Research Papers: Gas Turbines: Structures and Dynamics

Brush Seal Frictional Heat Generation—Test Rig Design and Validation Under Steam Environment

[+] Author and Article Information
Markus Raben

TU Braunschweig,
Institute of Jet Propulsion and Turbomachinery,
Hermann-Blenk-Str., 37,
Braunschweig 38108, Germany
e-mail: m.raben@ifas.tu-bs.de

Jens Friedrichs

TU Braunschweig,
Institute of Jet Propulsion and Turbomachinery,
Hermann-Blenk-Str., 37,
Braunschweig 38108, Germany
e-mail: j.friedrichs@ifas.tu-bs.de

Johan Flegler

Siemens AG,
Power and Gas Division,
Rheinstr. 100,
Mülheim a. d. Ruhr 45478, Germany
e-mail: johan.flegler@siemens.com

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 10, 2016; final manuscript received July 18, 2016; published online October 4, 2016. Editor: David Wisler.The content of this paper is copyrighted by Siemens Energy, Inc. and is licensed to ASME for publication and distribution only. Any inquiries regarding permission to use the content of this paper, in whole or in part, for any purpose must be addressed to Siemens Energy, Inc. directly.

J. Eng. Gas Turbines Power 139(3), 032502 (Oct 04, 2016) (9 pages) Paper No: GTP-16-1320; doi: 10.1115/1.4034500 History: Received July 10, 2016; Revised July 18, 2016

Sealing technology is a key feature to improve efficiency of steam turbines for both new power stations and modernization projects. One of the most powerful sealing alternatives for reducing parasitic leakages in the blade path of a turbine as well as in shaft sealing areas is the use of brush seals, which are also widely used in gas turbines and turbo compressors. The advantage of brush seals over other sealing concepts is based on the narrow gap that is formed between the brush seal bristle tips and the mating rotor surface together with its radial adaptivity. While the narrow gap between the bristle tips and the rotor leads to a strongly decreased flow through the seal compared with conventional turbomachinery seals, it is important to be aware of the tight gap that can be bridged by relative motion between the rotor and the brush seal, leading to a contact of the bristles and the rotor surface. Besides abrasive wear occurrence, the friction between the bristles and the rotor leads to heat generation which can be detrimental to turbine operation due to thermal effects, leading to rotor bending connected to increasing shaft vibrations. In order to investigate the frictional heat generation of brush seals, different investigation concepts have been introduced through the past years. To broaden the knowledge about frictional heat generation and to make it applicable for steam turbine applications, a new testing setup was designed for the steam test rig of the Institute of Jet Propulsion and Turbomachinery—TU Braunschweig, Germany, enabling temperature measurements in the rotor body under stationary and transient operation in steam by using rotor-integrated thermocouples. Within this paper, the development of the instrumented new rotor design and all relevant parts of the new testing setup is shown along with the testing ability by means of the validation of the test rig concept and the achieved measurement accuracy. First results prove that the new system can be used to investigate frictional heat generation of brush seals under conditions relevant for steam turbine shaft seals.

Copyright © 2016 by Siemens AG
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Fig. 1

Clamped seal design (left) and welded seal design (right)

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Fig. 2

Schematic heat transfer mechanisms

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Fig. 4

New rotor design: (a) 3D CAD model and (b) section view center body (A-A)

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Fig. 5

Superficial TC wiring

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Fig. 6

Rotor instrumentation—measuring positions: (a) instrumentation displacement end and (b) instrumentation drive end

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Fig. 7

Radial rotor growth

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Fig. 8

Open rotor end—redesigned thrust bearing

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Fig. 9

Reference temperature (RTD)

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Fig. 10

Temperature measurements displacement end (a) TC nomenclature and (b) measurements 24 h (stationary operation 30 bar)

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Fig. 11

Calculated rotor temperature profile, displacement end

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Fig. 12

Difference between calculations (FEA) and measurements, displacement end

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Fig. 13

Compensation area—additional measuring positions




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