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MIT OpenCourseWarehttp://ocw.mit.edu

2.61 Internal Combustion EnginesSpring 2008

For information about citing these materials or our Terms of Use, visit: http://ocw.mit.edu/terms.
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2.61 Internal Combustion Engines Lecture 1

Engines

-There are two types of engines:1. Internal combustion - combustion occurs in the working fluid

- open cycle the working fluid is replenished in each cycle- ie) exhaust gas is dumped into the atmosphere

2. External combustion use of heat exchanger to transfer energy to theworking fluid

- Open or closed cycle

- Ex) steam engine, sterling engine

History

1860 Lenoir engine- air and fuel were hand pumped-spark, or ignition was a candle / kerosene lamp done all by hand- operated at about 10 RPM

- 500 sold- 2 stroke-ignition occurs while still in the expansion stage

limited expansion ratiolow efficiency (

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Other Developments

1870 Petroleum industry1888 Pneumatic tires1905 Spark plugs (Champion)1920 Internal Combustion Engine (ICE) takes over steam engine for transportation

- main advantage dont need to carry around water1920-1960 steady development1960 Emission standards start

Heagen Smith smog mechanism1970 Fuel crisis

1980 Global competition1990 Greenhouse gases2000 Fuel and CO2

4 stroke engine

intake exhaustcompression(work in)

expansion(work out)

2 Stroke engine

pressurizedintake

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Engine Size-Piston bore ranges from 1 cm to 1m (large diesels)

-heat loss and friction are surface phenomenon

bigger engine, less losses

Engine GeometryCrank radius aConnecting rod length l

Displacement volume - lB

Vd4

2

=

Compression ratio (geometric) -c

CDR

VVVC

+=

Piston position - 222 sincos)( alas ++=

Instantaneous volume - SB

VV C4

)(2

+=

[ ]5.022 )sin(cos1)1(2

11 ++= RRC

V

VR

C

Piston velocity

a

Rs

=

5.022

2

)sin(2

sinsin)(

where and N=RPMN 2=

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Pressures normally aspirated 4 stroke SI

Heat release normally aspirated 4 stroke SI

Pressure - normally aspirated 4-stroke Diesel

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Lecture 9

SI Engine Combustion

Movie of SI engine combustion

The spark discharge

The SI combustion flame propagation process

Heat release phasing

Heat release analysis from pressure data

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Square piston flow visualization engine

Bore 82.6 mm

Stroke 114.3 mmCompression ratio 5.8

Operating condit ion

Speed 1400 rpm

0.9

Fuel propane

Intake pressure 0.5 bar

Spark timing MBT

Fig. 3.1 in Constanzo, Vincent M. A Visualization Study ofMixture Preparation Mechanisms for Port Fuel Injected

Spark Ignition Engines. Masters Thesis, MIT. June 2004.

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Flame Propagation (Fig 9-14)

1400 rpm

0.5 bar inlet pressure

Image removed due to copyright restrictions. Please see Fig. 9-14 in Heywood, John B.

Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

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Burn duration

Burn duration as CA-deg. : measure of burn progressin cycle

For modern fast-burn engines under medium speed,

part load condition: 0-10% ~ 15

o

0-50% ~ 25o

0-90% ~ 35o

As engine speed increases,burn duration as CA-deg. :

Increases because there is less time per CA-deg. Decreases because combustion is faster due to

higher turbulence

Net effect: increases approximately as rpm0.2

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Spark dischargecharacteristics

Fig.9-39

Schematic of voltage and

current variation with

time for conventional coil

spark-ignit ion system.



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Flame Kernel Development (SAE Paper 880518)

Single cycle flame sequenceFlame from 4 consecutive cycles at f ixed

time after spark

=1,spk= 40oBTC,

1400 rpm, vol. eff. = 0.29

Image removed due to copyright restrictions. Please see Pischinger, Stefan, and John B.

Heywood. A Study of Flame Development and Engine Performance with Breakdown

Ignition Systems in a Visualization Engine.Journal of Engines 97 (February 1988): 880518.

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Energy associated

with Spark Discharge,Combustion and Heat

Loss

SAE Paper 880518

Image removed due to copyright restrictions. Please see Pischinger, Stefan, and John B.

Heywood. A Study of Flame Development and Engine Performance with Breakdown

Ignition Systems in a Visualization Engine.Journal of Engines 97 (February 1988): 880518.

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Schematic of SI engine flame propagation

Fig. 9-4 Schematic of flame propagation in SI engine: unburned gas (U) to left of

flame, burned gas to right. A denotes adiabatic burned-gas core, BL denotesthermal boundary layer in burned gas.

Heat transfer

Work

transfer



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Combustion produced pressure riseFlame

ub

ub

m m

Flame

time t time t + t

1. Pressure is uniform, changing with time

2. For mass m: hb = hu (because dm is allowed to expand against

prevailing pressure)

3. T rise is a function of fuel heating value and mixture composition e.g. at = 1, Tu ~ 700 K, Tb ~ 2800 K

4. Hence burned gas expands: b ~ u ; Vb ~ 4 Vu

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Combustion produced pressure rise5. Since total volume is constrained. The pressure must

rise by p, and all the gas in the cylinder is compressed.

6. Both the unburned gas ahead of flame and burned gas

behind the flame move away from the flame front

7. Both the unburned gas and burned gas temperatures

rise due to the compression by the newly burned gas

8. Unburned gas state: since heat transfer is relatively

small, the temperature is related to pressure byisentropic relationship

Tu/Tu,0 = (p/p0)(

u-1)/

u

9. Burned gas state:

u

Flame

Early burned gas,higher Tb

Later burned gas,

lower Tb

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Thermodynamic

state of charge

Fig. 9-5 Cylinder pressure,

mass fraction burned, and

gas temperatures as

function of crank angle

during combustion.



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Optimum Combustion Phasing

Heat release schedule has to phase correctly withpiston motion for optimal work extraction

In SI engines, combustion phasing controlled by spark

Spark too late heat release occurs far into expansion and workcannot be fully extracted

Spark too early Effectively lowers compression ratio increased heat transfer losses Also likely to cause knock

Optimal: Maximum Brake Torque (MBT) timing MBT spark timing depends on speed, load, EGR, ,temperature, charge motion,

Torque curve relatively flat: roughly 5 to 7oCA retard

from MBT results in 1% loss in torque

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Spark timing effects

Fig. 9-3 (a) Cylinder pressure versus crank angle for overadvanced spark timing

(50o

BTDC), MBT timing (30o

BTDC), and retarded t iming (10o

BTDC). (b) Effectof spark advance on brake torque at constant speed and A/F, at WOT



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Control of spark timing

WOT

Fig. 15-3

Fig. 15-17

Images removed due to copyright restrictions. Please see Fig. 15-3 and 15-17 in Heywood,

John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

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Obtaining combustion information from engine

cylinder pressure data

1. Cylinder pressure affected by:

a) Cylinder volume change

b) Fuel chemical energy release by combustion

c) Heat transfer to chamber wallsd) Crevice effects

e) Gas leakage

2. Obtaining accurate combustion rate information requiresa) Accurate pressure data (and crank angle indexing)

b) Models for phenomena a,c,d,e, above

c) Model for thermodynamic properties of cylinder contents

3. Available methods

a) Empirical methods (e.g. Rassweiler and Withrow SAE800131)

b) Single-zone heat release or burn-rate modelc) Two-zone (burned/unburned) combustion model

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Typical

piezoelectricpressure

transducer spec.

Kistler 6125

6.2mm

Image and data table removed due to copyright restrictions. Please see

http://www.intertechnology.com/Kistler/pdfs/Pressure_Model_6125B.pdf

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Sensitivity of NIMEP to crank angle phase error

SI engine;1500 rpm, 0.38 bar intake pressure

-15

-10

-5

0

5

10

15

-3 -2 -1 0 1 2 3

Crank angle phase error (deg)

Percent error in NIMEP

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Simple heat release analysis

Single zone, perfect gas, idealized model

Other effects:

Crevice effect

Blowby

Real gas properties

Non-uniformities (significant difference between burned and

unburned gas) Unknown residual fraction

v gross ht loss

v

gross

b

f

gross ht loss

1Q

d(mc T) Q Q pV

d

PVmc T1

whence

Mass fraction burned:

pV

Qx

m L

pV Q1

HV

1

= + +

=

=

=

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Cylinder pressure

Fig. 9-10 (a) Pressure-volume diagram; (b) log p-log(V/Vmax) plot; 1500 rpm,MBT timing, IMEP = 5.1 bar, = 0.8, rc = 8.7, propane fuel.



Ad i l

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Burned mass analysis Rassweiler and Winthrow

(SAE 800131)

0

f

u b

1/ nu,0 u

1/ nb,f b

u,0 b,f

b 0 f

During combustion v = v v

Unburned gas volume, back tracked

to spark (0)

V V (p /p )

Burned gas volume, forward tracked

to end of combustion (f)

V V (p /p )

Mass fraction bunred

V Vx 1

V V

+

=

=

= =

1/n 1/n0 0

b

1/n 1/nf f 0 0

Hence, after some algebra

p V p Vx

p V p V

=

Advantage: simple

Need only p(), p0, pfand n

xb

always between 0 and 1

Image removed due to copyright restrictions. Please see Rassweiler, Gerald M., and

Lloyd Withrow. Motion Pictures of Engine Flames Correlated with Pressure Cards.

SAE Transactions 33 (1938): 185-204. Reprinted as SAE Technical Paper 800131.

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Results of heat-release analysis

Fig. 9-12 Results of heat-release analysis showing the effects of heat

transfer, crevices and combustion inefficiency.

Pintake



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Lecture 10

SI Combustion (continue)

SI Engine Knock

Entrainment-and-Burn Model for SI engine combustion

Cycle-to-cycle fluctuation of SI engine heat release

SI engine knock

Spark knock and surface ignition

Spark knock mechanism

Knock chemistry and fuel effects

Knock control

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Model of turbulent combustion

Fig. 14-12

Schematic of entrainment-and-burn model



SI engine flame propagation

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SI engine flame propagation

Entrainment-and-burn model

Rate of entrainment:

Rate at which mixture burns:

= + b

t /e

u f L u f T

dmA S A u (1 e )

dt

Laminar diffusion

through flame front

Turbulent entrainment

L

Tb

b

beLfu

b

S;

mmSA

dt

dm =

+=

Laminar frontal burning Conversion of entrained mass

into burned mass

Critical parameters: uT and T

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Cycle-to-cycle variations

Crank angle (oATDC) Crank angle (oATDC)

Fig. 9-31

Measured cylinder pressure and calculated gross heat-release rate for ten

cycles in a single-cylinder SI engine operating at 1500 rpm, = 1.0, MAP = 0.7

bar, MBT timing 25

o

BTC



C l t l h i b ti h i

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Cycle-to-cycle change in combustion phasing

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SI ENGINE CYCLE-TO-CYCLE VARIATIONS

Phases of combustion1. Early flame development

2. Flame propagation

3. Late stage of burning

Factors affecting SI engine cycle-to-cycle variations:

(a) Spark energy deposition in gas (1)

(b) Flame kernel motion (1)(c) Heat losses from kernel to spark plug (1)

(d) Local turbulence characteristics near plug (1)

(e) Local mixture composition near plug (1)(f) Overall charge components - air, fuel, residual (2, 3)

(g) Average turbulence in the combustion chamber (2, 3)

(h) Large scale features of the in-cylinder flow (3)

(i) Flame geometry interaction with the combustion chamber (3)

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Cycle distributions

Charge variations

Charge andcombustion

phasing

variation

Very Slow-burn cycles Partial burn substantial combustion

inefficiency (10-70%)

Misfire significant combustioninefficiency (>70%)

(No definitive value for threshold)

Fig,. 9-33 (b)

Fig,. 9-36 (b)

Images removed due to copyright restrictions. Please see Fig. 9-33b and 9-36b in Heywood,

John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

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Auto-ignition and knock

1. Knock and surface ignition

2. Knock fundamentals

3. Fuel factor

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Abnormal Combustion:

Knock and surface-Ignition

Image removed due to copyright restrictions. Please see: Fig. 9-58 in Heywood, John B. Internal CombustionEngine Fundamentals. New York, NY: McGraw-Hill, 1988.

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SI Engine Knock

1.

Knock is most critical at WOT and at low speed because of itspersistence and potential for damage.Part-throttle knock is atransient phenomenon and is a nuisance to the driver.

2.

Whether or not knock occurs depends on engine/fuel/vehiclefactors and ambient conditions (temperature, humidity). Thismakes it a complex phenomenon.

3.

To avoid knock with gasoline, the engine compression ratio islimited to approximately 12.5 in PFI engines and 13.5 in DISIengines. Significant efficiency gains are possible if thecompression ratio could be raised. (Approximately, increasing

CR by 1 increases efficiency by one percentage point.)

4.

Feedback control of spark timing using a knock sensor isincreasingly used so that SI engine can operate close to itsknock limit.

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Knock damaged pistons

Images removed due to copyright restrictions. Please see p. 6 in "The Internal Combustion: Modeling Considers All Factors."Lawrence Livermore National Lab, December 1999.

Also see any other photos of knock damage to pistons, such as: http://cameronassociates.org.uk/assets/images/autogen/a_Piston_Damage.jpg

From Lichty, Internal Combustion Engines From Lawrence Livermore website
https://www.llnl.gov/str/pdfs/12_99.1.pdfhttps://www.llnl.gov/str/pdfs/12_99.1.pdfhttp://cameronassociates.org.uk/assets/images/autogen/a_Piston_Damage.jpghttp://cameronassociates.org.uk/assets/images/autogen/a_Piston_Damage.jpghttp://cameronassociates.org.uk/assets/images/autogen/a_Piston_Damage.jpghttp://cameronassociates.org.uk/assets/images/autogen/a_Piston_Damage.jpghttps://www.llnl.gov/str/pdfs/12_99.1.pdf

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End Gas Geometry, Knock, and Pressure Signal(SAE 960827,Stiebels, Schreiber, and Sakak)

Images (12 sapart) of flame

luminescence

and pressuretrace; RON= 90,

2400 rpm, ign. at

35o BTC; white

circle indicating

first autoignition

Images removed due to copyright restrictions. Please see Stiebels, B., et al. "Development of a New Measurement Technique for theInvestigation of End-gas Autoignition and Engine Knock." SAE Transactions105 (February 1996): 960827.

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Knock Fundamentals

Knock originates in the extremely rapid release of much of the fuel

chemical energy contained in the end-gas of the propagating

turbulent flame, resulting in high local pressures. The non-

uniform pressure distribution causes strong pressure waves or

shock waves to propagate across and excites the acoustic modes

of the combustion chamber.

When the fuel-air mixture in the end-gas region is compressed to

sufficiently high pressures and temperatures, the fuel oxidation

processstarting with the pre-flame chemistry and ending with

rapid heat releasecan occur spontaneously in parts or all of the

end-gas region.

Most evidence indicates that knock originates with the auto-

ignition of one or more local regions within the end-gas.

Additional regions then ignite until the end-gas is essentially fully

reacted. The sequence of processes occur extremely rapidly.

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Knock chemical mechanism

CHAIN BRANCHING EXPLOSION

Chemical reactions lead to increasing number of radicals,

which leads to rapidly increasing reaction rates

Formation of Branching AgentsChain Initiation

RO

2 +

RH

ROOH +

R

RH +

O2 R+HO2RO2 RCHO +RO

Chain PropagationDegenerateBranching

R+O2 RO2, etc. ROOH

RO+

OH

RCHO +

O2 RCO +

HO2

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FUEL FACTORS

The auto-ignition process depends on the fuelchemistry.

Practical fuels are blends of a large numberof individual hydrocarbon compounds, each

of which has its own chemical behavior.

A practical measure of a fuels resistance to

knock is the octane number. High octanenumber fuels are more resistant to knock.

Types of hydrocarbons(See text section 3.3)

NAPHTHENES

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PARAFFINS

Butane

Cyclohexane Cyclopentane

OLEFINS

AROMATICS

ISOMERS

Benzene

Cis-2-Butane

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Knock tendency of

individualhydrocarbons

Image removed due to copyright restrictions. Please see: Fig. 9-69 in

Heywood, John B. Internal Combustion Engine Fundamentals.

New York, NY: McGraw-Hill, 1988.

Fig 9-69

Critical compression ratio for

incipient knock at 600 rpm and

450 K coolant temperature for

hydrocarbons

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Fuel anti-knock rating(See table 9.6 for details)

Blend primary reference fuels (iso-octane and normal heptane) so

its knock characteristics matches those of the actual fuel.

Octane no. = % by vol. of iso-octane

Two different test conditions: Research method: 52oC (125oF) inlet temperature, 600 rpm

Motor method: 149oC (300oF) inlet temperature, 900 rpm

Research ON

ON

Motor ON

Sensitivity

Road ON = (RON+MON) /2

Engine

severityLess severe test condition scale

More severe test condition

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Octane requirement

From Balckmore and Thomas, Fuel Economy of the Gasoline Engine, Wiley 1977.

Remark: these are old data; modern engine octane requirement relates more to MON

100

OctaneRequiremen

t

95

90

85OctaneRequirement,RON

7 8 9 10 11

Engine on test stand

103

102

101

10099

98

97

96

95

94

93

92

91

Compression Ratio 90

897.0 8.0 9.0 10.0

Compression Ratio

Represents data from at least ten

cars of the same make and model

As above, from at least five cars

Cars on the road

Slope ~ 5~

Figure by MIT OpenCourseWare. Adapted from Blackmore, D. R., and Thomas, A. Fuel Economy of the Gasoline

Engine: Fuel, Lubricant, and Other Effects. London, England: Macmillan, 1977.

O t R i t I

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Octane Requirement Increase

18

14

10

6

2

50 100 150 200

0

0-2

No additive (ORI = 15)

Deposit controlling additive (ORI = 10)

Clean combustion chamber only

Clean combustion chamber and intake valves

Test 1 (no additive)

Test 2 (with additive)

Test 3 (with additive)

Deposit removal

Octanerequirem

entincrease(ORI)

Hours of operation

ACS Vol. 36, #1, 1991

Figure by MIT OpenCourseWare.

K k t l t t i

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Knock control strategies

1. Provide adequate cooling to the engine

2.

Use intercooler on turbo-charged engines3. Use high octane gasoline

4. Anti-knock gasoline additives

5.

Fuel enrichment under severe condition

6. Use knock sensor to control spark retard so as to

operate close to engine knock limit

7. Fast burn system

8. Gasoline direct injection

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Anti-knock Agents

Alcohols

Methanol CH3OH

Ethanol C2H5OH

TBA (Tertiary Butyl Alcohol) (CH3)3COH

Ethers

MTBE (Methyl Tertiary Butyl Ether) (CH3)3COCH3

ETBE (Ethyl Tertiary Butyl Ether) (CH3

)3

COC2

H5

TAME (Tertiary Amyl Methyl Ether) (CH3)2(C2H5)COCH3

MIT OpenCourseWare

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SI Engine Mixture Preparation

1. Requirements

2. Fuel metering systems

3. Fuel transport phenomena4. Mixture preparation during engine transients

5. The Gasoline Direct Injection engine

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MIXTURE PREPARATION

Fuel

Fuel Air

Metering Metering

Air

EGRMixing Control

EGR

Combustible

Mixture

Engine

MIXTURE PREPARATION

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MIXTURE PREPARATION

Parameters Impact

-Fuel Properties - Driveability

-Air/Fuel Ratio - Emissions-Residual Gas - Fuel Economy

Fraction

Other issues: Knock, exhaust temperature, starting and

warm-up, acceleration/ deceleration transients

Equivalence ratio and EGR strategies

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Equivalence ratio and EGR strategies

(No emissions constrain)

Enrichment to Enrichment to prevent

improve idle stability knocking at high load

Rich=1

Lean for good fuel economyLean

Load

EGR to decrease

NO emission andEGR pumping loss

No EGR toNo EGR to maximize air

improve idle flow for powerstability

Load

R i t f th 3 t l t

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Requirement for the 3-way catalyst

Image removed due to copyright restrictions. Please see: Fig. 11-57 in Heywood, John B.


FUEL METERING

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FUEL METERING

Carburetor A/F not easily controlled

Fuel Injection Electronically controlled fuel metering

Throttle body injection

Port fuel injection

Direct injection

Injectors

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Injectors

PFI injectors

Single 2-, 4-,, up to 12-holes

Injection pressure 3 to 7 bar Droplet size:

Normal injectors: 200 to 80 m

Flash Boiling Injectors: down to 20 m

Air-assist injectors: down to 20 m

GDI injectors

Shaped-spray

Injection pressure 50 to 150 bar

Drop size: 15 to 50 m

PFI Injector targeting

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PFI Injector targeting

Fuel Injector

Intake Valve

70

7

40

Geometrical Issues

Arrangement of fuel injector and intake valve


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Enginemanagement

system

From Bosch Automotive

Handbook

Image removed due to copyright restrictions. Please see anyillustration of an engine control system, such as that in the

Bosch Automotive Handbook. London, England: John Wiley & Sons, 2004.

Fuel Metering

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g

A/F ratio measured by

sensor (closed loop operation)

feedback on fuel amount to keep =1

Feed-forward control (transients): To meter the correct fuel flow for the targeted A/F target, need to

know the air flow Determination of air flow (need transient correction)

Air flow sensor (hot film sensor)

Speed density method

Determine air flow rate from MAP (P) and ambient temperature

(Ta) using volumetric efficiency (v) calibration

Nma = VD

2

v (N,) Displacement vol. VD,rev. per second N,

=PRTa

gas constant R

ENGINE EVENTS DIAGRAM

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ENGINE EVENTS DIAGRAM

Intake Exhaust Injection

BC TC BC TC BC TC BC TC BC

Ign Ign

Cyl.#4TC BC TC BC TC BC TC BC TC

Ign Ign

Cyl.#3

TC BC TC BC TC BC TC BC TC

Ign Ign

Cyl.#2

BC TC BC TC BC TC BC TC BC

Ign Ign

Cyl.#1

0 180 360 540 720 900 1080 1260 1440

Cyl. #1 CA (0 o is BDC compression)

Effect of Injection

Ti i HC

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Timing on HC

Emissions

Engine at 1300 rpm

275 kPa BMEP

Image removed due to copyright restrictions. Please see: Stache, I., and

Alkidas, A. C. "The Influence of Mixture Preparation on the HC ConcentrationHistories from an S.I. Engine Running Under Steady-state Conditions."

SAE Transactions106 (October 1997): 972981.

Injection

timing refers

to start of

injection

Mixture Preparation in PFI engine

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Mixture Preparation in PFI engine

Image removed due to copyright restrictions. Please see: Nogi, Toshiharu, et al. "Mixture Formation of Fuel InjectionSystems in Gasoline Engines." SAE Transactions97 (February 1988): 880558.

Intake flow phenomena in mixture preparation(At low to moderate speed and load range)

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(At low to moderate speed and load range)

Reverse Blow-down Flow

IVO to EVC:

Burned gas flows from exhaust port because Pe>Pi

IVO to Pc = Pi: Burned gas flows from cylinder into intake system until cylinder and

intake pressure equalize

Forward Flow

Pc

= Pito BC:

Forward flow from intake system to cylinder induced by downward piston

motion

Reverse Displacement Flow

BC to IVC:

Fuel, air and residual gas mixture flows from cylinder into intake due to

upward piston motion

Note that the reverse flow affects the mixture preparation

process in engines with port fuel injection

Mixture Preparation in Engine Transients

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Mixture Preparation in Engine Transients

Engine Transients

Throttle Transients

Accelerations and decelerations Starting and warm-up behaviors

Engine under cold conditions

Transients need special compensationsbecause:

Sensors do not follow actual air delivery into cylinder

Fuel injected for a cycle is not what constitutes the

combustible mixture for that cycle

Manifold pressure charging in thrott le transient

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p g g

Image removed due to copyright restrictions. Please see: Aquino, C. F. "Transient A/F Control Characteristics of the 5

Liter Central Fuel Injection Engine." SAE Transactions90 (February 1981): 810494.

Fuel-Lag in Throttle Transient

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g

The x-Model

dMf=

xmf Mf

dt

mc =(1- x)mf +Mf

mf =Injected fuel flow rate

mc =Fuel delivery rate

to cylinder

Mf =Fuel mass inpuddle

mf

.

xmf

.

Mf/

Mf


Fuel transient in throttle opening

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p g

Image removed due to copyright restrictions. Please see Fig. 7-28 in Heywood, John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

Fig 7-28

Uncompensated A/F behavior in throttle transient

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Engine start

up behavior

2.4 L, 4-cylinder

Image removed due to copyright restrictions. Please see: Fig. 1 in Santoso, Halim, and Cheng, Wai K. engine

"Mixture Preparation and Hydrocarbon Emissions in the First Cycle of SI Engine Cranking."

SAE Journal of Fuels and Lubricants111 (October 2002): 2002-01-2805. Engine starts

with Cyl#2

piston in mid

stroke of

compression

Firing order1-3-4-2

Pertinent Features of DISI Engines

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1. Precise metering of fuel into cylinder Engine calibration benefit: better driveability and

emissions2. Opportunity of running stratified lean at part load

Fuel economy benefit (reduced pumping work; lowercharge temperature, lower heat transfer; better

thermodynamic efficiency)3. Charge cooling by fuel evaporation

Gain in volumetric efficiency

Gain in knock margin (could then raise compression

ratio for better fuel economy) Both factors increase engine output

Toyota DISI Engine (SAE Paper 970540)

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Images removed due to copyright restrictions. Please see: Harada, Jun, et al. "Development of Direct-injection

Gasoline Engine." SAE Journal of Engines106 (February 1997): 970540.

Charge cooling by in-air fuel evaporation

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Images removed due to copyright restrictions. Please see: Anderson, R. W., et al. "Understanding the Thermodynamics of

Direct-injection Spark-ignition (DISI) Combustion Systems: An Analytical and Experimental Investigation." SAE Journal of Engines105 (October 1996): 962018.

Full load performance benefit

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Images removed due to copyright restrictions. Please see: Iwamoto, Y., et al. "Development of

Gasoline Direct Injection Eengine." SAE Journal of Engines106 (February 1997): 970541.

Part load fuel economy gain

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Images removed due to copyright restrictions. Please see Kume, T., et al. "Combustion Control Technologies for

Direct Injection SI Engine." SAE Journal of Engines105 (February 1996): 960600.

DISI Challenges

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1. High cost

2. With the part-load stratif ied-charge concept :

High hydrocarbon emissions at light load

Significant NOx emission, and lean exhaust not amenable to3-way catalyst operation

3. Particulate emissions at high load

4. Liquid gasoline impinging on combustion chamber walls

Hydrocarbon source Lubrication problem

5. Injector deposit

Special fuel additive needed for injector cleaning

6. Cold start behavior

Insufficient fuel injection pressure

Wall wetting

http://ocw.mit.edu/http://ocw.mit.edu/

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http://ocw.mit.edu/termshttp://ocw.mit.edu/terms

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Diesel Engine Combustion

1. Characteristics of diesel combustion

2. Different diesel combustion systems

3. Phenomenological model of dieselcombustion process

4. Movie of combustion in diesel systems

5. Combustion pictures and planar laser

sheet imaging

DIESEL COMBUSTION PROCESS

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PROCESS

Liquid fuel injected into compressed charge

Fuel evaporates and mixes with the hot airAuto-ignition with the rapid burning of the fuel-

air that is premixed during the ignition delay

periodPremixed burning is fuel rich

As more fuel is injected, the combustion is

controlled by the rate of diffusion of air into theflame

DIESEL COMBUSTION PROCESS

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NATURE OF DIESEL COMBUSTION

Heterogeneous liquid, vapor and air

spatially non-uniform

turbulent

diffusion flame

The Diesel Engine

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Intake air not throttled

Load controlled by the amount of fuel injected

>A/F ratio: idle ~ 80

>Full load ~19 (less than overall stoichiometric)

No end-gas; avoid the knock problem

High compression ratio: better efficiency

Combustion:

Turbulent diffusion flame Overall lean

Diesel as the Most Efficient Power Plant

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Theoretically, for the same CR, SI engine has higher f; butdiesel is not limited by knock, therefore it can operate at

higher CR and achieves higher f Not throttled - small pumping loss

Overall lean - higher value of - higher thermodynamicefficiency

Can operate at low rpm - applicable to very large engines

slow speed, plenty of time for combustion

small surface to volume ratio: lower percentage of parasiticlosses (heat transfer and friction)

Opted for turbo-charging

Large Diesels: f~ 55%

~ 98% ideal efficiency !

Disadvantages of Diesel Engines

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Cold start difficulty

Noisy - sharp pressure rise: cracking noise

Inherently slower combustion

Lower power to weight ratio

Expensive components

NOx and particulate matters emissions

Diesel Engine Characteristics

(compared to SI engines)

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Better fuel economy

Overall lean, thermodynamically efficient

Large displacement, low speed lower FMEP

Higher CR

> CR limited by peak pressure, NOx emissions, combustion andheat transfer loss

Turbo-charging not limited by knock: higher BMEP over domain of

operation, lower relative losses (friction and heat transfer)

Lower Power density Overall lean: would lead to smaller BMEP

Turbocharged: would lead to higher BMEP

> not knock limited, but NOx limited

> BMEP higher than SI engine

Lower speed: overall power density (P/VD) not as high as SI engines

Emissions: more problematic than SI engine

NOx: needs development of efficient catalyst

PM: regenerative and continuous traps

Applications

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Small (7.5 to 10 cm bore; previously mainly IDI; new

ones are high speed DI)

passenger cars Medium (10 to 20 cm bore; DI)

trucks, trains

Large (30 to 50 cm bore; DI) trains, ships

Very Large (100 cm bore)

stationary power plants, ships

Common Direct-Injection Compression-Ignition Engines(Fig. 10.1 of text)

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Image removed due to copyright restrictions. Please see Fig. 10-1 in Heywood, John B. Internal Combustion Engine

Fundamentals.New York, NY: McGraw-Hill, 1988.

(a) Quiescent chamber with multihole nozzle typical of larger engines

(b) Bowl-in-piston chamber with swirl and multihole nozzle; medium to small size engines

(c) Bowl-in-piston chamber with swirl and single-hole nozzle; medium to small size engines

Common types of small Indirect-injection diesel engines(Fig. 10.2 of text)

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Image removed due to copyright restrictions. Please see Fig. 10-2 in Heywood, John B. Internal Combustion EngineFundamentals.New York, NY: McGraw-Hill, 1988.

(a) Swirl prechamber (b) Turbulent prechamber

Common Diesel Combustion Systems (Table 10.1)

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Image removed due to copyright restrictions. Please see Table 10-1 in Heywood, John B. Internal Combustion

Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

Typical Large Diesel Engine Performance Diagram

140

Max Pressure120

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Max Pressure

Scavenge Air Pressure (gauge)

Exh. Temp, Turbine Inlet and Outlet

Specific air quantity

Specific fuel consumption

Compression

Pressure

120100

80

60

40

20

2.5

2.0

1.5

1.0

0.5

(g/kWh)

(kg/kWh)

(oC)

(bar)

(bar)

Sulzer RLB 90 - MCR 1

Turbo-charged 2-stroke Diesel

1.9 m stroke; 0.9 m bore

Rating:0 Speed: 102 Rev/ min

500

Piston speed 6.46 m/s

BMEP: 14.3 bar

Configurations

4 cyl: 11.8 MW (16000 bhp)

5 cyl: 14.7 MW (20000 bhp)

6 cyl: 17.7 MW (24000 bhp)

450

400

350

300

250

200

13

12

11

10

9

7 cyl: 20.6 MW (28000 bhp) 8

7 8 cyl: 23.5 MW (32000 bhp) 210

9 cyl: 26.5 MW (36000 bhp)

10 cyl: 29.4 MW (40000 bhp)

205

200

195

190

12 cyl: 35.3 MW (48000 bhp) 185180

4 6 10 14

8

12

BMEP (bar)

16

Diesel combustion processdirect injection

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1)

Ignition delayno significant

heat release

2)

Premixed rapid combustion

3)

Mixing controlled phase of

combustion

4)

Late combustion phase

Note:

(2) is too fast;

(4) is too slow

Rate of Heat Release in Diesel Combustion(Fig. 10.8 of Text)

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Image removed due to copyright restrictions. Please see Fig. 10-9 in Heywood, John B. Internal CombustionEngine Fundamentals. New York, NY: McGraw-Hill, 1988.

A Simple Diesel Combustion Concept (Fig. 10-8)

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Image removed due to copyright restrictions. Please see Fig. 10-8 in Heywood, John B. Internal Combustion



2.61 Internal Combustion Engines

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Diesel injection, ignition, and fuel air mixing

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1. Fuel spray phenomena

2. Spontaneous ignition

3. Effects of fuel jet and charge motion on mixing-

controlled combustion

4. Fuel injection hardware5. Challenges for diesel combustion

DIESEL FUEL INJECTION

The fuel spray serves multiple purposes:

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The fuel spray serves multiple purposes: Atomization

Fuel distribution

Fuel/air mixingTypical Diesel fuel injector Injection pressure: 1000 to 2200 bar

5 to 20 holes at ~ 0.15 - 0.2 mm diameter

Drop size 0.1 to 10 m For best torque, injection starts at about 20o BTDC

Injection strategies for NOx control

Late injection (inj. starts at around TDC)

Other control strategies:Pilot and multiple injections, rate shaping, water emulsion

Diesel Fuel Injection System

(A Major cost of the diesel engine)

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(A Major cost of the diesel engine) Performs fuel metering

Provides high injection pressure

Distributes fuel effectively Spray patterns, atomization etc.

Provides fluid kinetic energy for charge mixing

Typical systems: Pump and distribution system (100 to 1500 bar)

Common rail system (1000 to 1700 bar)

Hydraulic pressure amplification Unit injectors (1000 to 2500 bar)

Piezoelectric injectors (to 1800 bar)

Electronically controlled

EXAMPLE OF DIESEL INJECTION

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(Hino K13C, 6 cylinder, 12.9 L turbo-charged diesel

engine, rated at 294KW@2000 rpm)

Injection pressure = 1400 bar; duration = 40oCA

BSFC 200 g/KW-hr

Fuel delivered per cylinder per injection at rated

condition

0.163 gm ~0.21 cc (210 mm3)

Averaged fuel flow rate during injection

64 mm3/ms

8 nozzle holes, at 0.2 mm diameter

Average exit velocity at nozzle ~253 m/s

Fuel Atomization Process

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Liquid break up governed by balance between

aerodynamic force and surface tension

gasu2

d

Webber Number (Wb ) =

Critical Webber number: Wb,critical ~ 30; diesel fuel

surface tension ~ 2.5x10-2 N/m

Typical Wb at nozzle outlet > Wb,critical; fuel shattered

into droplets within ~ one nozzle diameter

Droplet size distribution in spray depends on further

droplet breakup, coalescence and evaporation

Droplet size distribution

f(D) Size distribution:f(D)dD b bilit f f i di

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f(D)dD = probability of f inding

particle with diameter in

the range of (D, D + dD)

1=

f(D)dD

D 0

Average diameter Volume distribution

1 dV f(D) D

3

D =

f(D) D dDV dD =

0

f(D)D3dD0

Sauter Mean Diameter (SMD)

f (D) D 3dD

D 32 = 0

f (D) D 2dD0

Droplet Size Distribution

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Fig. 10.28 Droplet size distribution measured well downstream; numbers on the curves are

radial distances from jet axis. Nozzle opening pressure at 10 MPa; injection into air at 11 bar.

Droplet Behavior in Spray

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Small drops (~ micron size) follow gas stream;

large ones do not

Relaxation time d2

Evaporation time d2

Evaporation time small once charge is ignited

Spray angle depends on nozzle geometry and

gas density : tan(/2) (gas/liquid)

Spray penetration depends on injectionmomentum, mixing with charge air, and droplet

evaporation

Spray Penetration: vapor and liquid (Fig. 10-20)

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Shadowgraph image

showing both liquid

and vapor penetration



Back-lit image

showing liquid-

containing core

Auto-ignition Process

PHYSICAL PROCESSES (Physical Delay)

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Drop atomization

Evaporation

Fuel vapor/air mixing

CHEMICAL PROCESSES (Chemical Delay)

Chain initiation

Chain propagationBranching reactions

CETANE IMPROVERS

Alkyl Nitrates 0.5% by volume increases CN by ~10

Ignition Mechanism: similar to SI engine knock


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Chemical reactions lead to increasing number of radicals,

which leads to rapidly increasing reaction rates

Formation of Branching Agents

ChainInitiation RO2 +RH ROOH +R

RH +

O2 R +HO2 RO2 RCHO +RO

ChainPropagation DegenerateBranching

R +O2 RO2,etc. ROOH RO +OHRCHO +

O2 RCO +

HO2

Cetane Rating

(Procedure is simi lar to Octane Rating for SI Engine; for details,

see10 6 2 of text)

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see10.6.2 of text)

Primary Reference Fuels:

Normal cetane (C16H34): CN = 100

Hepta-Methyl-Nonane (HMN; C16H34): CN = 15

(2-2-4-4-6-8-8 Heptamethylnonane)

Rating:

Operate CFR engine at 900 rpm with fuel

Injection at 13o BTC

Adjust compression ratio until ignition at TDC

Replace fuel by reference fuel blend and change blend proportion to

get same ignition point

CN = % n-cetane + 0.15 x % HMN

Ignition Delay

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Igniti on delays measured in a

small four-stroke cycle DI

diesel engine with rc=16.5, as aImage removed due to copyright restrictions. Please see Fig. 10-36 in Heywood, John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988. function of load at 1980 rpm, at

various cetane number

(Fig. 10-36)

Fuel effects on Cetane Number (Fig. 10-40)

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Ignition Delay Calculations

Difficulty: do not know local conditions (species concentrationand temperature) to apply kinetics information

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a d e pe a u e) o app y e cs o a o

Two practical approaches:

Use an instantaneous delay expression

(T,P) = P-nexp(-EA/ T)

and solve ignition delay (id) from

1=

tsi

+id

1

dttsi (T(t),P(t)) Use empirical correlation of id based on T, P at an appropriate

charge condition; e.g. Eq. (10.37 of text)

1 1

21.2 0.63

id(CA) =

(0.36 +

0.22Sp(m/s))expEA(

R~

T(K)

17190)) +

(P(bar) 12.4

)

EA (Joules per mole) = 618,840 / (CN+25)

Diesel Engine Combustion

Air Fuel Mixing Process

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Importance of air utilization

Smoke-limit A/F ~ 20

Fuel jet momentum / wall interaction has a larger influenceon the early part of the combustion process

Charge motion impacts the later part of the combustion

process (after end-of-injection)

CHARGE MOTION CONTROL

Intake created motion: swirl, etc.

Not effective for low speed large engine

Piston created motion - squish

Interaction of fuel jet and the chamber wall

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Sketches of outer vapor boundary

of diesel fuel spray from 12

successive frames (0.14 ms apart)Image removed due to copyright restrictions. Please see Fig. 10-21 in

of high-speed shadowgraphHeywood, John B. Internal Combustion Engine Fundamentals. New York,NY: McGraw-Hill, 1988. movie. Injection pressure at 60

MPa.

Fig. 10-21

Interaction of fuel jet with air swirl

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Schematic of fuel jet air swirl interaction;

is the fuel equivalenceImage removed due to copyright restrictions. Please see Fig. 10-22in

ratio distributionHeywood, John B. Internal Combustion Engine Fundamentals. New York,

NY: McGraw-Hill, 1988.

Fig. 10-22

Rate of Heat Release in Diesel Combustion(Fig. 10.8 of Text)

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DIESEL FUEL INJECTION HARDWARE

High pressure system

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High pressure system

precision parts for flow control

Fast action high power movements

Expensive system

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CHALLENGES IN DIESEL COMBUSTION

Heavy Duty Diesel Engines

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Heavy Duty Diesel Engines NOx emission

Particulate emission Power density

Noise

High Speed Passenger Car Diesel Engines

All of the above, plus

Fast burn rate



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Diesel Emissions and Control

Diesel emissions

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Diesel emissions

Regulatory requirements Diesel emissions reduction

Diesel exhaust gas after-treatment

systems

Clean diesel fuels

Diesel Emissions

CO not significant until smoke-limit is reachedOverall fuel lean

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higher CR favors oxidation

HC not significant in terms of mass emission

Crevice gas mostly air Significant effects:

Odor

Toxics (HC absorbed in fine PM)

Mechanisms:Over-mixing, especially during light load

Sag volume effect

NOx very important

No attractive lean NOx exhaust treatment yet PM very important

submicron particles health effects

Demonstration of over-mixing effect

Diesel HC

emissionmechanisms

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Images removed due to copyright restrictions. Please see: Fig. 11-35 and 11-36 in Heywood, John B.

Internal Combustion Engine Fundamentals.New York, NY: McGraw-Hill, 1988.

Effect of nozzle sac vol. on HC emissions

NOx mechanisms

NO: Extended Zeldovich mechanismN2 + O NO + N

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N2 ONO N

N + O2NO + O

N + OH

NO + H Very temperature sensitive: favored at high temperature

Diffusion flame: locally high temperature

More severe than SI case because of higher CR

NO2 : high temperature equilibrium favors NO, but NO2 isformed due to quenching of the formation of NO by mixingwith the excess air

NO + HO2

NO2 + OH

NO2 + ONO + O2

Gets 10-20% of NO2 in NOx

NOx formation in Diesel engines

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Images removed due to copyright restrictions. Please see: Fig. 11-15 and 11-16 in Heywood, John B.


Normalized NO concentration fromcylinder dumping experiment. NOx and NO emissions as a funct ion of

Injection at 27o BTC. Note most of the overall equivalence ratio . Note that NO2

NO is formed in the diffusion phase of as a fraction of the NOx decreases with

burning increase of .

Diesel combustion

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Image removed due to copyright restrictions. Please see: Flynn, Patrick F., et al. "Diesel Combustion: An Integrated View Combining Laser

Diagnostics, Chemical Knetics, and Empirical Validation." SAE Journal of Engines108 (March 1991): SP-1444.

Particulate Matter (PM)

As exhaust emission:

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visible smoke

collector of organic and inorganic materialsfrom engine

Partially oxidized fuel; e.g. Polycyclic Aromatic

Hydrocarbons (PAH)Lubrication oil (has Zn, P, Cu etc. in it)

Sulfates (fuel sulfur oxidized to SO2, and

then in atmosphere to SO3 which hydratesto sulfuric acid (acid rain)

Particulate Matter

In the combustion process, PM formed

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initially as soot (mostly carbon)

partially oxidized fuel and lub oil condenseon the particulates in the expansion,

exhaust processes and outside the engine

PM has effective absorption surface area of200 m2/g

Soluble Organic Fraction (SOF) 10-30%

(use dichloromethane as solvent)

Elementary soot particle structure

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Source: Environmental Protection Agency, www.epa.gov.

PM formation processes

NucleationDehydrogenation

Oxidation
http://www.epa.gov/http://www.epa.gov/

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Surface growth

Agglomeration

Adsorption,

condensation

Dehydrogenation

Oxidation

Time

DehydrogenationOxidation

In-cylinder

Inatmosp

here

Diesel NOx/PM regulation

1

US

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0.01

0.1

2007

1991-93

1990

PM

(g/bhp-hr)

1994

1998

2004

US

Euro II (1998)

Euro III(2000)

Euro IV(2005)

Euro V(2008)Euro VI (proposed-2013)

EU

0.1 1 10

NOx (g/bhp-hr)

(Note: Other count ries regulations are originally in terms of g/KW-hr)

Diesel Emissions Reduction

1. Fuel injection: higher injection pressure; multiple

pulses per cycle, injection rate shaping; improved

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pu ses pe cyc e, jec o a e s ap g; p o ed

injection timing control

2.

Combustion chamber geometry and air motionoptimization well matched to fuel injection system

3. Exhaust Gas Recycle (EGR) for NOx control

Cooled for impact4. Reduced oil consumption to reduce HC contribution

to particulates

5. Exhaust treatment technology: NOx, PM

6. Cleaner fuels

Effect of EGR

1.35 L single cylinder engine,

Direct Injection, 4-stroke

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Images removed due to copyright restrictions. Please see: Uchida, Noboru, et al. "Combined Effects of EGR and Supercharging on Diesel

Combustion and Emissions." SAE Journal of Engines102 (March 1993): 930601.

Split Injection

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Images removed due to copyright restrictions. Please see: Nehmer, D. A., and Reitz, R. D. "Measurement of the Effect of Injection

Rate and Split Injections on Diesel Engine Soot and NOx Emissions." SAE Journal of Engines103 (February 1994): 940668.

PM Control

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Images removed due to copyright restrictions. Please see: Zelenka, P., et al. "Ways Toward the Clean

Heavy-duty Diesel." SAE Journal of Engines 99 (February 1990): 900602.

Post injection filter regeneration

Regeneration needs ~550oC

Normal diesel exhaust under city

d i i 150 200 C

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Image removed due to copyright restrictions. Please see: Fig. 8 in Salvat, O., et al.

"Passenger Car Serial Application of a Particulate Filter System on a Common

Rail Direct Injection Diesel Engine." SAE Journal of Fuels and Lubricants109

(March 2000): SP-1497.

Increase exhaust gas temperature by injection of

additional fuel pulse late in cycle.

driving ~150-200oC

Need oxidation catalyst (CeO2) to

lower light off temperature

Control engine torque

Minimized fuel penalty

Peugeot SAE 2000-01-0473

Diesel particulate filters use porous ceramics

and catalyst to collect and burn the soot

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Please see slide 9 in Johnson, Tim. "Diesel Exhaust Emission Control." Environmental Monitoring, Evaluation, and Protection in

New York: Linking Science and Policy, 2003.

State-of-the Art SCR system has NO2 generation and

oxidation catalyst to eliminate ammonia slip

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Image removed due to copyright restrictions. Please see p. 9 in "Recent Developments in Integrated Exhaust Emission

Control Technologies Including Retrofit of Off-Road Diesel Vehicles." Manufacturers of Emissions Controls Association,

February 3, 2000.

Integrated DPF and NOx trap
http://www.arb.ca.gov/msprog/offroad/techreview/MECA-ci.pdfhttp://www.arb.ca.gov/msprog/offroad/techreview/MECA-ci.pdfhttp://www.arb.ca.gov/msprog/offroad/techreview/MECA-ci.pdfhttp://www.arb.ca.gov/msprog/offroad/techreview/MECA-ci.pdf

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Image removed due to copyright restrictions. Please see: Fig. 3 in Nakatani, Koichiro, et al. "Simultaneous PM and NOx

Reduction System for Diesel Engines." SAE Journal of Fuels and Lubricants111 (March 2002): SP-1674.

From Toyota SAE Paper 2002-01-0957

Clean Diesel Fuels

1. Lower sulfur levels

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350 ppm15 ppm

2. Lower percentage aromatics

3. Oxygenated fuels

4. Higher cetane number5. Narrower distillation range

Diesel Emission Control

Summary

Emission regulations present substantial

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Emission regulations present substantial

challenge to Diesel engine system Issues are:

performance and sfc penalty

cost

reliability

infra-structure support




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Engine Heat Transfer

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1. Impact of heat transfer on engine operation

2. Heat transfer environment

3. Energy flow in an engine

4. Engine heat transfer Fundamentals Spark-ignition engine heat transfer

Diesel engine heat transfer

5. Component temperature and heat flow

Engine Heat Transfer

Heat transfer is a parasitic process that

contributes to a loss in fuel conversion

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efficiency

The process is a surface effect

Relative importance reduces with:

Larger engine displacement

Higher load

Engine Heat Transfer: Impact

Efficiency and Power: Heat transfer in the inlet decrease volumetricefficiency. In the cylinder, heat losses to the wall is a loss of

availability.

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Exhaust temperature: Heat losses to exhaust influence theturbocharger performance. In- cylinder and exhaust system heat

transfer has impact on catalyst light up.

Friction: Heat transfer governs liner, piston/ ring, and oiltemperatures. It also affects piston and bore distortion. All of these

effects influence friction. Thermal loading determined fan, oil and

water cooler capacities and pumping power.

Component design: The operating temperatures of critical engine

components affects their durability; e.g. via mechanical stress,

lubricant behavior

Engine Heat Transfer: Impact

Mixture preparation in SI engines: Heat transfer to the fuel

significantly affect fuel evaporation and cold start calibration

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Cold start of diesel engines: The compression ratio of dieselengines are often governed by cold start requirement

SI engine octane requirement: Heat transfer influences inlet

mixture temperature, chamber, cylinder head, liner, piston andvalve temperatures, and therefore end-gas temperatures, which

affect knock. Heat transfer also affects build up of in-cylinder

deposit which affects knock.

Engine heat transfer environment

Gas temperature: ~300 3000oK

Heat flux to wall: Q /A

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Materials limit:

Cast iron ~ 400oC

Aluminum ~ 300oC

Liner (oil film) ~200oC

Hottest components Spark plug > Exhaust valve > Piston crown > Head

Liner is relatively cool because of limited exposure to burned

gas

Source Hot burned gas

Radiation from particles in diesel engines

Energy flow diagram for an IC engine

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Energy flow distribution for SI and Diesel

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Image removed due to copyright restrictions. Please see Table 12-1 in Heywood, John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

Energy distribution in SI engine

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Images removed due to copyright restrictions. Please see: Fig. 12-4 in Heywood, John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

Heat transfer process in engines

Areas where heat transfer is important

Intake system: manifold, port, valves

In cylinder: cylinder head piston valves liner

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In-cylinder: cylinder head, piston, valves, liner

Exhaust system: valves, port, manifold, exhaust pipe Coolant system: head, block, radiator

Oil system: head, piston, crank, oil cooler, sump

Information of interest

Heat transfer per unit time (rate)

Heat transfer per cycle (often normalized by fuel heating

value)

Variation with time and location of heat flux (heat transferrate per unit area)

Schematic of temperature distribution and heat flow across

the combustion chamber wall (Fig. 12-1)

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Image removed due to copyright restrictions. Please see: Fig. 12-1 in Heywood, John B.Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988.

Combustion Chamber Heat Transfer

Turbulent convection: hot gas to wall

.Q =Ahg(Tg Twg )

Conduction through wall

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.

Q =A (Twg Twc)tw

Turbulent convection: wall to coolant.

Q =Ahc (Twc Tc )

Overall heat transfer.

Q = Ah(T g Tc )

Overall thermal resistance: three resistance in series

1 1 t 1= + w +h hg hc ( alum ~180 W/m-k

cast iron ~ 60 W/m-k

stainless steel ~18 W/m-k)

Turbulent Convective Heat Transfer Correlation

Approach: Use Nusselt- Reynolds number correlations similar tothose for turbulent pipe or flat plate flows.

e.g. In-cylinder:

Nu =hL

= a(Re) 0 .8

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Nu = = a(Re)

h = Heat transfer coefficient

L = Characteristic length (e.g. bore)

Re= Reynolds number, UL/

U = Characteristic gas velocity

= Gas thermal conductivity

= Gas viscosity

= Gas density

a = Turbulent pipe flow correlation coefficient

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IC Engine heat transfer

Heat transfer mostly from hot burned gasThat from unburned gas is relatively small

Flame geometry and charge motion/turbulence

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Flame geometry and charge motion/turbulence

level affects heat transfer rate

Order of MagnitudeSI engine peak heat flux ~ 1-3 MW/m2

Diesel engine peak heat flux ~ 10 MW/m2

For SI engine at part load, a reduction inheat losses by 10% results in animprovement in fuel consumption by 3%Effect substantially less at high load

SI Engine Heat Transfer

Heat transfer dominated by that

from the hot burned gas

Burned gas wetted area determine

b li d / fl t

Unburned Zone

Burned Zone

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Heat transfer

by cylinder/ flame geometry

Gas motion (swirl/ tumble) affects

heat transfer coefficient

Burned zone: sum over area wetted Qb = Aci,bhb(Tb Tw,i)by burned gas i

Unburned zone: sum over area Qu = Aci,uhu(Tu Tw,i) wetted by unburned gas i

Note: Burned zone heat f lux >> unburned zone heat f lux

Acij

Cooling Surface Area


SI engine heat transfer environment

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Fig. 14-9 5.7 L displacement, 8 cylinder engine at WOT, 2500 rpm; fuel equivalence

ratio 1.1; GIMEP 918 kPa; specif ic fuel consumption 24 g/kW-hr.

SI engine heat flux

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Images removed due to copyright restrictions. Please see: Gilaber, P., and P. Pinchon. "Measurements and Multidimensional

Modeling of Gas-wall Heat Transfer in a S.I. Engine." SAE Journal of Engines97 (February 1988): 880516.

Heat transfer scaling

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Nu correlation: heat transfer rate 0.8N0.8

Time available (per cycle) 1/N

Fuel energy BMEP

Thus Heat Transfer/Fuel energy BMEP-0.2N-0.2

Diesel engine heat transfer

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Fig. 12-13 Measured surface heat fluxes at di fferent locations in cylinder head andliner of naturally aspirated 4-stroke DI diesel engine. Bore=stroke=114mm; 2000

rpm; overall fuel equivalence ratio = 0.45.

Diesel engine radiative heat transfer

Fig. 12-15Radiant heat flux as

Image removed due to copyright restrictions. Please see: Fig. 12-15 in Heywood, John B. fraction of total heat flux

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g py g g y

Internal Combustion Engine Fundamentals. New York, NY: McGraw-Hill, 1988. over the load range ofseveral different diesel

engines

Heat transfer effect on component temperatures

Temperature distribution in head

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Fig. 12-20 Variation of cylinder head temperature with measurement location n SI

engine operating at 2000 rpm, WOT, with coolant water at 95oC and 2 atmosphere.

Heat transfer paths from piston

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Fig. 12-24 Heat outflow form various zones of piston as percentage of heat flow in

from combustion chamber. High-speed DI diesel engine, 125 mm bore, 110 mm

stroke, CR=17

Piston Temperature Distribution

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Figure 12-19

Isothermal contours (solid lines) and heat flow paths (dashed lines) determined from measured

temperature distribution in piston of high speed DI diesel engine. Bore 125 mm, stroke 110

mm, rc=17, 3000 rev/min, and full load

Thermal stress

Simple 1D example : column constrained at ends

T2>T1 induces

compressionStress-strain relationship

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Stress strain relationshipT1 stress

x=[x-(y+z)]/E + (T2-T1)

REAL APPLICATION - FINITE ELEMENT ANALYSIS

Complicated 3D geometry

Solution to heat flow to get temperature distribution Compatibility condition for each element

Heat Transfer Analysis

Example of Thermal

Stress Analysis:Piston

Design

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Images removed due to copyright restrictions. Please see Castleman, Jeffrey L. "Power Cylinder Design Variables and Their

Effects on Piston Combustion Bowl Edge Stresses." SAE Journal of Engines102 (September 1993): 932491.

Thermal-Stress-Only

Loading Structural Analysis

Power Cylinder Design

Variables and TheirEffects on Piston

Combustion Bowl Edge

Stresses

J. Castleman, SAE 932491

Heat Transfer Summary

1.

Magnitude of heat transfer from the burned gas much greater than inany phase of cycle

2. Heat transfer is a significant performance loss and affects engine

operation

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Loss of available energy

Volumetric efficiency loss

Effect on knock in SI engine

Effect on mixture preparation in SI engine cold start

Effect on diesel engine cold start

3. Convective heat transfer depends on gas temperature, heat transfer

coefficient, which depends on charge motion, and transfer area,

which depends on flame/combustion chamber geometry

4. Radiative heat transfer is smaller than convective one, and it is only

significant in diesel engines


2.61 Internal Combustion Engines

Spring 2008


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Engine Friction and Lubrication

Engine friction

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terminology

Pumping loss

Rubbing friction loss

Engine Friction: terminology

Pumping work: Wp Work per cycle to move the working fluid through the engine

Rubbing friction work: Wrf

Accessory work: Wa

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Total Friction work: Wtf = Wp + Wrf + Wa

Normalized by cylinder displacement MEP

tfmep = pmep + rfmep + amep

Net output of engine

bmep = imep(g) tfmep

Mechanical efficiency m = bmep / imep(g)

Friction components

1.

Crankshaft friction Main bearings, front and rear bearing oil seals

2. Reciprocating friction

C ti d b i i t bl

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Connecting rod bearings, piston assembly3. Valve train

Camshafts, cam followers, valve actuation mechanisms

4. Auxiliary components

Oil, water and fuel pumps, alternator5. Pumping loss

Gas exchange system (air filter, intake, throttle, valves,

exhaust pipes, after-treatment device, muffler)

Engine fluid flow (coolant, oil)

Engine Friction

Fig. 13-1

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Comparison of major categories of

Image removed due to copyright restrictions. Please see: Fig. 13-1 in frict ion losess: fmep at dif ferentHeywood, John B. Internal Combustion Engine Fundamentals. loads and speeds for 1.6 L four-New York, NY: McGraw-Hill, 1988.

cylinder overhead-cam automotive

Spark Ignit ion (SI) and

Compression-Ignition (CI) engines.

Pumping loss

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Fig. 13-15 Puming loop diagram for SI engine under f iring

conditions, showing throttling work Vd(pe-p i), and valve flow work

Sliding friction mechanism

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Energy dissipation processes:

Detaching chemical binding between surfaces Breakage of mechanical interference (wear)

Bearing Lubrication

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Stribeck Diagram

for journal bearing

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Motoring break-down analysis

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Fig. 13-14

Motored fmep versus engine speed for engine breakdown tests.

(a) Four-cyl inder SI engine.

(b) Average results for several four- and six-cylinder DI diesel engines

Breakdown of engine mechanical fr iction

1 F.A. Martin, Friction in Internal Combustion

Engines, I.Mech.E. Paper C67/85, Combust ion

Engines Frict ion and Wear, pp.1-17,1985.

T. Hisatomi and H. Iida, Nissan Motor Companys

New 2.0 L. Four-cylinder Gasoline Engine, SAE

Trans. Vol. 91, pp. 369-383, 1982; 1st engine.

2nd engine

18

19Piston + Rod

Piston

Rings

RingsTypical

4800 r/min

4800 r/min

Full Load

Rod

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2nd engine.

M. Hoshi, Reducing Friction Losses in Automobile

Engines, Tribology International, Vol. 17, pp 185-

189, Aug. 1984.

J.T. Kovach, E.A. Tsakiris , and L.T. Wong, Engine

Friction Reduction for Improved Fuel Economy,

SAE Trans. Vol. 91, pp. 1-13, 1982

19

20

21

Mechanical Friction (%)

Rings + Piston + Rod

RodRings + Piston

Piston + RodRings

Motoringr/min

Motoringr/min

4800 r/min

Full Load

6000

4000

2000

2000

4000

Valvetrain

Crankshaft

100806040200


Valve train friction

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Image removed due to copyright restrictions. Please see illustrations of "Valve Timing-gear Designs."

In the Bosch Automotive Handbook. London, England: John Wiley & Sons, 2004.

Valve train friction depends on: Total contact areas

Stress on contact areas

Spring and inertia loads

Low friction valve train

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Valve train friction reduction

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Engine speed (x1000 rpm)

Friction loss reduction by new lighter valve train system,

JSAE Review 18 (1977), Fukuoka, Hara, Mori, and Ohtsubo

Courtesy of Elsevier, Inc., http://www.sciencedirect.com. Used with permission.

Piston ring pack
http://www.sciencedirect.com/http://www.sciencedirect.com/

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Hydrodynamiclubrication of the

piston ring

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Friction force and associated power loss

150

100

50Force(N)

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0

0

800

600

400

200

TDC TDC TDCBDC BDC

Crank Angle

Intake Compression Expansion Exhaust

Power(N

-m/s)


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Bore distortion

4thOrder

Cylinder Distortion

2ndOrder 2ndOrder 3rdOrder

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1

4 Order 2 Order

2ndOrder

3rdOrder 4thOrder

2 Order 3 Order

2 3 4

Three orders of bore distortion.

Top deck of hypothetical engine.


Lubricants

Viscosity is a strong function of temperature

Multi-grade oils (introduced in the 1950s)

Temperature sensitive polymers to stabilize

viscosity at high temperatures

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y g p

Cold: polymers coiled and inactive

Hot: polymers uncoiled and tangle-up:

suppress high temperature thinning

Stress sensitivity: viscosity is a function of

strain rate

Viscosity

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Image removed due to copyright restrictions. Please see: Linna, Jan-Roger, et al. "Contribution of Oil Layer Mechanism to the

Hydrocarbon Emissions from Spark-ignition Engines." SAE Journal of Fuels and Lubricants106 (October 1997): 972892.

Modeling of engine friction

Overall engine friction model:

tfmep (bar) = fn (rpm, Vd, , B, S, .)

See text, ch. 13, ref.6; SAE 900223, )

Detailed model

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Detailed model

tfmep = (fmep)components

With detailed modeling of component friction as a function of rpm, load,

FMEP distribution

Image removed due to copyright restrictions. Please see: Patton, Kenneth J., et al. "Development and Evaluation of a Friction

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Model for Spark-ignition Engines." SAE Journal of Engines98 (February 1989): 890836.

Distr ibution of FMEP for a 2.0L I-4 engine; B/S = 1.0, SOHC-rocker arm, flat

fol lower, 9.0 compression ratio

C = crankshaft and seals

R = reciprocating components

V = valve train components

A = Auxiliary components

P = Pumping loss




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Engine Turbo/Super Charging

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Super and Turbo-charging

Why super/ turbo-charging?

Fuel burned per cycle in an IC engine is air limited

(F/A)stoich = 1/14.6

f,v fuel conversion and volumetricefficienciesfm QHV

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fm f QHV mf fuel mass per cycleTorq = QHV fuel heating value2nR nR 1 for 2-stroke, 2 for 4-stroke engine

N revolution per second

Power =

Torq

2N VD engine displacementa,0 air densityFmf =(A)Va,0VD

Super/turbo-charging: increase air density

Super- and Turbo- Charging

Purpose: To increase the charge density Supercharge: compressor powered by engine output

No turbo-lag

Does not impact exhaust treatment

Fuel consumption penalty

T b h d b h t t bi

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Turbo-charge: compressor powered by exhaust turbine Uses wasted exhaust energy

Turbo- lag problem

Affects exhaust treatment

Intercooler Increase charge density (hence output power) by cooling the

charge

Lowers NOx emissions

Exhaust-gas turbocharger for trucks

1.Compressor housing, 2. Compressor

impeller, 3. Turbine housing, 4. Rotor, 5.

Bearing housing, 6. inflowing exhaust gas, 7.

Charge-air pressure regulation with Out-flowing exhaust gas, 8. Atmospheric fresh

wastegate on exhaust gas end. 1.Engine,air, 9. Pre-compressed fresh air, 10. Oil inlet,

2. Exhaust-gas turbochager, 3. Wastegate 11. Oil return

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Images removed due to copyright restrictions. Please see illustrations of "Charge-air Pressure Regulation with Wastegate on Exhaust

Gas End", and "Exhaust-gas Turbocharger for Trucks." In the Bosch Automotive Handbook. London, England: John Wiley & Sons, 2004.

From Bosch Automotive Handbook

Compressor: basic thermodynamics

Compressor efficiency c

W = Widealc Wactual

T

Wideal =mcpT1

T

T

2

1

1

1

1

2

m

P2

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1

T2 =P2

T1

P1

1 1 P2

Wactual = mcpT1 1c P1

Wactuals T2 =T1 +

mcp

P1

1

22

Ideal

process

Actual

process

Turbine: basic thermodynamics

Turbine efficiency t4

W

t =

W

actual

Wideal3

mWideal =mcpT3

1

T

T

4

3

T 1

P3

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T4 =P4

T3

P3

1

P Wactual = tmcpT3

1

P3

4

Wactuals T4 =T3

mcp

P4

4

3

4

Ideal

process

Actual

process

Properties of Turbochargers

Power transfer between fluid and shaft RPM3

Typically operate at ~ 60K to 120K RPM

RPM limited by centrifugal stress: usually tip

velocity is approximately sonic

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Flow devices, sensitive to boundary layer (BL)

behavior

Compressor: BL under unfavora

Documents

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