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).

http://rotorlab.tamu.eduLearn more at:

Copyright© 2017 Luis San Andres

GT2017-64839

Thanks Honeywell Transportation Systems (HTT)