Reciprocating Internal Combustion Engines two-stage turbocharger Duplicated Configuration per Cylinder Bank EGR Cooler EGR Cooler EGR Valve EGR Valve LP stage Turbo-Charger with Bypass

1 CEFRC1-2 , 2014

Reciprocating Internal Combustion Engines

Prof. Rolf D. Reitz

Engine Research Center

University of Wisconsin-Madison

2014 Princeton-CEFRC

Summer School on Combustion

Course Length: 15 hrs

(Mon.- Fri., June 23 – 27, 2014)

Copyright ©2014 by Rolf D. Reitz.

This material is not to be sold, reproduced or distributed without

prior written permission of the owner, Rolf D. Reitz.

Part 2: Turbochargers, Engine Performance Metrics

2 CEFRC1-2, 2014

Short course outine:

Engine fundamentals and performance metrics, computer modeling supported

by in-depth understanding of fundamental engine processes and detailed

experiments in engine design optimization.

Day 1 (Engine fundamentals)

Part 1: IC Engine Review, 0, 1 and 3-D modeling


Day 2 (Combustion Modeling)

Part 3: Chemical Kinetics, HCCI & SI Combustion

Part 4: Heat transfer, NOx and Soot Emissions

Day 3 (Spray Modeling)

Part 5: Atomization, Drop Breakup/Coalescence

Part 6: Drop Drag/Wall Impinge/Vaporization/Sprays

Day 4 (Engine Optimization)

Part 7: Diesel combustion and SI knock modeling

Part 8: Optimization and Low Temperature Combustion

Day 5 (Applications and the Future)

Part 9: Fuels, After-treatment and Controls

Part 10: Vehicle Applications, Future of IC Engines


3 CEFRC1-2, 2014

Turbocharging

Improved


Pulse-driven turbine was invented and

patented in 1925 by Büchi to increase

the amount of air inducted into the engine.

- Increased engine power more than offsets

losses due to increased back pressure

- Need to deal with turbocharger lag

Turbocharging

Purpose of turbocharging or supercharging is to increase inlet air density,

- increase amount of air in the cylinder.

Mechanical supercharging

- driven directly by power from engine.

Turbocharger - connected compressor/turbine

- energy in exhaust used to drive turbine.

Supercharging necessary in two-strokes

for effective scavenging:

- intake P > exhaust P

- crankcase used as a pump

Some engines combine engine-driven and

mechanical (e.g., in two-stage configuration).

Intercooler after compressor

- controls combustion air temperature.

4 CEFRC1-2, 2014


Turbocharging

Energy in exhaust is used to drive

turbine which drives compressor

Wastegate used to by-pass turbine

Charge air cooling after compressor

further increases air density

- more air for combustion

5 CEFRC1-2, 2014


Regulated two-stage turbocharger

Duplicated Configuration per Cylinder Bank

EGR Cooler

EGR Cooler

EGR Valve

EGR Valve

LP stage Turbo-Charger

with Bypass

LP stage Turbo-Charger

with Bypass

HP stage Turbo

charger

HP stage Turbo

charger

Regulating valve

Regulating valve Charge Air

Cooler

Charge Air

Cooler

Compressor

Bypass

Compressor

Bypass

LP TURBINE

Regulating Valve

LP Stage Bypass

HP TURBINE Compressor Bypass

GT-Power R2S Turbo Circuit

6 CEFRC1-2, 2014


Intercooler for IVC temperature control

ln V

Q

ln T

ln V

TDC IVC

Tign

Q

Reduced Peak Temp (NOx)

Improved phasing

IVC

IVC

VP

P V

Isentropic

( 1)

IVC

IVC

VT

T V

Boost explains 20% of the improved fuel efficiency of diesel vs. SI

ln P

TDC IVC

Pressure

/time of

ignition

Boost

Compressor

7 CEFRC1-2, 2014


Centrifugal compressor typically used in

automotive applications

Provides high mass flow rate at

relatively low pressure ratio ~ 3.5

Rotates at high angular speeds

- direct coupled with exhaust-driven

turbine

- less suited for mechanical

supercharging

Consists of:

stationary inlet casing,

rotating bladed impeller,

stationary diffuser (w or w/o vanes)

collector - connects to intake system

Automotive compressor

8 CEFRC1-2, 2014


Compressible flow – A review

/Tds dh dp

dh VdV

Gibbs

Energy

Euler dP VdV

2

2

(1 )dA MdP

A V

0d dA dV

A V

AV Const

2( 1)dA dV

MA V

for M<1 for M>1

Subsonic nozzle Subsonic diffuser Supersonic diffuser Supersonic nozzle

dA<0 dA >0 dA <0 dA >0

from AV dV>0 dV <0 dV <0 dV >0 from Euler dP<0 dP >0 dP >0 dP <0

kinetic energy pressure recovery kinetic energy

Traffic flow behaves like a supersonic flow!

9 CEFRC1-2, 2014

Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990

Area-velocity relations

Model passages as compressible flow in converging-diverging nozzles

A*/A

0 P/P0

0 1

1

0.528

Subsonic Supersonic

0 ∞ M 1

reservoir throat exit

2 solutions for

same area

P0

A*

1

*2( 1)

1 0

0

2( )

1Mm P A

RT

With M=1: Fliegner’s formula

1/ 2

0 0 0

0

( / ) /( / )

P Vm AV A RT

RT c

P AM P P T TRT

Choked flow, M=1

Minimum area point

1/ 21

1 11

*

0 0

2 1( ) 1 ( )

1 2

A P P

A P P

1

2( 1)2

*

1 2 ( 1)(1 )

1 2

AM

A M

Area Mach number relations

10 CEFRC1-2, 2014


201

1

11

2

TM

T

2 101

1

1(1 )

2

PM

P

P0 P=Pb

P/P0 Pb

0

1

x

0.528

reservoir ambient

M=1 Manifold pressure, P1 cmHg

m

Choked

WOT

y

Ex. Flow past throttle plate

Choked flow for P2 < 53.5 kPa = 40.1cmHg

40.1 76

1

Isentropic nozzle flows

y

0

11 CEFRC1-2, 2014


P1 P0

Application to turbomachinery

Reduced flow passage

area

P0 /P Total/static pressure ratio

1/0.528=1.89 1.0

Choked flow

Increased speed

0

0

/

/

ref

ref

m T T

P P

Variable Geometry Compressor/

turbine performance map

“Corrected mass

flow rate”

A measure of effective flow

area

1

*2( 1)

1 0

0

2( )

1Mm P A

RT

Fliegner’s Formula:

12 CEFRC1-2, 2014


P 0

P 3 T

S

P 1

P 2

P 0 3

= P 0,in

= P out

V 1 2 / 2 c P

Air at stagnation state 0,in accelerates to

inlet pressure, P1, and velocity V1.

Compression in impeller passages

increases pressure to P2, and velocity V2.

Diffuser between states 2 and out,

recovers air kinetic energy at exit of impeller

producing pressure rise to, Pout and

low velocity Vout

Compressor

1

0,

1

a

a

a

c a out in

a P in out

c in

W m h h

m c T pW

p

c

)(

)(

inout

inisenoutc

TT

TT

Heywood, Fig. 6-43

Heywood, 1988

13 CEFRC1-2, 2014


Note: use exit static pressure and inlet total

pressure, because kinetic energy of gas

leaving compressor is usually not recovered

Compressor maps Work transfer to gas occurs in impeller via change in gas

angular momentum in rotating blade passage

Surge limit line

– reduced mass flow

due to periodic flow

reversal/reattachment in

passage boundary layers.

Unstable flow can lead

to damage At high air flow rate,

operation is limited by

choking at the minimum

area point within compressor Pressure ratio evaluated

using total-to-static

pressures since exit flow

kinetic energy is not

recovered

Non-dimensionalize blade

tip speed (~ND) by speed

of sound

Speed/pressure limit line

Supersonic flow

Shock

wave

Heywood, Fig. 6-46

14 CEFRC1-2, 2014

Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988

Compressor maps

0.5

0.6

0.7

0.8

0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 0.18

Corrected Air Flow (kg/s)

Efficiency

(T/T)

35000 40000 50000 70000

90000 110000 130000 150000

170000 180000 190000

35000 4000050000

70000

90000

110000

130000

150000

170000

180000

190000

1.0

1.2

1.4

1.6

1.8

2.0

2.2

2.4

2.6

2.8

3.0

0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 0.18

Corrected Air Flow (kg/s)

Pressure

Ratio (t/t)GM 1.9L diesel engine

Serrano, 2007

15 CEFRC1-2, 2014


P

V

TDC BDC

Pexhst

Pintake

Compression

Expansion

Available work

(area 5-6-7)

Blowdown

Automotive turbines

P-V diagram showing available exhaust energy

- turbocharging, turbocompounding, bottoming cycles and

thermoelectric generators further utilize this available energy

1

2

3 4

5

6

7 8

9

Pamb 6’

Naturally aspirated:

Pintake=Pexhst=Patm (5-7-8-9-1)

Boosted operation:

Negative pumping work:

P7<P1 – but hurts scavenging

6’’

Compressor

Turbine

0,( )t g in outW m h h 1

0,1

g

gout

t g P in t

in

PW m c T

P

Reitz, 2007

16 CEFRC1-2, 2014


Turbochargers

out

Radial flow – automotive;

axial flow – locomotive, marine

0

3

0

3

0

3

TT

NN

pp

TT

mm

corrected

gcorrected

T

S

P 1

P 2

P 0 3

P 0 = P 0,in

P 3 = P out

V 1 2 / 2 c P

)(

)(

inisenout

inoutt

TT

TT

17 CEFRC1-2, 2014


Compressor selection

To select compressor, first determine engine breathing lines.

The mass flow rate of air through engine for a given pressure ratio is:

= IMP = PR * atmospheric pressure (no losses)

= IMT = Roughly constant for given Speed

18 CEFRC1-2, 2014


Engine breathing lines

Engine Breathing Lines1.4L Diesel, Air-to-Air AfterCooled, Turbocharged

1

1.2

1.4

1.6

1.8

2

2.2

2.4

2.6

2.8

3

3.2

3.4

3.6

3.8

0.000 1.000 2.000 3.000 4.000 5.000 6.000 7.000 8.000 9.000 10.000 11.000 12.000 13.000 14.000

Intake Mass Flow Rate (lb/min)

Co

mp

resso

r P

ressu

re R

ati

o

Torque Peak (1700rpm)

Trq Peak Operating Pnt

Rated (2300rpm)

Rated Operating Pnt

Parameter Torque Peak Rated Units

Horsepower 48 69 hp

BSFC 0.377 0.401 lb/hp-hr

A/F 23.8 24.5 none

19 CEFRC1-2, 2014


1

1

3

4

1

3

1

2 111

a

a

g

g

p

p

m

m

TCp

TCp

p

pmechct

air

fuel

a

g

Wt = Wc

Heywood, 1988

20 CEFRC1-2, 2014


. .

Maximum possible closed-cycle

efficiency (“ideal efficiency”)

State (1) to (2) isentropic

(i.e., adiabatic and reversible)

compression from max (V1) to

min cylinder volume (V2)

Compression ratio rc = V1/V2.

State (2) to (3) adiabatic

and isochoric (constant volume)

combustion,

State (3) to (4) isentropic

expansion.

State (4) to (1) exhaust process

- available energy is rejected

- can be converted to mechanical

or electrical work:

Ideal engine efficiency – Otto cycle

21 CEFRC1-2, 2014


Ideal engine efficiency – Otto cycle

T

s

1

2

3

4

Otto

)/()]()[( 231243 TTTTTT

)/()(1 2314 TTTT

However,

1 1 12 1 1 2 4 3 3 4/ ( / ) ( / ) /cT T V V r V V T T

1/11 cr1.25

1.3

=1.4

0.2

0.4

0.6

0.8

8 16 24 0 rc

Efficiency = net work / energy supplied

Wexpansion

Wcompression

22 CEFRC1-2, 2014


ηideal Function of only two variables, compression ratio (rc)

and ratio of specific heats (γ)

Increasing rc increases operating volume for compression and expansion

Increasing γ increases pressure rise during combustion and increases work

extraction during expansion stroke.

Both effects result in an increase in net system work for a given energy release

and thereby increase engine efficiency.

Actual closed-cycle efficiencies to deviate from ideal:

1.) Assumption of isochoric (constant volume) combustion:

Finite duration combustion in realistic engines. Kinetically controlled combustion has shorter combustion duration than diesel or SI

- duration limited by mechanical constraints, high pressure rise rates with audible

engine noise and high mechanical stresses

2.) Assumption of calorically perfect fluid: Specific heats decrease with increasing gas temperature; species conversion during

combustion causes γ to decrease

3.) Adiabatic assumption: Large temperature gradient near walls results in energy being lost to heat transfer

rather than being converted to crank work

23 CEFRC1-2, 2014


Other assumptions:

In engine system models, compressors, supercharger, turbines modeled with

constant isentropic efficiency instead of using performance map. - typically, compressors, superchargers, and fixed geometry turbines have isentropic

efficiencies of 0.7. VGT has isentropic efficiency of 0.65.

Charge coolers - intercooler, aftercooler, and EGR cooler modeled with zero

pressure drop, a fixed effectiveness of 0.9, constant coolant temperature of 350 K.

24 CEFRC1-2, 2014

Part 2: Turbochargers, Engine Performance Metrics Herold, 2011

Zero-dimensional closed-cycle analysis:

Combustion represented as energy addition to closed system

Fuel injection mass addition from user-specified start of injection crank angle

(θSOI) and injection duration (Δθinj).

Pressure and mass integrated over the closed portion of cycle with specified

initial conditions at IVC of pressure (p0), temperature (T0), and composition

(xn,0 for all species considered - N2, O2, Ar, CO2, and H2O) and initial trapped

mass (m0), including trapped residual mass

Post-combustion composition determined assuming complete combustion of

delivered fuel mass.

Minor species resulting from dissociation during combustion not considered

Herold, 2011

25 CEFRC1-2, 2014


Combustion model - Wiebe function

Heat transfer model - Woschni

First law energy balance: de=dq - Pdv

Combustion:

Wall heat transfer:

26 CEFRC1-2, 2014

Part 2: Turbochargers, Engine Performance Metrics Herold, 2011

Friction model

Chen-Flynn model ( SAE 650733).

FMEP = C + (PF*Pmax) + (MPSF*Speedmp)

+ (MPSSF*Speedmp2)

where: C = constant part of FMEP (0.25 bar)

PF = Peak Cylinder Pressure Factor (0.005)

Pmax = Maximum Cylinder Pressure

MPSF = Mean Piston Speed Factor (0.1)

MPSSF = Mean Piston Speed Squared Factor (0)

Speedmp = Mean Piston Speed

BTE*LHV=IMEPg-PMEP-FMEP

DOE goal BTE=55%

Engine brake thermal efficiency BTE

0 5 10 15 20 25 3020

30

40

50

60

70

Load -- Gross IMEP [bar]

BT

E [%

]

GIE = 55%

GIE = 60%

GIE = 65%

PMEP = 0.4 barFMEP = 1 bar

150 bar PCP Limit

UW Dyno limit

UW RCCI

SCOTE

results

(Exp/Sim)

{1 }PMEP FMEP

BTE GIEIMEPg

45

55

Chen-Flynn, 1965

27 CEFRC1-2, 2014


Lavoie, 2012

1-D modeling for engine performance analysis

28 CEFRC1-2, 2014


Mid load

29 CEFRC1-2, 2014

Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012

Burn duration Heat transfer

Friction

Turbocharger equation

m~0.8, Re increases with Bore and (boost)

Woshni, 1967

30 CEFRC1-2, 2014


Effect of combustion phasing on efficiency

Constant volume combustion

Without HT: Best efficiency CA50~TDC

With HT: best efficiency with CA50~10 deg – tradeoff between heat loss/late expansion C

um

ula

tive h

eat re

lease

Crank angle

10-90 Burn

CA50

10%

50%

90%

100%

31 CEFRC1-2, 2014


63%

F0 air standard efficiency

Adiabatic

Decreasing Incomplete combustion

Energy budget

32 CEFRC1-2, 2014


Fuel-to-charge equivalence ratio, f’

Bu

rned

gas

tem

per

atu

re

f ranges from 0.2 to 1 with air, EGR ranges from 0 to 80% with f=1

Effect of dilution

33 CEFRC1-2, 2014


Effect of boost on efficiency

Reduced heat transfer loss

Reduced friction losses

34 CEFRC1-2, 2014


Potential brake efficiencies of naturally aspirated engines

Increased pumping losses

35 CEFRC1-2, 2014


Summary

Turbocharging can increase engine efficiency by using available energy in exhaust

and by reducing pumping work

Air standard “ideal cycle” analysis provides a bound on engine efficiency

estimates.

0-D engine system models provide estimates of engine system efficiencies,

if combustion details (e.g., timing and duration) and heat transfer losses are assumed

The goal of multi-dimensional models (to be discussed next) is to predict the

combustion details

36 CEFRC1-2, 2014


Reciprocating Internal Combustion Engines two-stage turbocharger Duplicated Configuration per Cylinder Bank EGR Cooler EGR Cooler EGR Valve EGR Valve LP stage Turbo-Charger with Bypass

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