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Research Papers: Gas Turbines: Structures and Dynamics

Leakage, Drag Power, and Rotordynamic Force Coefficients of an Air in Oil (Wet) Annular Seal

[+] Author and Article Information
Luis San Andrés

Mast-Childs Chair Professor
Fellow ASME
Mechanical Engineering Department,
Texas A&M University,
College Station, TX 77843
e-mail: Lsanandres@tamu.edu

Xueliang Lu

Mechanical Engineering Department,
Texas A&M University,
College Station, TX 77843
e-mail: luliang413@gmail.com

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received July 3, 2017; final manuscript received July 5, 2017; published online September 19, 2017. Editor: David Wisler.

J. Eng. Gas Turbines Power 140(1), 012505 (Sep 19, 2017) (11 pages) Paper No: GTP-17-1268; doi: 10.1115/1.4037622 History: Received July 03, 2017; Revised July 05, 2017

Wet gas compression systems and multiphase pumps are enabling technologies for the deep sea oil and gas industry. This extreme environment determines both machine types have to handle mixtures with a gas in liquid volume fraction (GVF) varying over a wide range (0–1). The gas (or liquid) content affects the system pumping (or compression) efficiency and reliability, and places a penalty in leakage and rotordynamic performance in secondary flow components, namely seals. In 2015, tests were conducted with a short length smooth surface annular seal (L/D = 0.36, radial clearance = 0.127 mm) operating with an oil in air mixture whose liquid volume fraction (LVF) varied to 4%. The test results with a stationary journal show the dramatic effect of a few droplets of liquid on the production of large damping coefficients. This paper presents further measurements and predictions of leakage, drag power, and rotordynamic force coefficients conducted with the same test seal and a rotating journal. The seal is supplied with a mixture (air in ISO VG 10 oil), varying from a pure liquid to an inlet GVF = 0.9 (mostly gas), a typical range in multiphase pumps. For operation with a supply pressure (Ps) up to 3.5 bar(a), discharge pressure (Pa) = 1 bar(a), and various shaft speed (Ω) to 3.5 krpm (ΩR = 23.3 m/s), the flow is laminar with either a pure oil or a mixture. As the inlet GVF increases to 0.9 the mass flow rate and drag power decrease monotonically by 25% and 85% when compared to the pure liquid case, respectively. For operation with Ps = 2.5 bar(a) and Ω to 3.5 krpm, dynamic load tests with frequency 0 < ω < 110 Hz are conducted to procure rotordynamic force coefficients. A direct stiffness (K), an added mass (M), and a viscous damping coefficient (C) represent well the seal lubricated with a pure oil. For tests with a mixture (GVFmax = 0.9), the seal dynamic complex stiffness Re(H) increases with whirl frequency (ω); that is, Re(H) differs from (K−ω2M). Both the seal cross coupled stiffnesses (KXY and −KYX) and direct damping coefficients (CXX and CYY) decrease by approximately 75% as the inlet GVF increases to 0.9. The finding reveals that the frequency at which the effective damping coefficient (CXXeff = CXX − KXY/ω) changes from negative to positive (i.e., a crossover frequency) drops from 50% of the rotor speed (ω = 1/2 Ω) for a seal with pure oil to a lesser magnitude for operation with a mixture. Predictions for leakage and drag power based on a homogeneous bulk flow model match well the test data for operation with inlet GVF up to 0.9. Predicted force coefficients correlate well with the test data for mixtures with GVF up to 0.6. For a mixture with a larger GVF, the model under predicts the direct damping coefficients by as much as 40%. The tests also reveal the appearance of a self-excited seal motion with a low frequency; its amplitude and broad band frequency (centered at around ∼12 Hz) persist and increase as the gas content in the mixture increase. The test results show that an accurate quantification of wet seals dynamic force response is necessary for the design of robust subsea flow assurance systems.

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References

Figures

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Fig. 1

(a) Cross section of seal test with direction of flow noted and (b) schematic top view of test seal with coordinate system [10]

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Fig. 2

Seal diametrical clearance (2 c) versus seal cartridge temperature. c = 0.203 mm at 34 °C. Measurements without shaft rotation and predictions based on model in Ref. [16].

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Fig. 3

Air and lubricant circulation flow system

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Fig. 4

Flow visualization of wet seal operating with a gas and oil mixture. Inlet GVF = 0.9 and journal speed = 1.8 krpm (30 Hz). Pictures taken with a stroboscope light at 30 Hz. Seal inlet pressure (Ps) = 2.0 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 5

Normalized wet seal leakage (m¯˙m) versus mixture inlet GVF. Supply pressure (Ps) to 3.5 bar(a) and discharge pressure (Pa) = 1 bar(a). Shaft speed N = 0, 1.5 krpm, and 3.5 krpm (ΩR = 23.3 m/s).

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Fig. 6

Seal drag power loss (Pow¯) (normalized) versus mixture inlet GVF. Journal speed (N) = 1.5–3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 7

Example of waterfall plot of SC response (Y). Test with oil (GVF = 0), Ps = 2.5 bar(a), shaft speed N = 3.5 krpm (58.3 Hz), and external load excitation frequency at 80 Hz.

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Fig. 8

Real part of seal complex stiffnesses Re(HXX)seal and Re(HYY)seal versus whirl frequency (ω). Inlet GVF = 0–0.9. Shaft speed = 3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 9

Imaginary part of seal complex stiffnesses Ima(HXX)seal and Ima(HYY)seal versus whirl frequency (ω). Inlet GVF = 0–0.9. Shaft speed = 3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 10

Real part of seal complex stiffnesses Re(HXY)seal and−Re(HYX)seal versus whirl frequency (ω). Inlet GVF = 0–0.9. Shaft speed (N) = 3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 11

Seal damping coefficient: predicted (CXX = CYY)seal and experimental (1/2) (CXX + CYY)seal versus frequency (ω) and inlet GVF = 0–0.9. Shaft speed = 3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 12

Seal cross-coupled dynamic stiffness: predicted (KXX = −KXY)seal and experimental (1/2) (KXX − KYX)seal versus frequency (ω) and inlet GVF = 0–0.9. Shaft speed = 3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 13

Seal direct dynamic stiffness: predicted (KXX − ω2MXX)seal and experimental (1/2) (KXX + KYY)seal versus frequency (ω) and inlet GVF = 0–0.9. Shaft speed = 3.5 krpm. Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 14

Seal effective damping coefficient: (a) predicted (CXX = CYY)eff and experimental (1/2) (CXX+CYY)eff versus frequency (ω) and inlet GVF = 0–0.9. (b) Expanded graph showing region of cross frequency for Ceff ∼ 0. Shaft speed = 3.5 krpm (58.3 Hz). Supply pressure (Ps) = 2.5 bar(a) and discharge pressure (Pa) = 1 bar(a).

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Fig. 15

Frequency spectra of SC response (Y). Air in oil mixture with inlet GVF = 0.0–0.9. Tests with a supply pressure (Ps) = 2.5 bar(a) and ambient pressure (Pa) = 1 bar(a). Journal speed N = 3.5 krpm (58.3 Hz) and excitation frequency = 80 Hz.

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Fig. 16

Waterfall plot of SC response (Y) for air in oil mixture with inlet GVF = 0.9. Supply pressure (Ps) = 2.5 bar(a) and ambient pressure (Pa) = 1 bar(a). Journal speed 3.5 krpm (58.3 Hz) and excitation frequency = 80 Hz.

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Fig. 17

Acoustic resonance frequency (symbols) versus mixture inlet GVF and two pressures, 1 bar(a) and 2.5 bar(a). 12 Hz denotes center of SSV.

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