Research Papers: Gas Turbines: Combustion, Fuels, and Emissions

A Novel Damping Device for Broadband Attenuation of Low-Frequency Combustion Pulsations in Gas Turbines

[+] Author and Article Information
Mirko R. Bothien

e-mail: mirko.bothien@power.alstom.com

Bruno Schuermans

5401 Baden, Switzerland

The acoustic impedance Z is related to the reflection coefficient R by the bilinear transform Z/ρc=(1+R)/(1-R).

1Corresponding author.

Contributed by the Combustion and Fuels Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received August 16, 2013; final manuscript received August 26, 2013; published online December 10, 2013. Editor: David Wisler.

J. Eng. Gas Turbines Power 136(4), 041504 (Dec 10, 2013) (9 pages) Paper No: GTP-13-1310; doi: 10.1115/1.4025761 History: Received August 16, 2013; Revised August 26, 2013

Damping of thermoacoustically induced pressure pulsations in combustion chambers is a major focus of gas turbine operation. Conventional Helmholtz resonators are an excellent means to attenuate thermoacoustic instabilities in gas turbines. Usually, however, the damping optimum is in a narrow frequency band at one operating condition. The work presented here deals with a modification of the basic Helmholtz resonator design overcoming this drawback. It consists of a damper body housing multiple volumes that are connected to each other. Adequate adjustment of the governing parameters results in a broadband damping characteristic for low frequencies. In this way, changes in operating conditions and engine-to-engine variations involving shifts in the combustion pulsation frequency can conveniently be addressed. Genetic algorithms and optimization strategies are used to derive these parameters in a multidimensional parameter space. The novel damper concept is described in more detail and compared with cold-flow experiments. In order to validate the performance under realistic conditions, the new broadband dampers were implemented in a full-scale test engine. Pulsation amplitudes could be reduced by more than 80%. In addition, it is shown that, due to sophisticated damper placement in the engine, two unstable modes can be addressed simultaneously. Application of the damper concept allowed a considerable increase of the engine operating range, thereby reducing NOx emissions by 55%. Predictions obtained with the physics-based model excellently agree with experimental results for all tested damper geometries, bias flows, excitation amplitudes, and most importantly with the measurements in the engine.

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Fig. 1

Schematic block diagram for the damper model

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Fig. 2

Schematic of measurement setup in cold-flow acoustic test facility

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Fig. 3

Comparison of measured reflection coefficients (dashed with ×) and model (solid) for different geometries (a)–(c); top: magnitude, bottom: phase. From red to blue, the ratio between most downstream to most upstream volume decreases.

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Fig. 4

Comparison of measured reflection coefficients (dashed with symbols) and model (solid) for different excitation amplitudes; top: magnitude, bottom: phase

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Fig. 5

Comparison of measured reflection coefficients (dashed with symbols) and model (solid) for different bias flow velocities; top: magnitude, bottom: phase

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Fig. 6

Isosurfaces of damping performance versus combination of geometry parameters P1, P2, and P3. Red (gray) isosurface: 95% (90%) of maximum growth rate reduction.

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Fig. 7

Comparison of ratios of acoustic pressures inside HHD and combustion chamber for measurement (solid) and sixth order model (dashed). Top: magnitude, bottom: phase.

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Fig. 8

Normalized growth rate versus normalized frequency for system without dampers (), equipped with single-volume (°), and multivolume (□) dampers

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Fig. 12

C2n values for distributions 1–6 of damper configuration B. Solid with ×: third azimuthal mode, dashed with °: first azimuthal mode. The values are scaled with the maximum C2n for the respective mode.

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Fig. 11

Averaged spectra of scaled acoustic pressure p∧/p∧max for two different operating conditions

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Fig. 10

Normalized frequency of critical mode without dampers and NOx versus operating parameters A and B. Top: operating window of engine without dampers. Middle: enlarged operating window by using dampers. Bottom: normalized NOx improvement at B3 using dampers.

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Fig. 9

Comparison of reflection coefficients for single-volume (dashed) and multivolume (solid) damper; top: magnitude, bottom: phase




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