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1 CEFRC2 June 27, 2012 Reciprocating Internal Combustion Engines Prof. Rolf D. Reitz, Engine Research Center, University of Wisconsin-Madison 2012 Princeton-CEFRC Summer Program on Combustion Course Length: 9 hrs (Wed., Thur., Fri., June 27-29) Hour 2 Copyright ©2012 by Rolf D. Reitz. This material is not to be sold, reproduced or distributed without prior written permission of the owner, Rolf D. Reitz.

Reciprocating Internal Combustion Engines Lecture...Reciprocating Internal Combustion Engines ... Isentropic compressible flow theory ... operation is limited by

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1 CEFRC2 June 27, 2012

Reciprocating Internal Combustion Engines

Prof. Rolf D. Reitz,Engine Research Center,

University of Wisconsin-Madison

2012 Princeton-CEFRCSummer Program on Combustion

Course Length: 9 hrs

(Wed., Thur., Fri., June 27-29)

Hour 2

Copyright ©2012 by Rolf D. Reitz.This material is not to be sold, reproduced or distributed withoutprior written permission of the owner, Rolf D. Reitz.

2 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Short course outine:

Engine fundamentals and performance metrics, computer modeling supportedby in-depth understanding of fundamental engine processes and detailedexperiments in engine design optimization.

Day 1 (Engine fundamentals)Hour 1: IC Engine Review, 0, 1 and 3-D modelingHour 2: Turbochargers, Engine Performance MetricsHour 3: Chemical Kinetics, HCCI & SI Combustion

Day 2 (Spray combustion modeling)Hour 4: Atomization, Drop Breakup/CoalescenceHour 5: Drop Drag/Wall Impinge/VaporizationHour 6: Heat transfer, NOx and Soot Emissions

Day 3 (Applications)Hour 7: Diesel combustion and SI knock modelingHour 8: Optimization and Low Temperature CombustionHour 9: Automotive applications and the Future

Turbocharging

3 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Improved

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-strokesfor effective scavenging:

- intake P > exhaust P- crankcase used as a pump

Some engines combine engine-driven andmechanical (e.g., in two-stage configuration).

Intercooler after compressor- controls combustion air temperature.

4 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Turbocharging

Energy in exhaust is used to driveturbine which drives compressor

Wastegate used to by-pass turbine

Charge air cooling after compressorfurther increases air density- more air for combustion

5 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Regulated Two-Stage Turbocharger

Duplicated Configuration per Cylinder Bank

EGR Cooler

EGR Cooler

EGR Valve

EGR Valve

LP stage Turbo-Chargerwith Bypass

LP stage Turbo-Chargerwith Bypass

HP stage Turbocharger

HP stage Turbocharger

Regulating valve

Regulating valveCharge AirCooler

Charge AirCooler

CompressorBypass

CompressorBypass

LP TURBINE

Regulating Valve

LP Stage Bypass

HP TURBINE Compressor Bypass

GT-Power R2S Turbo Circuit

6 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Variable Valve Timing - IVC control

ln V

Q

ln T

ln V

TDC IVC

T ign

Q

Reduced Peak Temp (NOx)Improved phasing

IVC

IVC

VPP V

Isentropic

( 1)IVC

IVC

VTT V

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

ln P

TDC IVC

Pressure/time ofignition

Boost

Compressor

7 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Centrifugal compressor typically used inautomotive applications

Provides high mass flow rate atrelatively low pressure ratio ~ 3.5

Rotates at high angular speeds- direct coupled with exhaust-driventurbine- less suited for mechanicalsupercharging

Consists of:stationary inlet casing,rotating bladed impeller,stationary diffuser (w or w/o vanes)collector - connects to intake system

Automotive Compressor

8 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

201

1

11

2T

MT

2 10

11

1(1 )

2P

MP

P0 P=Pb

P/P0Pb

0

1

x

0.528

reservoir ambient

M=1Manifold pressure, P2 cmHg

m

Choked

WOT

speed

Ex. Flow past throttle plate

Choked flow for P2 < 53.5 kPa = 40.1cmHg

40.1 76

1

Isentropic compressible flow theory

0

9 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Anderson, 1990

Application to turbomachinery

Reduced flow passagearea

P0 /PTotal/static pressure ratio

1/0.528=1.891.0

Choked flow

Increased speed

0

0

/

/ref

ref

m T T

P P

Variable Geometry Compressor/turbine performance map

“Corrected massflow rate”

A measure of effective flowarea

1*2( 1)

1 00

2( )1Mm P A

RT

1/ 20 0 0

0

( / ) /( / )

P Vm AV A RT

RT c

P AM P P T TRT

Fliegner’s Formula:

10 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Anderson, 1990

P0

P3T

S

P1

P2

P03

= P0,in

= Pout

V12/2cP

Air at stagnation state 0,in accelerates toinlet pressure, P1, and velocity V1.

Compression in impeller passagesincreases pressure to P2, and velocity V2.

Diffuser between states 2 and out,recovers air kinetic energy at exit of impellerproducing pressure rise to, Pout andlow velocity Vout

Compressor

1

0,

1

a

aa

c a out in

a P in out

c in

W m h h

m c T pW

p

Note: use exit static pressure and inlet totalpressure, because kinetic energy of gasleaving compressor is usually not recovered

c

)()(

inout

inisenoutc TT

TT

Heywood, Fig. 6-43

11 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Heywood, 1988

12 CEFRC2 June 27, 2012

Compressor MapsWork transfer to gas occurs in impeller via change in gasangular momentum in rotating blade passage

Surge limit line– reduced mass flowdue to periodic flowreversal/reattachment inpassage boundary layers.Unstable flow can leadto damage

At high air flow rate,operation is limited bychoking at the minimumarea point within compressor

Pressure ratio evaluatedusing total-to-staticpressures since exit flowkinetic energy is notrecovered

Non-dimensionalize bladetip speed (~ND) by speedof sound

Speed/pressure limit line

Supersonic flow

Shockwave

Heywood, Fig. 6-46

Hour 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)

PressureRatio (t/t)

GM 1.9L diesel engine

13 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Serrano, 2007

Turbochargers

out

Radial flow – automotive;axial flow – locomotive, marine

0

3

0

3

0

3

TTN

N

pp

TT

mm

corrected

gcorrected

T

S

P1

P2

P03

P0 = P0,in

P3 = Pout

V12/2cP

)()(

inisenout

inoutt TT

TT

14 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

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

678

9

Pamb

6’

Naturally aspirated:Pintake=Pexhst=Patm (5-7-8-9-1)Boosted operation:Negative pumping work:P7<P1 – but hurts scavenging

6’’

CompressorTurbine

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

0,1

g

goutt g P in t

in

PW m c T

P

15 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Reitz & Hoag, 2007

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

16 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

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)

Com

pres

sorP

ress

ure

Rat

io

Torque Peak (1700rpm)

Trq Peak Operating Pnt

Rated (2300rpm)

Rated Operating Pnt

Parameter Torque Peak Rated UnitsHorsepower 48 69 hp

BSFC 0.377 0.401 lb/hp-hrA/F 23.8 24.5 none

17 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

1

1

3

4

1

3

1

2 111

a

a

g

g

pp

m

mTCpTCp

pp

mechct

air

fuel

a

g

18 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Wt = Wc

. .

Heywood, 1988

Maximum possible closed-cycleefficiency (“ideal efficiency”)

State (1) to (2) isentropic(i.e., adiabatic and reversible)compression from max (V1) tomin cylinder volume (V2)Compression ratio rc = V1/V2.

State (2) to (3) adiabaticand isochoric (constant volume)combustion,State (3) to (4) isentropicexpansion.

State (4) to (1) exhaust process- available energy is rejected- can be converted to mechanicalor electrical work:

Ideal Engine Efficiency – Otto cycle

19 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Heywood, 1988

ηideal Function of only two variables, compression ratio (rc)and ratio of specific heats (γ)

Increasing rc increases operating volume for compression and expansionIncreasing γincreases pressure rise during combustion and increases workextraction during expansion stroke.

Both effects result in an increase in net system work for a given energy releaseand thereby increase engine efficiency.

Actual closed-cycle efficiencies to deviate from ideal:

1.) Assumption of isochoric 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 stresses2.) Assumption of calorically perfect fluid:

Specific heats decrease with increasing gas temperature; species conversion duringcombustion causes γto decrease

3.) Adiabatic assumption:Large temperature gradient near walls results in energy being lost to heat transferrather than being converted to crank work

20 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Other Assumptions:

In engine system models, compressors, supercharger, turbines modeled withconstant 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 zeropressure drop, a fixed effectiveness of 0.9, constant coolant temperature of 350 K.

21 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

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 specifiedinitial conditions at IVC of pressure (p0), temperature (T0), and composition(xn,0 for all species considered - N2, O2, Ar, CO2, and H2O) and initial trappedmass (m0), including trapped residual mass

Post-combustion composition determined assuming complete combustion ofdelivered fuel mass.

Minor species resulting from dissociation during combustion not considered

Herold, SAE 2011-01-2216

22 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Combustion model - Wiebe function

Heat transfer model - Woschni

First law energy balance: du=dq+pdv

Combustion:

Wall heat transfer:

23 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Herold, SAE 2011-01-2216

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

24 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

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 RCCISCOTEresults(Exp/Sim)

{1 }PMEP FMEPBTE GIEIMEPg

45

55

Chen-Flynn SAE 650733).

Lavoie, 2012

1-D modeling for engine performance analysis – Lavoie et al. (2012)

25 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Mid load

26 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Lavoie, 2012

Burn duration Heat transfer

Friction

Turbocharger equation

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

27 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Woshni, 1967

Lavoie, 2012

Effect of combustion phasing on efficiency

Constant volume combustion

Without HT: Best efficiency CA50~TDCWith HT: best efficiency with CA50~10 deg – tradeoff between heat loss/late expansion

28 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Cum

ulat

ive

heat

rele

ase

Crank angle

10-90 Burn

CA50

10%

50%

90%

100%

Lavoie, 2012

63%

0 air standard efficiency

Adiabatic

Decreasing Incomplete combustion

Energy budget

29 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Lavoie, 2012

Fuel-to-charge equivalence ratio, ’

Bur

ned

gas

tem

pera

ture

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

Effect of dilution

30 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Lavoie, 2012

Effect of boost on efficiency

Reduced heat transfer loss

Reduced friction losses

31 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Lavoie, 2012

Potential brake efficiencies for naturally aspirated engines

Increased pumping losses

32 CEFRC2 June 27, 2012

Hour 2: Turbochargers, Engine Performance Metrics

Lavoie, 2012

Summary

33 CEFRC2 June 27, 2012

Turbocharging can increase engine efficiency by using available energy in exhaustand by reducing pumping work

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

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 thecombustion details

Hour 2: Turbochargers, Engine Performance Metrics