumps because of the reduction in leakage flow rate and suppression of the rotor forward whirling motion (with large radial and tangential dynamic reaction forces) as compared to conventional through-helically grooved seals.

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References

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Figures

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Fig. 2

Inner surface of seal stator with axially partial helical grooves

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Fig. 1

Liquid annular seal with axially partial helical grooves in the seal stator and coordinate system

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Fig. 3

Small whirling motion of rotor about the center of the seal and dynamic fluid-film forces

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Fig. 4

Relationship between leakage flow rate Q and rotor spinning velocity N for all seals tested; ε0 = 0: (a) pd = 1000 kPa and (b) pd = 294 kPa

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Fig. 5

Dynamic force coefficients versus ratio of whirling angular velocity Ω to spinning angular velocity ω at the center of the seal; pd = 1000 kPa, N = 3000 rpm: (a) tangential force coefficient Ft/ew and (b) radial force coefficient Fr/ew

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Fig. 7

Stiffness coefficients versus rotor spinning velocity N at the center of the seal; pd = 1000 kPa: (a) main stiffness Km and (b) cross‐coupled stiffness Kc

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Fig. 8

Damping coefficients versus N at the center of the seal; pd = 1000 kPa: (a) main damping Cm and (b) cross‐coupled damping Cc

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Fig. 6

Axial distributions of dynamic force coefficients per unit seal length at the center of the seal; pd = 1000 kPa, Ω/ω = −0.4 (N = 3000 rpm, Nw = −1200 rpm): (a) tangential dynamic force coefficient per unit seal length F¯t/ew and (b) radial dynamic force coefficient per unit seal length F¯r/ew

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Fig. 9

Whirl-frequency ratio versus N at the center of the seal; pd = 1000 kPa

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Fig. 10

Effective damping coefficient Ceff versus N at the center of the seal; pd = 1000 kPa

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Fig. 11

Numerical results for the effect of the clearance groove on the leakage flow rate Q of the partially helically grooved seals; ε0 = 0, pd = 1000 kPa, α = 0.1

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Fig. 12

Numerical results for the effect of the clearance groove on the dynamic characteristics of the partially helically grooved seals; ε0 = 0, pd = 1000 kPa, α = 0.1: (a) main stiffness Km, (b) cross‐coupled stiffness Kc, (c) main damping Cm, and (d) WFR

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