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1
THE INFLUENCE OF LUBRICANT SUPPLY
CONDITIONS AND BEARING CONFIGURATION
ON THE PERFORMANCE OF (SEMI) FLOATING
RING BEARING SYSTEMS FOR TURBOCHARGERS
Funded by Honeywell Transportation
Systems (HTT)
GT2017-64839
Luis San AndrésMast-Childs Chair Professor
ASME Fellow
Texas A&M University
Feng YuProduct Line Engineer
Honhua America LLC
Kostandin GjikaSenior Fellow
Honeywell TT
Proceedings of ASME Turbo Expo 2017: Turbomachinery Technical Conference
and Exposition
Accepted for
journal
publication
Fully Floating Bearing Semi Floating Bearing Ball Bearing
Bearing supports in commercial turbos
Increased IC engine performance & efficiency
demands of robust turbocharging solutions.
The
driver:
Oil lubricated bearings are cheap with longer
life span; but prone to harmful sub
synchronous whirl & depend heavily on
engine oil condition.
Expensive with
limited lifespan
2
Major challenges
extreme operating conditions:
• - Low Oil Viscosity, e.g. 0W30 or 0W20
• - High Oil Temperature (up to 150°C)
• - Low HTHS (2.4); Low Oil Pressure (1 bar),
• - Increased Max. Turbocharger Speed 5 kHZ
• - Variable Geometry Turbo Technology & Assisted e-power start up
• - High Engine Vibration Level
• - More Stringent Noise Requirements
Need predictive tool to reduce costly engine test stand
qualification.
Thermal management & reduced thermal loading
3
• TC linear and nonlinear rotordynamics codes including engine induced excitations
• Realistic bearing models: thermohydrodynamic
• Novel methods to estimate imbalance distribution and shaft temperatures
• NL analysis for frequency jumps (internal & combined resonances) and noise reduction
• Measured ring speeds with fiber optic sensors.
• Rapid design models for (S)FRB systems
HTT funded work at Texas A&M
Predictive tool for shaft motion benchmarked by test data
2004 IMEchE J. Eng. Tribology
2005 ASME J. Vibrations and Acoustics
ASME DETC 2003/VIB-48418
ASME DETC 2003/VIB-48419
2007 ASME J. Eng. Gas Turbines Power
ASME GT 2006-90873
2007 ASME J. Eng. Gas Turbines Power
ASME GT 2005-68177
2007 ASME J. Tribology
IJTC 2006-12001
2007 ASME DETC2007-34136
2010 ASME J. Eng. Gas Turbines Power
ASME GT2009-59108
2010
2012
IFToMM Korea
GT2012-68355 Best Paper Award
4
Thermal energy analysis in TCs complicated b/c of
a) Hot gas -work and heat flow from turbine
b)Cold gas +work and heat flow from the
compressor
c) internal heat flow across shaft from T to C and
radially thru bearings
d)Mechanical drag power in bearings
e) Heat flow to/from casing to ambient (convective and
radiant)
Conjugate heat transfer in TCs
Engine lubricated bearings enable low friction-load
support and effective cooling (oil carries away
heat)
5
turbinecompressor
Heat flows & energy transfer in a TCHot air
(energy)
in
-Hot air
(energy)
out
Heat
conducted
(casing)
Heat
conducted
(shaft)
Cold air
(energy)
in
Compressed
hot air (energy)
out
Heat
conducted
(casing)
Bearing
drag power
generation
Oil in
Oil out
Baines, Wygnant & Dris, 2010, J Eng Gas Turb. Power
Heat
conducted
(casing)
work
6
Past literature on CHT
Literature using 3D fluid/solid complex models and
interaction is not yet practical for routine engineering
Shaaban
(2004)
uses empirical formulas to model thermal energy flows oil
takes up to 49% of the total amount of turbine heat loss at a
fully load operation.
Baines (2010) correlates convective heat transfer coefficients based on test
data internal heat transfer is much higher (> 10) than the
amount convected to ambient.
Bohn (2005) uses CHT method to guarantee TC operation w/o thermal
overloading heat into the compressor heats the intake air.
San Andrés
(2012)
analyzes thermal energy transport in the films predict inner
film carries away most of the total energy (>70%).
7
a) Lumped parameter models with empirical
coefficients for heat transfer coefficients and
simplified formulas for drag power, flow & heat
flow in the oil bearings
b)3D CFD-FE modeling stresses on solids with
over-simplified coupling to the lubricant flows
Thermal energy analyses
Objective is to avoid oil coking, optimize
flow rates, ensure proper clearances,
eliminate seizure.
Engineered thermal management aims
to avoid severe thermal loading with
improved reliability of bearing system:
GOAL
8
9
Justification
Cooling vs. drag power loss
Lubricant
flow rate
Cooling effect
Drag power loss
Oil overheats, flashes &
burns (coke)
Lower TC mechanical
efficiency
Warmer air at
compressor intake,
lower engine efficiencyTrade off
Small
Large
For safe and optimal operating conditions, engineers
must quantify the transport of thermal energy in the
support bearing system.
10
Tasks
• Develop model for prediction of temperature
fields in the oil films and floating ring, as well as
the thermal energy flows in the bearing system.
• Validate thermo-hydrodynamic model of (S)FRB
• Analyze performance characteristics of al
(S)FRB for a typical PV TC.
• Assess impact of geometry and operating
parameters on (S)FRB performance.
Film temperature distribution
Ring temperature distributionThermal energy flows based on temperatures of films and
shaft/ring/casing.
Control Volume Method
Finite Element Method
11
The THD model for a (S)FRB
system
(S)FRB system in engine-oil lubricated TC
turbine
compressor
Semi-floating
ring bearing system
Compressor
side bearingoil supply
holes
to inner and
outer films
Oil supply
holes to
inner film
Turbine
side
bearing Outer film
with ½
moon
groove
Oil supply at
Psup, Tsup
shaft
casing
Hot oil discharges at
ambient pressure Pa
Lubricant flow paths into bearings on turbine and
compressor sides
X
Y
θ
r
½ moon groove
DJ
13
Laminar flow & steady-state;
Reynolds equationwith oil viscosity shear and
temperature dependent.
Inner film
h : fluid film thickness P : hydrodynamic pressure
RJ : journal radius RRo: ring radius (OD)
J: journal speed R: ring speed ( R=0, for semi FRB)
Equations for pressure (P) in thin filmsOil Supply hole, TSUP, PSUP
Casing
Outer film
Inner film
Ring
Shaft
DRo
Below i and o denote inner and outer films
Outer film
3 3
2
1
12 12 2
i i i i J R i
J i i i i
h P h P h
R z z
3 3
2
1
12 12 2o
o o o o oR
R o o o o
h P h P h
R z z
0
shaft
Ring
TS
casing
Outer film
Inner film
TRi
TRo
TC
Ti
To
Heat flow
from shaft Mechanical
drag power
Heat flow
into ring
Mechanical
drag power
Heat flow
into casing
Flow outer
Flow inner
Heat flow
carried
by oil
Heat flows & drag power in a FRB
14
15
Thermal energy transport in thin film flows
T: temperature h : film thickness U,W: circ. & axial flow velocities
, r, Cv : oil viscosity & density, specific heat HJ, HR, HC : heat convection coefficients
TS, TR,TC : Journal, ring and casing temperatures : journal speed R: ring speed
Shear drag energy
dissipation + heat
convection = thermal
energy advected by
films
1
( ) ( )o o
o
o R R o C o C v zo oR o
H T T H T T C m T m TR z
1
( ) ( )i ii J S i R i R v zi i
J i
H T T H T T C m T m TR z
Inner film
Outer film
Floatingring
Casing
Dragpower
Heat flow from shaft
Heat flow into casing
Heat flow
through ring
Energy
carried
by oil
Outer film
Shaft
Inner film
ΩR
Ω
In a (S)FRB, Ф o~0
2 22 1
1212 i
i i S R i m
i
W U U U Uh
22 21 1
12 212 o
o o Ro o Ro
o
W U U Uh
12
, , iS J R R R m S RU R U R U U U
R
R R RR R R P R
T T Tr C r
r r r r r r r
Ring
Ring temperature varies along
circ. & radial directions0
16
Validate thermo-hydrodynamic
model for a (S)FRB
There is no (S)FRB test data available.
Compare predictions of pressure and
temperature to (published) test data for a
journal bearing.
17
View of an (old) journal bearing test rig
Pressure sensors
(×16, equally)
X
Y
W
Driving
motor
Support
bearing
Support
bearing
Shaft
(hollow)
Seals
Bearing housing
(steel)
Test bearing bushing
(bronze)
Steel disks
(adding load
to the journal)
Bearing midplane
Loading
system
Oil supply
Lubricant
Thermocouples
(×16, equally)
Oil inlets
Bearing Diameter D=100 mm
Length L =55 mm
Cold clearance at 20ºC, C0*=75 µm
Tonnesen & Hansen,
1981, “Some Experiments
on the Steady State
Characteristics of a
Cylindrical Fluid-Film
Bearing Considering
Thermal Effects,” J. Lub.
Tech., 103.
Measurement vs. Prediction: Pressure field
0
5
10
15
20
25
30
0 40 80 120 160 200 240 280 320 360
Circumferential angle (θ )
Pre
ssu
re (
ba
r)
Test data
Prediction
Shaft speed = 3,200 rpm, W = 5,000 N, (e,Ф) ≈ (45 μm,
50°). TSUP = 40°C, PSUP = 2 bar. Tshaft = 62 °C
(measured).
Oil cavitation zone16 pressure sensors
equally spaced at
bearing mid plane
e
ē = 0.6
Load
Xθ
Y
18
Tonnesen & Hansen, 1981
16 thermocouples
equally spaced at the
bearing mid plane
30
40
50
60
70
80
90
0 30 60 90 120 150 180 210 240 270 300 330 360
Circumferential angle (θ )
Te
mp
era
ture
(°C
)
Test data
Prediction
TSUP = 40°C
10° Axial supply groove x 2
(180° apart)
Upper pad
Y
X
θ
ē ≈ 0
No load
N = 4,800 rpm, no load, (e,Ф) ≈ (0, 0)
30
40
50
60
70
80
90
0 30 60 90 120 150 180 210 240 270 300 330 360
Circumferential angle (θ )
Te
mp
era
ture
(°C
)
Test data
Prediction
Bottom pade
Load
X
Y
θ
ē= 0.6
Upper pad
Bottom pad
N = 6,400 rpm, W = 5 kN, (e,Ф) ≈ (45μm, 50°)
Bottom pad
Tavg = 77 °C
Due to assuming bearing
temperature does not vary axially
19
Measurement vs. Prediction: Temperature field
PSUP = 2 bar. TSUP = 40°C , Tshaft = 82 °C (test)
-1.0
-0.8
-0.6
-0.4
-0.2
0.0
0.0 0.2 0.4 0.6 0.8 1.0
Eccentricty (X) relative to the nominal (20°C) film clearance
Ecce
ntr
icty
(Y
) re
lative
to
th
e n
om
ina
l (2
0°C
) film
cle
ara
nce
Test data
Prediction
e
Load
X
Y
θ
Nominal film clearance at
20°C = 7.5 µm
9,000 N
5,000 N
2,200 N
N = 3.2 krpm, W = 100 ~ 9,000 N,
PSUP = 2 bar, TSUP = 40°C, Tshaft = 62°C
Caused by changes in clearance due to
mechanical over-stress as load is high
W = 9 kN (W/LD) = 16 bar.
Note:
Typical load in a (S)FRB
~ 0.25 bar
20
Measurement vs. Prediction: Journal locus
Tonnesen & Hansen, 1981
21
Predict performance
characteristics of a (S)FRB for a
typical PV TC.
22
Example (S)FRB turbine side bearing
SAE 5W-30100C-150C
200C
Shaft (journal)
RING
CASING
Oil inlet
Oil Supply temperature, TSUP* 100°C ~150°C
supply pressure, PSUP* 4 bar(gauge)
Shaft temperature, Tshaft ~200°C
Bearing
Length/diameter, Li/Ds, Lo/DC 0.6, 0.44
Nominal clearances ci/L , co/L 1.6×10-3 , 7.6×10-3
# holes and axial grooves 2~7
Brass ring thickness t/Ds=0.40
Static load, W/(LD) 0.25 bar
Maximum shaft speed 240 krpm
Z
Oil supply hole (×4)Inner film clearance (μm)
1 3 5 7 9 11 13 15 17 19 21 23 25 27 29 31 33 35 37 39 41 43 45 47 49 51S1
S3
S5
S7
S9
S11
circ coordinate (node #)
Axial
coordinate
1 3 5 7 9 11 13 15 17 19 21 23 25 27 29 31 33 35 37 39 41 43 45S1
S3
S5
S7
S9
S11
Circ. coordinate (node #)
Axial
coordinate
X
X
Mesh: outer film, NEX=45, NEZ =12
Mesh: inner film, NEX=52, NEZ =12
Mesh: ring
Nθ=40, NR =8 XR
0
10
20
30
40
Axial groove
(×4)
0
200
400
600
Circ. groove½ moon groove on the
casing
Moon groove
R
θ
YR
Outer film clearance (μm)
Meshes for analysis of flow in (S)FRB
23
W
θ
eJ = 0.02
eR = 0.38
X
YOil cavitation
Axial groove×4
Pressure & temperature fields in inner film
Pressure
Temperature
Film
thickness Hot spot
Shaft speed
= 105 krpm
24
Oil heats
quickly
hot spot at
exit plane.
Temperature field in floating ring
135
145
155
165
175
0 60 120 180 240 300 360
Circumferential coordinate(°)
Rin
g s
urf
ace
te
mp
era
ture
(°C
)
Ring inner surface Ring outer surface
Ring ID
145
150
155
160
165
170
Ro
Ri
θ
Dash line: Tavg
20°C
Axial groove×4
(on ring ID)
W
X
θ
Y
Je
Re = 0.19
= 0.01
Shaft speed =
240 krpm
15°C
Ring OD
25
Ring temperature
field depends on
material conductivity
and operating
conditions.
Large
temperature
gradient across
ring.
100
120
140
160
180
200
220
0 50 100 150 200 250
Journal speed (krpm)
Te
mp
era
ture
(°C
)
Mixing Temperature Inner film exit temperature Outer film exit temperature
Exit (oil) mixing temperature
Inner film exit temperature
Outer film exit temperature
Tshaft
Tsup
Tmax = 0.95 Tshaft
Film exit temperature vs. shaft speed
* Film exit temperature is average around circumference at exit plane (z= ½ L)
shaft speed increases
Inner film carries thermal energy generated in the film and draws heat from hot
shaft
Outer film carries away a small amount of heat with some heat conduction to
casing
flowinner > flowouterTmixing ~Touter and << Tinner
26
0%
20%
40%
60%
80%
100%
0 50 100 150 200 250
Journal speed (krpm)
En
erg
y p
ort
ion
s fo
r H
ea
t flo
ws
Heat carried by inner film Heat into ring
Heat carried away by inner film
Heat into ring
Heat flow & drag power loss vs. shaft speed
As shaft speed increases,
drag power loss steadily
raises.
Heat drawn from shaft
decreases because inner film temperature increases.
Heat carried
by inner film
Drag power
loss
Heat from
shaft
Heat into
ring= + -
Inner film:
Heat carried
by outer film
Drag power loss
Heat into ring
Heat into
casing= + -
Outer film: 0
27
0
50
100
150
200
250
300
350
0 50 100 150 200 250
Journal speed (krpm)
He
at flo
w a
nd
dra
g p
ow
er
loss (
W)
Heat from shaft Heat through ring
Heat carried by inner film Drag power loss
Heat to casing Total energy
Heat carried by inner film
Heat from shaft +
drag power loss
Drag power loss
Heat into ring
Heat from shaft
Heat into casing
Thermal energy transport and balance
29
Quantify impact of geometry and
operating parameters on (S)FRB
performance
X
Y
θ
Inner film
Outer film
Shaft
Casing
Depth_o
X
Y Y
Z
Length_o
Width_i
Depth_iGroove
Hole
Ring side viewRing front view
ci : inner film clearance
co: outer film clearance
Ng: number of axial
grooves (supply holes)
Ring ID:Width_i: inner groove
width
Depth_i: inner groove
depth
Ring OD:Length_o: outer groove
length
Depth_o: outer groove
depth
Floating ringRing isometric view
Notation for geometry of oil supply arrangement
30
Parameters to
vary:
Influence of supply temperature on film temperature
Inner film
T
120ºC(nominal)
31
0.2
0.4
0.6
0.8
1.0
0 60 120 180 240 300 360
Tem
pe
ratu
re / T
shaft
Circumferential coordinate(°)
0.75 Tsup* at z=0 Tsup* at z=0 1.2 Tsup* at z=0
0.75 Tsup* at z=L/4 Tsup* at z=L/4 1.2 Tsup* at z=L/4
z= ¼ L
z=0
(midplane)
Inner Film Max. Journal speed
4 x axial groove
0.75TSUP*
1.2TSUP*
1.0TSUP*
Tshaft
temperature vs
at z = 0 and z = L/4
Inner film temperature at θ = 240°
Oil supply temperature varies
1.2 Tsup
1 Tsup
0.75 Tsup
Max. shaft speed
Quick temperature
growth along axial plane
is due to large heat
drawn from hot shaft.
T/Tshaft
32
Exit
temperature
of inner film
~ shaft
temperature
Influence of oil supply temperature on film temperatures
Influence of TSUP on inner film temperature is felt only near the oil
inlet region. Oil average viscosity (inner film) is nearly the same for
all TSUP
0.2
0.4
0.6
0.8
1
0.7 0.8 0.9 1 1.1 1.2
Tem
pe
ratu
re / T
shaft
Oil supply temperature / nominal TSUP
Average temperature (inner f ilm) Peak temperature (inner f ilm)
Average temperature (outer f ilm) Peak temperature (outer f ilm)
Inner film average temperature
Max. shaft speed
Average
temperature
Outer film
Inner film peak temperature (exit plane)
Oil supply temperature increases
T/Tshaft
0.08
0.12
0.16
0.20
0.24
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 1.1
Lu
brica
tio
n flo
w (
-)
Journal speed / Max. journal speed (-)
0.83 Tsup* Tsup* 1.17 Tsup*
33
Larger thermal growth for
higher TSUP, leads to
increase in film flow
Average oil viscosity is
same for various TSUP,, as
shaft temperature
determines film
temperature.
SAE 5W-30, 100
1402
C
C
.
Lower average oil
viscosity for higher TSUP
Influence of oil supply temperature on flow rates
Inner film flow
Outer film flow
1.17 Tsup 1 Tsup
0.83 Tsup
1.17 Tsup
1 Tsup0.83 Tsup
shaft speed increases
0%
20%
40%
60%
80%
0.7 0.8 0.9 1.0 1.1 1.2 1.3
En
erg
y p
ort
ion
of
dra
g p
ow
er
loss (
%)
Oil supply pressure / Nominal PSUP
1/8Ωmax 3/8Ωmax 5/8Ωmax Ωmax
Psup nominal
34
Influence of oil supply pressure on drag power loss
PSUP has little influence on drag power loss since average
inner film temperature is ~ same. Fraction of total power
(heat from shaft + drag in film) increases with shaft speed.
Oil supply increases
Nmax
5/8 Nmax
3/8 Nmax
1/8 Nmax
shaft speed varies
0.0
0.2
0.4
0.6
0.8
1.0
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 1.1
He
at f
low
ca
rrie
d b
y f
ilm
an
d in
to c
asin
g (-)
Journal speed / Max. journal speed (-)
0.75 Psup* Psup* 1.25 Psup*
Influence of oil supply pressure on heat flows
carried by films
Psup
Oil supply pressure has little effect on heat carried by outer
film and the heat conducted into casing. As shaft speed
increases, a larger supply pressure forces inner film to carry more heat (mainly from shaft).
35
Heat carried by
inner film
Heat into casing
shaft speed varies & 3 oil supply pressures
Heat carried by
outer film
1.25 Psup
0.75 Psup
shaft speed increases
0.5
0.6
0.7
0.8
0.9
1.0
1.1
0.0 0.2 0.4 0.6 0.8 1.0
Tem
pe
ratu
re / T
shaft
Axial coordinate (dimentionless) relative to the bearing length
1/3 Nominal Ci Nominal Ci 4/3 Nominal Ci
36
Influence of clearance on inner film temperature
Inner film temperature grows much
faster for smallest film clearance and
its peak magnitude is ~Tshaft
Small Ci A
portion of shaft
surface receives
heat from inner film.
Likelihood of oil
flashing (burning).
Nominal (P,T)sup
Max shaft speed
Three clearances (small to large)
1/3 Cnom
1 Cnom
4/3 Cnom
Axial length (z)inlet exit
T/Tshaft
0.6
0.7
0.8
0.9
1.0
1 2 3 4 5 6 7 8
Te
mp
era
ture
/ T
shaft
Axial groove number
Average Temperature / Tshaft Peak temperature / Tshaft
37
Influence of # grooves on inner film temperature
Inner film (peak and mean) temperature decreases quickly
with # grooves more flow is drawn.
Nominal (P,T)sup
Max shaft speed
# grooves vary
Average
temperature
Peak temperature
# of grooves
T/Tshaft
38
Influence of clearance on drag power loss and heat from shaft
A larger inner film
clearance allows more
heat drawn from the shaft
but also increases drag
power loss in film (non
intuitive).
Drag power loss
Heat drawn from shaft
Larger ciLarger flow rate
Lower film
temperature
Larger heat from shaft
Larger drag
power loss
Higher oil
viscosity
nominal ci
nominal ci
0.00
0.15
0.30
0.45
0.60
0.75
0.90
Drag power loss (-)
0.90-1.000.75-0.900.60-0.750.45-0.600.30-0.450.15-0.300.00-0.15
39
Influence of # grooves on shear drag power
More grooves more thru flow less temperature
larger oil viscosity more drag power (inner film).
# groovesShaft speed
(min max)2
7
23%
Nominal (P,T)sup
0.6
0.7
0.8
0.9
1.0
1.1
0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5
Tem
pe
ratu
re / T
shaft
Groove depth / Nominal Groove Depth
Average Temperature / Tshaft Peak temperature / Tshaft
40
Influence of groove depth on film temperature
Deeper grooves (oil reserve) more thru flow less film
temperature.
Inner film
Nominal (P,T)sup
Max shaft speed
Average
temperature
Peak temperature
Groove depth increases
T/Tshaft
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5
Dra
g p
ow
er
loss /
To
tal h
ea
t ca
rrie
d b
y f
ilms
Groove depth / Nominal Groove Depth
3/8Ωmax Ωmax
41
Influence of groove depth on power loss
Deeper grooves (oil reserve) less film temperature
and higher oil viscosity more drag power.
Inner film
Nominal (P,T)sup
Max shaft speed
1/8 Max speed
Max Speed
Groove depth increases
0.0
0.2
0.4
0.6
0.8
Drag powerloss (-)
0.6-0.8
0.4-0.6
0.2-0.4
0.0-0.2
42
Influence of groove width on drag power
Narrow grooves more shear drag power loss (inner film).
Effect opposite to that of deep grooves.
Groove widthShaft speed
(min max)
Nominal (P,T)sup
2 nominal
½ nominal
43
Conclusion
THE INFLUENCE OF LUBRICANT
SUPPLY CONDITIONS AND BEARING
CONFIGURATION ON THE
PERFORMANCE OF (SEMI) FLOATING
RING BEARING SYSTEMS FOR
TURBOCHARGERS
44
ConclusionHeat flow (from hot shaft inner film) dominates thermal energy
process: 94% at 1/8 Nmax to 50% at Nmax of total energy transfer. At high shaft
speed, shear drag power loss is as large as heat drawn from shaft.
Shaft temperature influences most the inner film temperature.
Inner film flow carries away 70% or more of total energy, while
heat flow into the casing and thru outer film are small.
Oil supply temperature (75% to 125% nominal) does not affect
inner film temperature field and drag power loss.
Oil supply pressure (3 to 5 bar) large flow rates cools the lower
film; but lower T higher oil viscosity drag power increases while also
drawing heat from shaft.
Large inner film clearance demands of larger flow rate
reduces inner film T increases heat drawn from shaft but also
increases drag power loss.
Multiple parameter analysis allows optimization of (S)FRB system.
GT2017-64839
45
Questions (?)
Tool integrated into sponsor engineering design practice
to predict thermal loading and mechanical stresses and
to ensure lubricant does not overheat (coking).
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Copyright© 2017 Luis San Andres
GT2017-64839
Thanks Honeywell Transportation Systems (HTT)