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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
3 CEFRC1-2, 2014
Turbocharging
Improved
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 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-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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
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
Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
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
Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 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
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 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)
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
. .
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
Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988
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
Part 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 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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics
Lavoie, 2012
1-D modeling for engine performance analysis
28 CEFRC1-2, 2014
Part 2: Turbochargers, Engine Performance Metrics
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
Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012
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
Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012
63%
F0 air standard efficiency
Adiabatic
Decreasing Incomplete combustion
Energy budget
32 CEFRC1-2, 2014
Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012
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
Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012
Effect of boost on efficiency
Reduced heat transfer loss
Reduced friction losses
34 CEFRC1-2, 2014
Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012
Potential brake efficiencies of naturally aspirated engines
Increased pumping losses
35 CEFRC1-2, 2014
Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012
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
Part 2: Turbochargers, Engine Performance Metrics
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