Measurement of leakage and estimation of force coefficients ... TRC San Andres...Measurement of...
Transcript of Measurement of leakage and estimation of force coefficients ... TRC San Andres...Measurement of...
Measurement of leakage and estimation of force coefficients in a short length seal supplied with a
liquid/gas mixture (stationary journal)Luis San Andrés,
Xueliang Lu and Qing Liu
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TRC project 32513/1519NS
Year IIIMay 2015
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Bubbly & Foamy Annular Pressure Seals
• As oil fields deplete compressors work off-design with liquid in gas mixtures, mostly inhomogeneous.
• Similarly, oil compression station pumps operate with gas in liquid mixtures.
• The flow condition affects compressor or pump overall efficiency and reliability.
• Little is known about seals operating under 2-phase conditions, except that the mixture affects seal leakage, power loss and rotordynamic force coefficients; perhaps inducing random vibrations that are transmitted to the whole rotor-bearing system.
JustificationSeals operate with either liquids or gases, but not both……
Subsea Application: Wet Gas Compression
Brenne, L., Bjørge, T., Bakken, L., and Hundseid, Ø., 2008, "Prospects for Sub Sea Wet Gas Compression," ASME Paper GT2008-51158.
Brenne, L., Bjørge, T., Gilarranz, J., Koch, J., and Miller, H., 2005, “Performance Evaluation of A Centrifugal Compressor Operating under Wet Gas Conditions,” Proc. 34th
Turbomachinery Symposium, Houston, TX, pp. 111~120
Bertoneri, M., Duni, S., Ransom, D., Podesta, L., Camatti, M., Bigi, M., and Wilcox, M., 2012, “Measured Performance of Two-stage Centrifugal Compressor under Wet Gas Conditions,” ASME Paper GT2012-69819.
Vannini, G., Bertoneri, M., Del Vescovo, G., and Wilcox, M., 2014, “Centrifugal Compressor Rotordynamics in Wet Gas Conditions,” Proc. 43rd Turbomachinery & 30th
Pump Users Symposia, Houston, TX.
Bertoneri, M., Wilcox, M., Toni, L., and Beck, G., 2014, “Development of Test Stand for Measuring Aerodynamic, Erosion, and Rotordynamic Performance of a Centrifugal Compressor under Wet Gas Conditions,” ASME Paper GT2014-25349.
Announce / assess the effects of two phase flow in wet gas compression: drop in efficiency and reliability, excessive vibration and rotordynamic stability.
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Experimental – Seals (two phase)
Iwatsubo, T., and Nishino, T., 1993, “An Experimental Study on the Static and Dynamic Characteristics of Pump Annular Seals,“ 7th Workshop on Rotordynamic Instability Problems in High Performance Turbomachinery, Texas A&M University.
Computational – Seals (two phase)
Literature: Two Phase Flow in Annular Seals
Hendricks, R.C., 1987, "Straight Cylindrical Seals for High Performance Turbomachinery," NASA TP-1850.
Arauz, G., and San Andrés, L., 1998, “Analysis of Two Phase Flow in Cryogenic Damper Seals, I: Theoretical Model, II: Model Validation and Predictions,” ASME J. Tribol., 120.
Beatty, P.A., and Hughes, W.F., 1987, "Turbulent Two-Phase Flow in Annular Seals," ASLE Trans., 30.
Arghir, M., Zerarka, M., Pineau, G., 2009"Rotordynamic Analysis of Textured Annular Seals with Mutiphase (Bubbly) Flow, “Workshop : “Dynamic Sealing Under Severe Working Conditions” EDF – LMS Futuroscope, Universite de Poitiers.San Andrés, L., 2012, “Rotordynamic Force Coefficients of Bubbly Mixture Annular Pressure Seals,” ASME J. Eng. Gas Turbines Power, 134.
NEED EXPERIMENTAL
validation.
.
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Experimental – Seals (two phase)Iwatsubo, T., and Nishino, T., 1993, “An Experimental
Study on the Static and Dynamic Characteristics of Pump Annular Seals,“ 7th Workshop on Rotordynamic Instability
Problems in High Performance Turbomachinery.
NO description of water lubricated seal (L, D, c) or gas type…..
Tests at shaft speed 1.5-3.5 krpm and supply pressure=1.2 - 4.7 bar.
Gas/liquid volume fraction =0, 0.25, 0.45, 0.70
Mxx
Cxx
Kxx
, gas volume fraction increases
Stiffness, damping and mass coefficients decrease
as gas volume fraction increases.
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• Application–Subsea compression and pumping(wet compressors must operate with LVF up to 5%)
• Effect of two component flow on seal– Leakage rate– Dynamic force coefficients– Stability
• Status of current research: – Bulk-flow model available– Lack of experimental validation
Motivation
• Prepared test rig for measurements and improved lubrication system.
• Made oil-gas mixtures with liquid volume fraction (LVF) at inlet plane: gas to liquid.
• Measured flow rates with pure oil, all air, and with mixture.
• Designed (manufactured) new test rig.
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Past TRC work 2013-14 – Year I
• Revamped vertical test rig and troubleshoot.• Generated mixtures and measured flow rate.
– LVF = 0.02% (air) to 100% (liquid)– Supply pressure to 50 psig (3.5 bar (abs))– Rotor speed: stationary
• Conducted dynamic load tests and estimated force coefficients.– Installed shakers and performed single frequency
dynamic load measurements– Identification of force coefficients
• Visual inspection of bubbly mixture. 8
Completed work 2014-15 Year II
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Wet Seal Test Rig
Main frame
Support pipe
Thermocouple
Accelerometer
Load cell
Eddy current sensor
Dynamic pressure sensor
Bearing cartridge
Rotor
Top plate
Air-Oil Mixture
Shaker Rotor
Bearing
Shaker
Shaker
Stinger
seal clearance
X
ZY
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Seal cartridge & fluid propertiesSeal
Diameter 127 mm (5 in)Length 46 mm (1.8 in)clearance 0.127 mm(5 mil)
Supply pressure 1.0~3.5 bar (abs)
Oil ISO VG 10density 830 kg/m3
viscosity 18 cP at 20ºCAir density 1.2 kg/m3 at
1barviscosity 0.02 cP at 20ºC,
1bar
Flow
Seal
rotor & journal
L/D=0.36
X
ZY
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Flows & Mixture
Air InletOil Inlet
(ISO VG 10)
Test seal section
Valve
ValveSparger(mixing) element
Oil
Air
Ps/Pa=1.5, inlet LVF= 2%
Flow rate of mixture vs. supply pressure 1.1 ~3.5bar (abs)Room temperature ~20ºC– Seal inlet liquid volume fraction (at supply pressure):
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Flow Rate Measurement
Mass fraction is constant as there is no mass transfer between oil and air.
LVF= Qlβ =inlet PaQ +Ql g Ps
Oil
Air
0.012
0.014
0.016
0.018
0.02
0.022
0 0.2 0.4 0.6 0.8 1
Mass flow ra
te of m
ixture [k
g/s]
seal inlet LVF
Mass Flow Rate in Wet SealSupply PS=3.5 bar(abs), ambient Pa=1 bar(abs), (non-rotating) journal.
Liquid mass fraction () is large even for small LVF because of large density ratio (oil/air)~200 .Flow ranges from turbulent (pure air) to laminar (pure oil).Predicted flow ~ 6% lower than measured.
L/D=0.36 c=0.127 mm
Ps/Pa=3.5
Reynolds #
Liquid mass fraction
mass flow rate
Measurement
Prediction
Rapid drop
Flow visualization in Wet Sealambient pressure Pa=1 bar(abs),
inlet temperature 20oC. Non-rotating journal.
Inlet LVF=0.45
Smaller bubbles with
higher supply
pressureL/D=0.36 c=0.127 mm
Seal land length: 46mm
Top
Bottom
Top
Bottom
PS=1.5bar(abs)
=0.45
PS=3.0 bar(abs)
=0.45
Bubble expansion
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Force Coefficients in Annular SealsSeal reaction force is a function of the fluid
properties, flow regime, operating
conditions and geometry.
For small amplitudes of rotor motion: the force
is linearized with stiffness, damping
and inertia force coefficients:
c
rotor
L
D
Axial pressure field (liquid)
stator
Pa
PS Pe
W Axial velocity
X
Y
Film thickness H=c+eX coseY sin
rotor
X XX XY XX XY XX XY
Y YX YY YX YY YX YY
F K K C C M Mx x xF K K C C M My y y
Model system (2-DOF): structure + sealEOM: Time Domain
EOM: Frequency Domain
Measure:
Estimate Parameters:
K C M
K, C, M
Parameter Identification
S SEAL S SEAL BC SEAL(K +K )z+(C +C )z+(M +M )z=F
2iω ω K + C - M z = F
Load F=Fo sin(t) Displacement z,
R e ( )
Im ( ) (?)
2ω
ω
H K - M
H C
1( ) ( ) 2iω ω H z = F H K + C - M = F z
0
5
10
15
0 20 40 60 80 100
Dis
plac
emen
t (um
)
Frequency (Hz)
inlet LVF 0%inlet LVF 4%
-90
-60
-30
0
30
60
90
0 0.02 0.04 0.06 0.08 0.1
Load
(N)
Time (s)
inlet LVF 0% inlet LVF 4%0
10
20
30
40
50
0 0.02 0.04 0.06 0.08 0.1
Dis
plac
emen
t (um
)
Time (s)
inlet LVF 0% inlet LVF 4%
Applied load and seal motion X direction, frequency = 30 Hz
FFT
0
5
10
15
20
0 20 40 60 80 100
Load
(N)
Frequency (Hz)
inlet LVF 0%
inlet LVF 4%
FFT
Load Displacement
Unexplained drift at low frequencies
Seal coefficients
(K, C, M)SEAL = (K, C,M) – (K, C, MBC)S
SEALTest system(lubricated) Structure (dry)
•Instrumental Variable Filter Method (IVFM) (Fritzen, 1985, J.Vib, 108)
2Re H K M Im (?) H CComplex stiffness
Parameter Identification
-16
-12
-8
-4
0
4
0 50 100 150 200
Rea
l(H) (
MN
/m)
Frequency (Hz)
Hᵪᵪ
Hᵧᵧ
K - Mω²0
4
8
12
16
20
0 50 100 150 200
Imag
inar
y(H
) (M
N/m
)
Frequency (Hz)
Hᵪᵪ
Hᵧᵧ
Ps/Pa=2, inlet LVF 2%
Test System Dynamic Stiffness2.0 = Supply pressure PS/ambient pressure Pa
inlet temperature 20oC. (non-rotating) journal.
βinlet=0.02
βinlet=0.04
liquid volume fraction at inlet:
K-2M fits well the test
data. -16
-12
-8
-4
0
4
0 50 100 150 200
Rea
l(H) (
MN
/m)
Frequency (Hz)
Hᵪᵪ
Hᵧᵧ
K - Mω²
Uncertainty ±0.7MN/m
-16
-12
-8
-4
0
4
0 50 100 150 200
Rea
l(H) (
MN
/m)
Frequency (Hz)
Hᵪᵪ
Hᵧᵧ
K - Mω²
Test System Quadrature Stiffness2.0 = Supply pressure PS/ambient pressure Pa
inlet temperature 20oC. (non-rotating) journal.
liquid volume fraction at inlet (
Ima(H) shows damping is frequency
dependent. Very little
damping with all gas.
βinlet=0.0
0.0 (all air)
(mm=6±0.8 g/s)
βinlet=0.02
0.90
(mm=6±0.2 g/s)
βinlet=0.04
0.95
(mm=7±0.2 g/s)
20
0
0.2
0.4
0.6
0.8
1
0 50 100 150 200
Imag
inar
y(H
) (M
N/m
)
Frequency (Hz)
Hᵪᵪ
Hᵧᵧ
0
4
8
12
16
20
0 50 100 150 200
Imag
inar
y(H
) (M
N/m
)
Frequency (Hz)
Hᵪᵪ
Hᵧᵧ
Hᵪᵪ Hᵧᵧ
Inlet LVF 0%
Inlet LVF 4%
Uncertainty: ±0.7MN/m
10
100
1,000
10,000
100,000
0 50 100 150 200
Cyy
_sea
l (N
-s/m
)
Frequency (Hz)
10
100
1,000
10,000
100,000
0 50 100 150 200
Cxx
_sea
l (N
-s/m
)
Frequency (Hz)
Wet Seal Direct Damping C=Im(H)/2.0 = Supply pressure PS/ambient pressure Pa
Stationary (non-rotating) journal. βinlet=0.04
βinlet=0.02
βinlet=0 (all air)
βinlet=0.02
βinlet=0
βinlet=0.04
liquid volume fraction at inlet 0% (all gas), 2%, 4%.
Damping is frequency dependent.
Small LVF causes damping to increase 20
fold w/r to air only.
CXX & CYY differ due to uneven clearance and
inhomogeneous flow
seal stiffness & mass coefficients
LVF at seal Liquid mass KXXseal KYYseal MXXseal MYYseal
inlet fraction MN/m MN/m kg kg
0 0 1.4 1.0 0 0
2% 90% 1.5 1.3 0.3 0.7
4% 95% 2.5 1.3 1.2 0.9
With mixture, CXX>CYY due to uneven clearance and mixture not homogenous (one phase travels faster than the other).
Wet
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Conclusion• Mass flow rate through seal dictated by liquid
content (density ratio ~ = 830/1.2=692).
• System dynamic stiffness K-2M fits well with testdata.
• Damping coefficients are frequency dependent.For operating conditions with a small volume fractionof oil in gas (4%), damping can be up to twenty timeslarger than that obtained for the pure gas condition.
• The effect of a few droplets of liquid on affectingthe test system forced response is overwhelming.
LEAKAGE AND FORCE COEFFICIENTS IN AWET ANNULAR SEAL: INFLUENCE OFJOURNAL ROTATIONAL SPEED.
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2015-2016 Continuation Proposal Year III
• Redesign stingers and mount shakers rigidly to avoid low frequency drift.
• Install a bubble eliminator to remove air content from the bubbly lubricant.
• Improve DAQ process to perform ensemble averages over multiple periods of external excitation.
• Benchmark bulk-flow model predictions1.• Extend existing predictive model1 to consider
inhomogeneous flow.
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Proposed Work
San Andrés, L., 2012, “Rotordynamic Force Coefficients of Bubbly Mixture Annular Pressure Seals,” ASME J. Eng. Gas Turbines Power, 134 (2), p. 022503.
2015-2016 Year III
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TRC Budget 2015-2016 Year III
Support for GS(20 h/week) x $ 2,200 x 12 months $ 26,400
Support for UGS (10 h/week) x 9 months $ 2,250
Fringe benefits (2.7%) & medical insurance ($150/month) $ 2,574
Registration and travel to technical Conference $ 1,250
Tuition three semesters ($ 363 credit hour x 24 ch/year) $ 8,712
Supplies: Bubble eliminator $1,050+flow meter repair $790+ oil $600
$ 2,440
PC upgrade ($1,150) + Mathcad ($105) + storage ($100) $ 1,355
Total BUDGET $ 44,981
Acknowledgments
Questions (?)
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Thanks to the Turbomachinery Research Consortium for its support