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Research Papers

Improvement of Steam Turbine Stage Efficiency by Controlling Rotor Shroud Leakage Flows—Part II: Effect of Axial Distance Between a Swirl Breaker and a Rotor Shroud on Efficiency Improvement

[+] Author and Article Information
Chongfei Duan

Research & Innovation Center,
Mitsubishi Heavy Industries, Ltd.,
1-1 Shinhama, 2-chome, Arai-cho,
Takasago City 676-8686, Hyogo, Japan
e-mail: chongfei_duan@mhi.co.jp

Hisataka Fukushima

Turbomachinery Headquarters,
Mitsubishi Hitachi Power Systems, Ltd.,
2-1-1 Shinhama, Arai-cho,
Takasago City 676-8686, Hyogo, Japan
e-mail: hisataka_fukushima@mhps.com

Kiyoshi Segewa

Turbomachinery Headquarters,
Mitsubishi Hitachi Power Systems, Ltd.,
1-1, Saiwai-cho 3-chome,
Hitachi City 317-8585, Ibaraki, Japan
e-mail: kiyoshi_segawa@mhps.com

Takanori Shibata

Research & Innovation Center,
Mitsubishi Heavy Industries, Ltd.,
1-1 Shinhama, 2-chome, Arai-cho,
Takasago City 676-8686, Hyogo, Japan
e-mail: takanori_shibata@mhi.co.jp

Hidetoshi Fujii

Turbomachinery Headquarters,
Mitsubishi Hitachi Power Systems, Ltd.,
3-1, Minatomirai 3-chome, Nishi-ku,
Yokohama 220-8401, Kanagawa, Japan
e-mail: hidetoshi1_fujii@mhps.com

1Corresponding author.

Manuscript received August 10, 2018; final manuscript received September 9, 2018; published online November 1, 2018. Editor: Jerzy T. Sawicki.

J. Eng. Gas Turbines Power 141(4), 041002 (Nov 01, 2018) (9 pages) Paper No: GTP-18-1560; doi: 10.1115/1.4041648 History: Received August 10, 2018; Revised September 09, 2018

The basic principle of a distinct idea to reduce an aerodynamic mixing loss induced by the difference in tangential velocity between mainstream flow and rotor shroud leakage flow is presented in “Part I: Design Concept and Typical Performance of a Swirl Breaker.” When the swirl breaker is installed in the circulating region of leakage flow at the rotor shroud exit cavity, the axial distance between the swirl breaker and the rotor shroud is a crucial factor to trap the leakage flow into the swirl breaker cavity. In Part II, five cases of geometry with different axial distances between the swirl breaker and the rotor shroud, which covered a range for the stage axial distance of actual high and intermediate pressure (HIP) steam turbines, were investigated using a single-rotor computational fluid dynamics (CFD) analysis and verification tests in a 1.5-stage air model turbine. By decreasing the axial distance between the swirl breaker and the rotor shroud, the tangential velocity and the mixing region in the tip side which is influenced by the rotor shroud leakage flow were decreased and the stage efficiency was increased. The case of the shortest axial distance between the swirl breaker and the rotor shroud increased turbine stage efficiency by 0.7% compared to the conventional cavity geometry. In addition, the measured maximum pressure fluctuation in the swirl breaker cavity was only 0.7% of the entire flow pressure. Consequently, both performance characteristics and structural reliability of swirl breaker were verified for application to real steam turbines.

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References

Tanuma, T. , 2017, “ Introduction to Steam Turbines for Power Plants,” Advances in Steam Turbines for Modern Power Plants, Elsevier, Amsterdam, The Netherlands, Chap. 1.
Shibata, T. , Fukushima, H. , and Segawa, K. , 2018, “ Improvement of Steam Turbine Stage Efficiency by Controlling Rotor Shroud Leakage Flows—Part I: Design Concept and Typical Performance of a Swirl Breaker,” ASME J. Eng. Gas Turbines Power (accepted).
Denton, J. D. , 1993, “ Loss Mechanisms in Turbomachines,” ASME J. Turbomach., 115(4), pp. 621–656. [CrossRef]
Wallis, A. M. , Denton, J. D. , and Demargne, A. J. , 2001, “ The Control of Shroud Leakage Flows to Reduce Aerodynamic Losses in a Low Aspect Ratio,” ASME J. Turbomach., 123(2), pp. 334–341. [CrossRef]
Rosic, B. , and Denton, J. D. , 2006, “ The Control of Shroud Leakage Loss by Reducing Circumferential Mixing,” ASME Paper No. GT2006-90946.
Rosic, B. , Denton, J. D. , Curtis, E. M. , and Perterson, A. T. , 2007, “ The Influence of Shroud and Cavity Geometry on Turbine Performance—An Experimental and Computational Study—Part 2: Exit Cavity Geometry,” ASME Paper No. GT2007-27770.
Barmpalias, K. G. , Kalfas, A. I. , Abhari, R. S. , Hirano, T. , Shibukawa, N. , and Sasaki, T. , 2011, “ Design Considerations for Axial Steam Turbine Rotor Inlet Cavity Volume and Length Scale,” ASME Paper No. GT2011-45127.
Segawa, K. , Shikano, Y. , Tsubouchi, K. , and Shibashita, N. , 2001, “ Performance Verification of a Highly Loaded Steam Turbine Blade,” ASME Proceedings of the International Joint Power Generation Conference, Vol. 2, pp. 323–332.
Ozaki, S. , Yamashita, Y. , and Segawa, K. , 2012, “ Experimental and Numerical Investigations of the Influences of Axial Gap Between Blade Rows on Pressure Fluctuation,” 13th International Symposium on Unsteady Aerodynamics, Aeroacoustics and Aeroelasticity of Turbomachines, Tokyo, Japan, Sept. 11–14, Paper No. ISUAAAT13-S9-4.
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Figures

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Fig. 1

Shroud cavity geometries [2]: (a) base case and (b) swirl breaker

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Fig. 2

Typical calculated flow pattern contours and streamlines of single-stage CFD results (U/C0 = 0.56) (partly modified figure based on results of Part I [2]): (a) base case, (b) swirl breaker (G/Cx = 0.08), and (c) swirl breaker (G/Cx = 0.45)

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Fig. 3

Meridional shape (base shroud cavity case)

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Fig. 4

Shroud cavity geometry of swirl breaker (case 2)

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Fig. 5

Computational domains for single-stage CFD (base case)

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Fig. 6

Comparison between the EFD and single-stage CFD in the tip region: (a) rotor loss coefficient ξ/ξref,mid (-) and (b) absolute outlet flow angle (deg)

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Fig. 7

The definition of the absolute outlet flow angle

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Fig. 8

Typical calculated turbulent kinetic energy (U/C0 = 0.56): (a) single-stage CFD for base case, (b) single-stage CFD for case 1, (c) single-rotor CFD for base case, and (d) single-rotor CFD for case 1

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Fig. 9

Comparison between the EFD and single-rotor CFD in the tip region: (a) rotor loss coefficient ξ/ξref,mid (-) and (b) absolute outlet flow angle (deg)

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Fig. 10

Schematic diagram of air turbine test facility

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Fig. 11

Cross section of 1.5-stage air model turbine

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Fig. 12

Photograph for the inside of the air model turbine

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Fig. 13

Photograph of 5-hole pneumatic probe tip for traversing at the rotor outlet

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Fig. 14

Unsteady pressure transducers and installation place: (a) pressure transducer and (b) pressure transducer

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Fig. 15

Single-rotor CFD prediction flow distribution (U/C0 = 0.56): (a) stage efficiency η/ηref,mid (%), (b) difference of efficiency (%), (c) absolute tangential velocity (m/s), and (d) absolute outlet flow angle (deg)

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Fig. 16

Improvement of stage efficiency (U/C0 = 0.56)

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Fig. 17

Typical measured flow distribution at the rotor outlet for base case, case 1 and case 2 (U/C0 = 0.56): (a) stage efficiency η/ηref,mid (%), (b) difference of efficiency (%), (c) absolute tangential velocity (m/s), and (d) absolute outlet flow angle (deg)

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Fig. 18

Typical measured flow distribution at the rotor outlet for base case, case 4 and case 5 (U/C0 = 0.56): (a) stage efficiency η/ηref,mid (%) and (b) absolute outlet flow angle (deg)

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Fig. 19

Typical standard deviation results of pressure fluctuations in the swirl breaker cavity

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