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Research Papers: Gas Turbines: Structures and Dynamics

An Investigation of Flow, Mechanical, and Thermal Performance of Conventional and Pressure-Balanced Brush Seals

[+] Author and Article Information
Michael J. Pekris

Transmissions, Structures & Drives,
Rolls-Royce plc.,
Derby DE24 8BJ, UK
e-mail: michael.pekris@rolls-royce.com

Gervas Franceschini

Transmissions, Structures & Drives,
Rolls-Royce plc.,
Derby DE24 8BJ, UK
e-mail: gervas.franceschini@rolls-royce.com

David R. H. Gillespie

Department of Engineering Science,
University of Oxford,
Oxford OX1 3PJ, UK
e-mail: david.gillespie@eng.ox.ac.uk

Contributed by the International Gas Turbine Institute (IGTI) of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received September 18, 2012; final manuscript received November 30, 2013; published online January 24, 2014. Editor: David Wisler.

J. Eng. Gas Turbines Power 136(6), 062502 (Jan 24, 2014) (11 pages) Paper No: GTP-12-1364; doi: 10.1115/1.4026243 History: Received September 18, 2012; Revised November 30, 2013

Compliant contacting filament seals such as brush seals are well known to give improved leakage performance and hence specific fuel consumption benefit compared to labyrinth seals. The design of the brush seal must be robust across a range of operating pressures, rotor speeds, and radial build-offset tolerances. Importantly the wear characteristics of the seal must be well understood to allow a secondary air system suitable for operation over the entire engine life to be designed. A test rig at the University of Oxford is described which was developed for the testing of brush seals at engine-representative speeds, pressures, and seal housing eccentricities. The test rig allows the leakage, torque, and temperature rise in the rotor to be characterized as functions of the differential pressure(s) across the seal and the speed of rotation. Tests were run on two different geometries of bristle pack with conventional, passive, and active pressure-balanced backing ring configurations. Comparison of the experimental results indicates that the hysteresis inherent in conventional brush seal design could compromise performance (due to increased leakage) or life (due to exacerbated wear) as a result of reduced compliance. The inclusion of active pressure-balanced backing rings in the seal designs are shown to alleviate the problem of bristle–backing ring friction, but this is associated with increased blow-down forces which could result in a significant seal-life penalty. The best performing seal was concluded to be the passive pressure-balanced configuration, which achieves the best compromise between leakage and seal torque. Seals incorporating passive pressure-balanced backing rings are also shown to have improved heat transfer performance in comparison to other designs.

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References

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Figures

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Fig. 5

Typical run-down test results

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Fig. 6

Effective clearance versus rotor speed, backing ring type, and centered housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 7

Effective clearance versus rotor speed, backing ring type, and 250 μm offset housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 8

Idealized pressure-balanced brush seal domain

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Fig. 9

Pressure-balanced brush seal domain CFD results on a near radial—axial plane aligned with the bristle lay angle: (a) pressure (Pa) and (b) velocity (ms−1)

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Fig. 10

Variation of normalized seal leakage with backing ring lip thickness (Pcav = Pup)

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Fig. 11

Maximum two-seal torque versus rotor speed, Pup, and centered housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 12

Maximum two-seal torque versus rotor speed, Pup, and 250 μm offset housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 13

Rotor temperature versus rotor speed and centered housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 14

Rotor temperature versus rotor speed, Pup, and 250 μm offset housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 4

Typical rotor temperature history—pulsed pressure run-down tests

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Fig. 3

Engine seal test facility air system schematic [17]

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Fig. 2

Schematic diagram of the Oxford Engine Seal Test Facility

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Fig. 1

Typical brush seal configurations

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Fig. 15

Effective clearance versus rotor speed, balance cavity pressure ratio, and centered housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 16

Seal torque versus rotor speed, balance cavity pressure ratio, and centered housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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Fig. 17

Rotor temperature versus rotor speed, balance cavity pressure ratio, and centered housing: (a) Pup = 2 bar g and (b) Pup = 10 bar g

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