IC Engine

Reciprocating Internal Combustion Engines

Michele MannoDepartment of Industrial EngineeringUniversity of Rome Tor Vergata

Last update: May 15th, 2014

Internal Combustion Engines 1

Contents

1. General remarks and engine classification

2. Main operating parameters

3. Air intake

4. Supercharging and turbocharging

5. Fuel metering in spark ignition engines

6. Fuel injection in compression ignition engines

7. Operating characteristics and performance maps

8. Load matching: torque and rotational speed requirements

9. Pollutant formation and control


General remarks and engine classification


Engine components

Image sources: (left) M. Ehsani, Y. Gao, S.E. Gay, A. Emadi, Modern Electric, Hybrid

Electric, and Fuel Cell Vehicles, CRC Press LLC, New York, 2005.(right) J.B. Heywood, Internal Combustion Engine Fundamentals, McGraw-

Hill, New York, 1988.


1. Method of ignitiona) Spark Ignition, SI (Otto engines)b) Compression Ignition, CI (Diesel engines)

2. Working cyclea) full cycle in 4 piston strokes (four-stroke engine, 4S)b) full cycle in 2 piston strokes (two-stroke engine, 2S)

3. Fuelgasoline, fuel oil (diesel), natural gas, liquid petroleum gas (LPG), alcohols (methanol, ethanol)...

4. Air intakea) Naturally aspirated engineb) supercharged enginec) turbocharged engine

5. Air/fuel mixture preparationa) Carburetionb) Indirect fuel injectionc) Direct fuel injection

6. ApplicationPropulsion (automobile, truck, light aircraft, marine), portable power systems, power generation


Engine classification


Method of ignition

Spark ignition enginesA mixture of air and fuel (usually gasoline) vapor is ignited by an electrical discharge (spark) across the spark plug.

Compression ignition enginesThe fuel is injected directly into the engine cylinder just before the combustion process is required to start. The liquid fuel jet is atomized into tiny droplets, and evaporates inside the hot compressed air; after a short delay period, the air/fuel mixture spontaneously ignites, because temperature and pressure are above the fuels ignition point (thanks to high compression ratios).

Working cycle

o Four-stroke enginesthe working cycle takes four piston strokes, or two crankshaft revolutions, and more than half is dedicated to exhaust gas expulsion (scavenging) from the cylinder and to fresh mixture intake inside the cylinder: the working fluid is thus replaced efficiently at each cycle.

o Two-stroke enginesthe working cycle takes just two piston strokes, i.e. one crankshaft revolution: power density is therefore higher than in 4S engines, but the scavenging process is less efficient.


Engine classification


Bore (cylinder diameter) Crank radius Connecting rod length Piston stroke = 2

Unit displacement = Displacement =

(: number of cylinders) Minimum chamber volume Volumetric compression ratio = ( + )/ Working cycle frequency = /

(2S: = 1; 4S: = 2) Rotational speed = 2 Crank angle = = 2 Mean piston velocity = 2


Geometric and kinematic parameters

Image source: R. Stone, Introduction to Internal Combustion Engines, Palgrave Macmillan, 2012

TC or TDC: Top Dead CenterBC or BDC: Bottom Dead Center


Geometric and kinematic parameters


Typical values of geometric and kinematic parameters:

Ratio of cylinder bore to piston stroke = / 0.8 - 1.2 Ratio of connecting rod length to crank radius = / 3.0 - 4.0 Volumetric compression ratio (SI engines) 8 - 12 Volumetric compression ratio (CI engines) 12 - 24 Mean piston speed 8 - 15 m/s

0 45 90 135 1800

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

[deg]

u

p

/

u

p

R = 3.0R = 3.5R = 4.0

1. Intake stroke: fresh mixture is drawn into the cylinder by the depression induced by the piston stroke. To increase the mass inducted, the inlet valve opens shortly before the stroke starts and closes after it ends.

2. Compression stroke: air (or air/fuel mixture) is compressed to a small fraction of its initial volume, reaching pressure and temperature levels that depend on initial pressure and volumetric compression ratio . Toward the end of the compression stroke, combustion is initiated and the cylinder pressure rises more rapidly.

3. Power stroke: high temperature and pressure gases push the piston down, forcing the crank to rotate. As the piston approaches BDC the exhaust valve opens to initiate the exhaust process and drop the cylinder pressure to close to the exhaust pressure.

4. Exhaust stroke: exhaust gases exit the cylinder, first spontaneously because the pressure inside the cylinder is higher than inside the exhaust manifold (blowdownprocess), then because they are displaced by the piston as it moves toward TDC. As the piston approaches TDC the inlet valve opens, and just after TDC the exhaust valve closes, so the cycle starts again.


Four-stroke engine working cycle


Image source: R. della Volpe, Macchine, Liguori Editore, Napoli, 2011

1 2 3 4


I. Compression stroke: as the piston moves toward TDC, the mixture inside the cylinder is compressed, while at the same time the pressure inside the crankcase decreases, so fresh air is drawn as soon as the inlet ports (2) are uncovered by the piston. As the piston approaches TDC, combustion is initiated.

II. Power (or expansion) stroke: all ports are closed by the piston in the first part of the stroke, then the exhaust ports (1) are first uncovered, and most of the burnt gases exit the cylinder in an exhaust blowdown process. Then the transfer ports (3) are uncovered and fresh charge, which has been compressed in the crankcase during the compression stroke, flows into the cylinder. The piston and the ports are generally shaped to deflect the incoming charge from flowing directly into the exhaust ports, so as to achieve effective scavenging of the residual gases.


Two-stroke engine working cycle

Image source: (top) R. della Volpe, Macchine, Liguori Editore, Napoli, 2011


Naturally aspirated SI 4S engine cutaway


9. Crankshaft10.Sump11.Oil pump12.Camshaft13.Pushrod14.Coil ignition15.Spark plug16.Exhaust valve17.Rocker arm

1. Air filter2. Carburetor3. Engine head4. Exhaust pipe5. Cylinder block6. Piston7. Alternator8. Connecting rod




3-cylinder SI 4S engine cutaway

Image source: MTZ worldwide, February 2013


Naturally aspirated CI (Diesel) 4S engine cutaway


7. Oil pump8. Oil filter9. Injection pump10.Glow plug11.Injector12.Camshaft

1. Engine head2. Piston3. Cylinder block4. Connecting rod5. Crankshaft6. Sump



CI engine limitations:

o heavier weight, due to higher compression ratios necessary to reach pressures and temperatures required for fuel auto-ignition;

o lower specific power (with reference to engine displacement), and consequently higher footprint for the same power, due to lower rotational speed;

o higher noise level, because of the different nature of the combustion process.

CI engine advantages:

o higher overall efficiency, thanks to higher compression ratios, which in SI engines are limited in order to avoid combustion anomalies (knocking);

o better part-load performance, thanks to the different power control system (no need to throttle air in the intake manifold: partialization is obtained simply through the control of the total amount of fuel injected);

o lower quality fuel required (even though emission regulations have increased manufacture costs).


Comparison between SI and CI engines


2S engine limitations:

o lower efficiency, because work is not delivered during the whole expansion stroke, but only during the first part of it, before exhaust ports are uncovered to allow exhaust gas blowdown and scavenging;

o worse exchange of working fluid -> higher pollutant emissions;

o higher thermal and mechanical stresses, because every stroke (compression and expansion) is marked by high pressures and temperatures, while in 4S engines two strokes out of four are dedicated to the gas exchange process (air induction and exhaust gas expulsion), which takes place at low pressures and temperatures.

2S engine advantages:

o higher specific power: theoretically double than 4S engines, because in 2S engines there is a power stroke at every crankshaft revolution; in practice, specific power is only about 5060% higher due to worse gas exchange process;

o simpler construction, because ports or automatic valves can be used to control air intake and exhaust discharge, rather than cam-controlled valves necessary in 4S engines;

o more uniform torque, in particular for low-power engines, because a useful work phase takes place for every crankshaft revolution.


Comparison between 4S and 2S engines


Ideal thermodynamic cycles

Operating parameters

Thermodynamic processes1-2 Adiabatic compression2-3 Heat input:

Otto: constant volumeDiesel: constant pressureSabath: mixed (isochoric 2-3, isobaric 3-3)

3-4 Adiabatic expansion4-1 Constant volume heat rejection

Constant-volumeOtto

Constant-pressureDiesel

Limited-pressureSabath


Ideal cycles parameters

Volumetric compression ratio: = / Heat input ratio at constant volume: = / Heat input ratio at constant pressure: = /


Heat input: = + ( )

Rejected heat: = ( )

Cycle efficiency:

= 1

+ Reversible adiabatic compression:

=

Isochoric heat input:

= =

Isobaric heat input: = =

Reversible adiabatic expansion:

=

=

=

1

=

=


Ideal cycle analysis

Sabath cycle efficiency:

= 1 1

1 1 + 1

Otto cycle efficiency:

= 1 1

Diesel cycle efficiency:

= 1 1

1 1

Ideal cycle efficiency:

depends on fluid properties though the ratio of specific heats = /

increases with volumetric compression ratio , because it increases the average temperature of heat input

decreases as the ratio = / increases


Ideal cycle efficiency


8 12 16 20 240

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

rv

i

d

k = 1.4b = 2.0

OttoDiesel


In order to take into account the real

thermodynamic behavior of the working fluid,

a fuel/air cycle is defined. Its characteristics are:

air and combustion products are perfect gases,

with specific heat (, ) and specific heat

ratio dependent on temperature;

the combustion process is complete and

instantaneous;

no heat transfer takes place with engine walls;

reversible compression and expansion.

The fuel/air cycle efficiency (or /) is

thus defined as the ratio between the work output

of such a cycle and its heat input:

= /

The plot on the right shows the influence of volumetric compression ratio

on fuel/air cycle efficiency for different equivalence ratios

(defined later as /), in the case of a constant-volume (Otto) cycle.

is the absolute humidity; the fraction of exhaust gas residuals.


Fuel/air cycle efficiency

Image source: R. Stone, Introduction to Internal Combustion Engines, Palgrave Macmillan, 2012


Further differences between air/fuel cycle and real cycle are due to:

Finite combustion time: the combustion process usually lasts at least 50 crank angle degrees

Incomplete combustion and chemical dissociation

Heat transfer between burned gases and cylinder walls and between air and intake manifold

Crevice effects and leakage

Exhaust blowdown loss due to anticipated opening of exhaust valve

Pressure losses at intake and exhaust valves: in naturally aspirated engines, this means in

particular that work must be done by the piston on the gas during the intake and the exhaust

processes (pumping work)

Finite valve opening and closing time

The real working cycle is defined as indicated cycle (see the following slide), so an

indicated thermodynamic efficiency may be defined as the ratio of the actual work output and the

work output of a corresponding fuel/air cycle:

=


Air/fuel cycle vs real thermodynamic cycle



Indicated cycle

The real working cycle and engine performance are

measured with a dynamometer: the engine is clamped on

a test bed and the shaft is connected to the dynamometer

rotor, which is coupled to the stator by electromagnetic,

hydraulic or mechanical (friction) means.

The force required to balance the stator gives the

engine torque:

=

Engine power is then given by the product of torque and

rotational speed

= =

This is the usable power that is delivered by the engine to

the load, which is in this case a brake (brake power):

hence the suffix b.

Besides, pressure inside the engine is measured by

sensors called indicators (hence the definition of

indicated cycle) as a function of crank angle and,

consequently, cylinder volume .


Real (indicated) cycle


Pressure vs. crank angle (-) diagram

0 Intake valve opening (IVO)

1 Intake valve closing (IVC)

* Start of combustion (SOC)

4 Exhaust valve opening (EVO)

6 Exhaust valve closing (EVC)

Indicated cycle

(pressure vs cylinder volume diagram, -)

Left: indicated cycle of a 4S, SI engine

Right: indicated cycle of a 4S, CI engine


Intake

Compression Expansion

Exhaust

EV

C

IVC

EV

O

IVO

TDC BDC TDC BDC TDC


Indicated work per cycle (per cylinder)

=

Indicated mean effective pressure (imep)

=

=

Indicated fuel conversion efficiency

= =

Indicated cycle power output

= = / = /

Mechanical efficiency

=

=

is the fraction of the engine power needed

to drive accessories and overcome friction

Brake mean effective pressure (bmep)

=/

bmep is the work available per cycle and unit

displacement

Friction mean effective pressure (fmep)

=/

fmep is here defined to include work spent to

drive accessories

Mechanical balance in terms of power =

Mechanical balance in terms of effective pressures:

=

=


Indicated cycle


Overall fuel conversion efficiency must take into account energy losses due to friction and work

necessary to drive accessories, so it is defined in terms of brake power:

=

It is therefore the product of indicated fuel conversion efficiency and mechanical efficiency:

= =

The specific fuel consumption is, by definition, the fuel mass flow rate that must be burned in the

engine to obtain a unit power output: thus, it is given by the inverse of the product of fuel conversion

efficiency and heating value. It is usually expressed in [g/kWh]:

=

=


Overall fuel conversion efficiency and specific consumption


Volumetric efficiency is used to measure the effectiveness of an engines induction process, and is

defined as the mass of air that effectively flows into the intake system divided by the mass of air

that would fill a volume equal to the displacement at inlet air conditions (inlet air density ):

=

=

In naturally aspirated engines volumetric efficiency is lower than 1 because of pressure losses in the

intake system (distributed losses in the intake manifold and concentrated losses in the intake valve).

Typical maximum values are in the range 8090% for SI engines, and somewhat higher in CI

engines (because there are no throttling losses, which in some measure are always present in SI

engines even at full load).

In supercharged and turbocharged engines inlet air density is higher than the ambient value, so

volumetric efficiency is higher than 1 (and for this reason it is not totally appropriate to talk about

efficiency in this case).

Inlet air mass flow rate can be expressed as follows:

= /


Volumetric efficiency


An obviously important parameter for the

combustion process is the air/fuel ratio :

=

The fuel/air equivalence ratio is an even

more informative parameter for defining

mixture composition, because it compares the

stoichiometric air/fuel ratio to the actual one:

=

Its inverse is the relative air/fuel ratio :

=

According to these definitions:

o for fuel-lean mixtures: < 1, > 1

o for stoichiometric mixtures: = 1, = 1

o for fuel-rich mixtures: > 1, < 1

The energy input into the system per unit time

can thus be expressed as:

=

/ =

/

Power output:

=

/

Brake mean effective pressure

=

Torque is proportional to bmep and engine

displacement:

=

=


Air/fuel ratio, torque and power output


Fundamental energy balance equation:

= = + + + +

Heat flux absorbed by refrigerating fluid and lubrication oil:

= +

Heat flux rejected as sensible heat in the exhaust gases:

,( )

Other forms of heat dissipation ( ):

o incomplete combustion;

o radiation.


Thermal energy balance

Spark Ignition Engines Compression Ignition Engines

20 30% 28 40%

16 33% 15 37%

30 50% 24 40%

+ 4 20% 4 12%

Source: R. della Volpe, Macchine, Liguori Editore, Napoli, 2011


Taking into account indicated thermodynamic efficiency , if rotational speed increases the

working cycle gets shorter, so:

leaks of working fluid and heat transfer between fluid and engine walls both decrease

energy losses due to imperfect and incomplete combustion increase

On the other hand, the air/fuel cycle efficiency slightly increases with the rotational speed because

of higher dilution of fuel (the amount of residual gas is higher) and thus lower temperatures (which

reduce the effect of the specific heats variability).

In the case of mechanical efficiency , if rotational speed increases:

friction losses and pumping work also increase

Therefore, varying rotational speeds:

indicated thermodynamic efficiency has a maximum, albeit with only a slight variation

mechanical efficiency decreases as rotational speed increases

fuel conversion efficiency has therefore a maximum, but with a significant decrease only at

high rotational speeds, due to the marked drop in mechanical efficiency


Influence of rotational speed on efficiencies

Fuel/air cycle efficiency : for fuel-rich mixtures ( < ,

> 1) a fraction of the fuel cannot burn, so it decreases almost

linearly with the air/fuel ratio, for fuel-lean mixtures, ( > ,

< 1), there is a slight increase due to higher dilution

Indicated thermodynamic efficiency : it is highly dependent

on air/fuel ratio because of its influence on reaction speed. For

air/fuel ratios markedly higher or lower than the stoichiometric

value the speed of the chemical reactions decreases

significantly, bringing about higher energy losses and therefore

efficiency losses.

Maximum reaction speeds are obtained with slightly rich

mixtures ( 0,9); reactions are effectively frozen for < 0,5

or > 1,5.

Friction mean effective pressure does not depend on air/fuel

ratio: mechanical efficiency thus depends on air/fuel ratio only

through its influence on imep, and therefore it varies as the

indicated efficiency does

(because = 1 / and ).


Influence of air/fuel ratio on efficiencies in SI engines


Image source (bottom): G. Ferrari, Motori a combustione interna, Il Capitello, Torino, 1996

Equivalence ratio

Rich mixture Lean mixture

Air/fuel ratio

Specific

pow

er

outp

ut [k

W/d

m3]

Specific

fuel consum

ption [g/M

J]

Overa

ll fu

el convers

ion e

ffic

iency [%

]

/

Since CI engines work exclusively with lean mixtures

( 0,7), increasing air/fuel ratios is always beneficial

with reference to dissociation and incomplete

combustion, so the indicated thermodynamic

efficiency increases.

The different combustion mechanism makes the effects

related to reaction speeds much less important.

Only with extremely lean mixtures the indicated

thermodynamic efficiency drops significantly.

Regarding the influence of air/fuel ratio on mechanical

efficiency, the same considerations apply to CI and SI

engines, but the mechanical efficiency curve for CI

engines is different than for SI engine because imep

behaves differently.


Influence of air/fuel ratio on efficiencies in CI engines




Typical design and operating data for internal combustion engines

Source: J.B. Heywood, Internal Combustion Engine Fundamentals, McGraw-Hill, New York, 1988.


Polar valve timing diagram

Air intake

Intake Valve

Intake valve opens (IVO) before TDC in order to have the valve completely open when the induction stroke begins, so as to maximize air induction into the cylinder.

Intake valve close (IVC) after BDC in order to take advantage of kinetic energy of exhaust gases in the intake manifold, so as to achieve a good air induction, thanks to inertial effects, even after BDC.

Image sources:

(left) R.N. Brady, Internal Combustion (Gasoline and Diesel) Engines, In: Encyclopedia of Energy, Elsevier, New York, 2004, Pages 515-528, ISBN 9780121764807. (right) A. Paul, P.K. Bose, R. S. Panua, R. Banerjee, An experimental investigation of performance-emission trade off of a CI engine fueled by dieselcompressed natural gas (CNG) combination and dieselethanol blends with CNG enrichment, Energy, Volume 55, 15 June 2013, Pages 787-802, ISSN 0360-5442, http://dx.doi.org/10.1016/j.energy.2013.04.002.


Polar valve timing diagram

Air intake

Exhaust Valve

Exhaust Valve Closes (EVC) after TDC in order to take advantage of the inertia of the exhaust gases, which draw even more fresh air into the cylinder thanks to the overlap period (time while

intake and exhaust valve are simultaneously open).

Exhaust Valve Opens (EVO) before BDC so as to discharge initially the burned gases due to the pressure difference between the cylinder and the exhaust system (blowdown); after BDC the

cylinder is scavenged by the piston as it moves toward TDC (displacement process).

Advanced EVO allows to reduce pumping work (pressure decreases in the cylinder) but it also

reduces the power stroke, so an optimum value exists as a compromise between these effects.

Image sources:

(left) R.N. Brady, Internal Combustion (Gasoline and Diesel) Engines, In: Encyclopedia of Energy, Elsevier, New York, 2004, Pages 515-528, ISBN 9780121764807. (right) A. Paul, P.K. Bose, R. S. Panua, R. Banerjee, An experimental investigation of performance-emission trade off of a CI engine fueled by dieselcompressed natural gas (CNG) combination and dieselethanol blends with CNG enrichment, Energy, Volume 55, 15 June 2013, Pages 787-802, ISSN 0360-5442, http://dx.doi.org/10.1016/j.energy.2013.04.002.

Volumetric efficiency increases as the cross-section

available to induced air and to exhaust gases

increases.

The influence of intake cross-section is much more

pronounced than that of the exhaust cross-section, by

a factor equal to (because it affects the fluid filling

the whole cylinder, and not only the dead volume).

In order to increase intake and exhaust cross-sections,

it is obviously better to adopt multivalve systems

rather than increasing the size of a single valve.

Therefore the following systems are commonly used:

o 3 valves (2 intake and 1 exhaust)



(anyway, exhaust valves are usually smaller)


Intake and exhaust valves

Air intake


Air intake


Phenomena influencing gas exchange processes

In naturally aspirated engines, the volumetric efficiency is

lower than 1 because of several phenomena taking place

in the gas exchange process. The most important are:

1. Exhaust gas expansion: at the end of the exhaust

stroke, the dead volume is filled with exhaust gases,

whose pressure is higher than the atmospheric pressure

( > ): these gases expand as the induction process

begins and the pressure in the cylinder decreases, and

occupy a volume larger than the dead volume.

2. Pressure losses through the intake valve: when the

induction stroke ends, pressure inside the cylinder is

lower than atmospheric pressure (1 < ) because of

energy losses taking place as the fresh air (or mixture)

flows through the valve, so the air density is also lower

than atmospheric air density (1 < ).

3. Heat exchanges between engine walls and induced air

also make airs density lower than atmospheric density.

Image source: G. Ferrari, Motori a combustione interna, Il Capitello, Torino, 1996

Volumes

Pre

ssure

[kP

a]

Intake Exhaust


Volumetric efficiency: quasi-static effects

Air intake

Fuel vapor (and also water vapor) reduces the air partial pressure

below the mixture pressure. If is the mixture pressure (at the

beginning of the compression stroke), then = , + , + ,

(a -> air, f -> fuel, w -> water). Partial pressure of air is given by:

,

= 1 +

+

1

This effect can be quite large for gaseous fuels.

Fuel vaporization: the mixture temperature decreases as liquid fuel is

vaporized. The temperature difference that occurs after evaporation is:

=,

, + ,/

with: fraction of fuel evaporated; , enthalpy of vaporization.

For isooctane, at = 1, 19 K; for alcohols, given their large

enthalpy of vaporization, the effect can be quite large ( 128 K

for methanol) and compensate for the reduction in air partial pressure.

As the pressure in the exhaust manifold increases, the volume

occupied by the residual gas in the cylinder also increases, so the

volumetric efficiency decreases.

Image source: J.B. Heywood, Internal Combustion Engine Fundamentals, McGraw-Hill, New York, 1988.

Volumetric efficiency increases with ambient

temperature (according to a power 0,5),

because less heat is exchanged between

mixture and walls. In any case, the mass of

induced air decreases as ambient

temperature increases:

=

1

=

1

1

0,5

Refrigerant temperature affects wall

temperature and thus heat transfer during

the gas exchange process: therefore,

volumetric efficiency slightly increases as

refrigerant temperature decreases.


Volumetric efficiency: influence of ambient and refrigerant temperature

Air intake


Volu

metr

ic e

ffic

iency r

atio /

Volu

metr

ic e

ffic

iency r

atio /

Refrigerant temperature [K]

Ambient temperature [K]

= 302,4 K

= 363 K

0,5

Volumetric losses due to charge heating

decrease as (and therefore ) increases

because the time available for heat exchange

between fluid and engine walls also decreases.

The effect of friction, both at the intake and at

the exhaust, is proportional to 2 at low-medium

speeds, and increases even more strongly with

at high speeds, when sonic flow conditions

(choking) are reached at the intake valve.

Backflow: because the inlet valve closes after

the start of the compression stroke, a reverse

flow of fresh charge from the cylinder back into

the intake can occur as the piston moves

toward TDC. This flow is larger at lower speeds.

Dynamic effects: inertial effects and wave

effects (tuning).


Volumetric efficiency: influence of rotational speed

Air intake

Image source: J.B. Heywood, Internal Combustion Engine Fundamentals, McGraw-Hill, New York, 1988.

Early Exhaust Valve Opening (EEVO)

If the exhaust valve closes much earlier than BDC,

energy losses for incomplete expansion increase,

but the blowdown process is more effective, thus

reducing the work in the exhaust stroke.

Therefore, a compromise value must be sought.

Late Intake Valve Closing (LIVC)

As the difference between BDC and intake valve

closing increases, the volumetric efficiency vs

engine speed curve shifts towards higher rotational

speeds. With fixed valve timing, this also means

that backflow increases at low speeds.


Volumetric efficiency: effect of valve timing

Air intake


Cylinder volume

Pre

ssure

[M

Pa]

Volu

metr

ic e

ffic

iency

Engine speed [rps]

[rpm]

IVC point

(degrees after BDC)

EVO point

(degrees before BDC)

TDC BDC

The pressure in the inlet manifold varies during each cylinders

intake process due to the piston velocity variation, and the unsteady

gas-flow effects that result from these geometric variation.

At higher engine speeds, the inertia of the gas as the intake valve is

closing increases the pressure in the inlet port and continues the

charging process as the piston slows down around BDC and starts

the compression stroke. The inlet valve is closed some 40 to 60

after BDC, in part to take advantage of this ram phenomenon.

The fluid filling intake manifold and cylinder can be approximated as

a 1-degree of freedom oscillator (Helmholtz resonator), whose

natural frequency is:

0

2

Pressure oscillations increase volumetric efficiency when the intake

systems natural frequency is twice the rotational speed: 0 2.

Therefore, given the geometry of the intake system, the rotational

speed that maximizes volumetric efficiency is:


Volumetric efficiency: inertial effect

Air intake


Volu

metr

ic e

ffic

iency r

atio /

Frequency ratio 0/

0

2

Pressure oscillations due to wave propagation in the

intake system influence the gas exchange process.

When the intake valve opens, a rarefaction wave

propagates upstream; it is then reflected back as a

pressure wave at the first discontinuity.

If the reflected wave reaches the intake valve in the

second half of the induction phase (about 90 after TDC),

pressure at the valve increases just as the intake stroke is

almost complete, aiding the induction process.

In this case the intake system is said to be tuned.

The time required for a wave to travel along the system is

= /, which, in terms of crankshaft angle, is:

= 2 = 2/

The system is tuned if: 2 /2 /

Given actual values of sound speed , an intake system is

usually tuned at high rotational speeds; it could also be

tuned at relatively low speeds with very long pipes.


Volumetric efficiency: wave effects (tuning)

Air intake


Pre

ssure

[M

Pa]

Crankshaft angle [deg]

TDC BDC

Reflected

wave

Resulting pressure

Without reflection

LIVC EIVO

Air intake

a)Mechanical supercharging

b)Turbocharging

c) Engine-driven compressor and turbocharger

d)Two-stage turbocharging

e)Turbocharging with turbocompounding

f) Turbocharger with intercooler

C Compressor

E Engine

I Intercooler

T Turbine


Supercharging and turbocharging: layouts

Image source: J.B. Heywood, Internal Combustion Engine Fundamentals, McGraw-Hill, New York, 1988

Air intake


Supercharging: Roots compressor (positive displacement)

Image source:

(right) R. della Volpe, Macchine, Liguori Editore, Napoli, 2011

(left) P.W.Wetzel, J.P.Trudeau, New supercharger for downsized Engines, MTZ worldwide, February 2013

Air intake


Turbocharger

Image source: E. Chebli et al., Development of an exhaust-gas turbocharger for HD Daimler CV engines, MTZ Worldwide, February 2013

Air intake


Turbocharging with intercooler



Turbocharging

Air intake

Turbocharging is based on the residual energy of the exhaust gases

Backpressure increases as exhaust gases flow through the turbine

Area 2-3-4-5 represents the energy related to a spontaneous discharge in atmosphere, and it is the

maximum energy that could theoretically be extracted from the exhaust gases in an ideal impulse

layout (where each cylinder is directly connected to the turbine inlet)

In a constant pressure system, each cylinder discharges to an exhaust manifold that is large enough

to dampen pressure oscillations; kinetic energy is then lost while exhaust gases expand to pressure

(2-3 process), with an increase in total enthalpy and temperature


pre

ssure

Volume

TDC BDC

Adiabatic expansion

Air intake

For positive displacement compressors, pressure ratio is almost independent on the engines

rotational speed: therefore, good performances can be achieved even at low rotational speeds and

when accelerating from low speeds.

The turbocharger on the other hand provides increasing pressure ratios as the engine speed

increases: inlet pressure may be either insufficient at low speeds or too high at high speeds.

The turbocharger is more reliable; installation and maintenance are easier.

The turbocharger weighs less and is smaller than a supercharger, other things being equal

(in particular, for the same pressure ratio and air flow rate).

Pressure ratio is approximately up to 3 for a turbocharger and up to 2 for a Roots compressor.

During transient response, turbochargers do not respond as fast

as mechanically driven compressors (turbo lag), due to the time

needed for the exhaust system and turbocharger to generate

the required boost.

Inertia, friction, and compressor load are the primary contributors

to turbo lag.

The mechanically driven compressor has a faster response

because its rotational speed is directly coupled to the engines speed.


Supercharging vs. turbocharging

Time [s] T

orq

ue

Mechanically driven

Turbo

Naturally Aspirated


Air intake

Spark Ignition engines:

o The degree of supercharging in SI engines is mainly limited by the knock: supercharging reduces

ignition delay which increases the knocking tendency.

o Before the advent of direct injection systems, turbocharging required a decrease of the

compression ratio in order to reduce the risk of knocking -> higher specific fuel consumption.

o With the advent of direct injection systems and the availability of better fuels (higher octane

numbers), turbocharging could be more easily adopted, because fuel vaporization, together with a

wide use of charge cooling, allows to avoid the need to reduce compression ratios.

Compression Ignition engines:

o Since in CI engines there is no risk of knocking, supercharging is limited only by the maximum

permissible mechanical and thermal loads. Indeed, supercharging has even a positive influence on

the combustion process.

o Therefore, turbocharging is extensively used in CI engines, and it makes their performance reach

the same level of naturally aspirated SI engines.


Turbocharging: applications


Turbocharging: exhaust gas by-pass valve

Air intake

The turbocharger is sized in such a way that it can

provide maximum supercharging pressure at

4050% of the maximum rotational speed, in order

to achieve good torque output even at low speeds.

At high rotational speeds, the excess exhaust

gases are discharged through a bypass valve.


Bypass valve

Exhaust gas

Turbocharger

Fresh air

Engine

Discharge Turbine Inlet

Air intake

As an alternative, in order to achieve

good performances on a wide range

of rotational speeds, a variable-

geometry turbine can be used, which

is equipped with variable-pitch nozzle

blades.


Turbocharging: variable geometry turbine


Equivalence ratio

Rich mixture Lean mixture

Air/fuel ratio

Specific

pow

er

outp

ut [k

W/d

m3]

Specific

fuel consum

ption [g/M

J]

Overa

ll fu

el convers

ion e

ffic

iency [%

]

/

Fuel metering in SI engines

Fuel volatility

Fuel must be premixed with air, before the

spark starts the combustion process.

Air/fuel ratio

The metering system must provide the

appropriate quantity of fuel, so as to obtain the

required air/fuel ratio for every operating point.

Mixture homogeneity

The air/fuel mixture must be homogeneous, in

order to burn fuel rapidly and completely.


Requirements of a 4S, SI engine



1. Air filter

2. Carburetor

3. Throttle valve

4. Intake manifold

5. Fuel tank

6. Fuel filter

7. Cam

8. Diaphragm pump


Carburetor



Fundamental components:

Venturi (convergent-divergent nozzle)

In its throat the depression originated by the

inlet air flow draws fuel from its tank.

Fuel discharge tube

Connects the fuel tank to the Venturi throat.

The fuel flow is metered by a calibrated orifice.

Float chamber

The fuel level is maintained at a constant

height in a float chamber: a pressure

equalizing passage makes the pressure inside

the float chamber equal to air pressure at the

Venturi inlet. Thus, hydrostatic pressure on the

calibrated orifice depends only on the flow rate

of air.

Throttle valve

It controls air flow rate, and as a consequence

power output, acting on volumetric efficiency.


Elementary carburetor


Air

Fuel

Float

chamber

Calibrated

orifice

Throttle

plate

Pressure

equalizing

passage

Venturi

throat

Fuel discharge

nozzle


Limits of carburetors

It is difficult to control precisely the required

air/fuel ratio in different operating conditions

Pressure loss in aspiration

It is difficult to control its operation in transient

conditions, due to the inertia of the fuel mass

Significant variation of specific fuel

consumption with load, particularly pronounced

in case of frequent transient operating conditions

It is necessary to heat the intake manifold in

particular operating conditions, with the

consequent decrease in volumetric efficiency, in

order to avoid that fuel condenses on engine walls

Limited control on pollutant emissions in the

whole operating range

Advantages of fuel injection

better control on air/fuel ratio, both in terms of

precision and repeatability, in all operating

conditions; more uniform distribution among

cylinders; consequently, positive influence on:

o specific fuel consumption

o performance

o pollutant emissions

better transient operation, because of lower fluid

volumes in the system (lower fluid inertia)

higher volumetric efficiency, thanks to lower

pressure losses and no intake manifold heating

it makes possible to increase compression ratio,

because the fuel/air mixture has less time

available for autoignition, or alternatively to use a

fuel with lower octane rating


Carburetor vs. injection systems


Position of fuel injectors

o Indirect injection (IDI): it takes place in the intake manifold

o Direct injection (DI): fuel is directly injected inside the cylinder

Control of injectors

o Mechanical injection: an engine-driven pump pressurizes the fuel and meters the injected volume

by means of an automatic mechanical injector

o Electronic injection: an electromagnetic fuel injector is used; metering and control of injection are

bestowed upon the ECU

Distribution among cylinders (only for indirect injection)

o Single-point systems: only one or two fuel injectors meter the fuel into the air flow directly above

the throttle body

o Multipoint port injection: a fuel injector is used for each cylinder; fuel is injected into the intake

port of each cylinder


Classification of injection systems


Injection systems


Bosch Electromagnetic Injector

(electronic injection system)

1. Filter

2. Electric wire

3. Winding

4. Winding armature

5. Needle valve

6. Needle tip

7. Fuel pipe

8. Tightening ring

9. Seal ring

10.Seal ring

Direct Injection (DI, left)

and Indirect Injection (IDI, right)

1. Injector

2. Intake manifold

3. Intake valve



MultiPoint Fuel Injection (MPFI)


11.Throttle switch

12.Air flow sensor

13.Air temperature sensor

14.Lambda sensor

15.Coolant temperature sensor

16.Auxiliary air device

17.Crankshaft angle sensor

18.Battery

19.Ignition and starting switch

20.HVAC switch

1. Fuel tank

2. Fuel pump

3. Fuel filter

4. Pressure regulator

5. Electronic Control Unit (ECU)

6. Ignition coil

7. HT distributor

8. Spark plug

9. Injector

10.Throttle valve



Direct injection system (Bosch)


Solenoid (HDEV5) and piezo (HDEV 4.1) injectors

Metering accuracy 2 mg/inj

up to 3 injections per cycle (up to 5 inj/cycle for

stratified-charge combustion)

Source: A. Heinstein et al., High-pressure Direct Injection Systems for Gasoline Engines , MTZ Worldwide, March 2013

HDEV5 HDEV4.1


Problems arising with direct injection systems:

Higher mechanical and thermal stress

The injectors must withstand pressure and temperature levels that are reached inside the cylinder;

furthermore, they are subject to dirt and soot particles originating from the combustion process.

Installation complexity

Direct injection requires that the injector be mounted directly on top of the cylinder, where spark plug

and intake and exhaust valves are placed too.

Higher injection pressure

Direct injection systems need at least 3540 bar in order to achieve the correct fuel vaporization and

mixing with the air inside the cylinder (modern injection systems reach pressures up to 200 bar).

In the case of indirect injection systems, an injection pressure of 45 bar is good enough.

More difficult mixture homogenization

In order to achieve a mixture homogenization comparable to indirect injection systems, the air intake

system needs to be more sophisticated, so as to increase turbulence (through swish, squish and

tumble motions): a high degree of turbulence speeds evaporation of fuel, enhances air-fuel mixing

and increases combustion speed and efficiency.


Direct vs. Indirect injection systems


Advantages of direct injection systems:

Evaporative cooling

In indirect injections systems, fuel vaporization takes place in the intake manifold and it subtracts

heat from the manifold walls and from the intake valve; in direct injection systems, on the contrary,

fuel vaporization takes place inside the cylinder, so that it cools the induced air, with a double

benefit:

o higher density -> higher volumetric efficiency: torque output increases by approximately 56%;

o lower temperature -> lower risk of detonation (knocking) -> compression ratio can be increased

(by approx. 20%) -> significant improvements in efficiency and fuel consumption are possible.

Cold start and transients

Direct injection removes the problem of fuel condensation on the intake manifolds walls or on the

intake valve, which is particularly important during transient behaviors and at cold start: both fuel

consumption and CO and HC pollutant emissions are reduced.

Longer valve overlap period possible

Fuel injection is carried out when both valves are closed, so the induced air does not contain any

fuel, and thus in the overlap period there is no risk of fuel flowing back to the intake manifold (with

the possible risk of backfire) or of fuel loss with exhaust gases.




Indirect

injection

Direct

injection

[cm3] 3498 3498

10,7 12,2

[bar] 45 200

[kW] 200 215

[rpm] 6000 6400

[bar] 11,4 11,5

/ [kW/l] 57,2 61,5

[Nm] 350 365

[bar] 12,6 13,1

/ [Nm/l] 100,0 104,3

, [g/kWh] 240 235

(part load) [g/kWh] 360 290




T

n [rpm]

Fuel injection in CI engines

CI engines use low-volatility fuels: the fuel must

therefore be injected into the induced air in the form

of droplet spray, and the drops should be as

small as possible.

The air/fuel ratio must be close to the stoichiometric

value only locally, i.e. close to the injector; overall,

the fuel/air mixture can be lean, which allows to

control power output just through the mass of fuel

injected, without the need of controlling also the

mass of air induced.

On the other hand, the fuel/air equivalence ratio

cannot exceed threshold levels given by the

following reasons:

o pollutant emission

o mechanical stresses

o thermal stresses


General remarks


Rotational speed n

Bra

ke m

ean e

ffective p

ressure

(bm

ep)

[MP

a]

rps

rpm

Power


Indirect injection (prechamber) and direct injection


Indirect injection (prechamber)

1. Injector

2. Prechamber

3. Glow plug (used as a cold-starting aid)

Direct injection

Reduction of heat losses (no heat exchange

with the prechamber walls) -> increase in fuel

efficiency

On the other hand, injection pressure must be

significantly higher




Common Rail Fuel Injection System



Common Rail: injector

Open injector Closed injector

High pressure

fuel inlet

Solenoid

actuator

Two-way valve

High pressure

fuel inlet

Solenoid

actuator

Two-way valve

Electrical

connection

Fuel leak back

(return)

Nozzle open Nozzle closed

Injector valve

open

Injector valve

closed

Solenoid energized Solenoid not energized

Fuel pressure is relieved above

the valve control plunger

Balance of forces: Fa > Fc+Fe

Fuel pressure is the same above

and below the valve control plunger

Balance of forces: Fa < Fc+Fe



Common Rail: injector characteristics

Source: D. Schppe et al., Servo-Driven Piezo Common Rail Diesel Injection System, MTZ Worldwide, March 2012



Common Rail: multiple injections


Common Rail: application examples


Cursor 11 Euro VI engine series, used for commercial truck propulsion (FPT Industrial)



Common Rail: application examples

Vector V20 engine series, used in power generators (FPT Industrial)

Performance maps are usually drawn with reference

to full-load operation (full-load performance maps)

and represent power, torque and fuel consumption vs.

rotational speed.

Torque depends on and , so its maximum lies

where volumetric efficiency is highest, and falls

rapidly because of the decrease in mechanical

efficiency.

Beyond a particular value of rotational speed, the

speed increase cannot compensate the decrease in

mechanical and volumetric efficiency: therefore that

operating point () corresponds to the maximum

power output of the engine.

The operating range of the engine is usually limited to

a maximum speed only slightly higher than

, because there is no point in using the

decreasing part of the power curve.

Fuel specific consumption depends only on the

overall fuel conversion efficiency .


Power, torque and fuel specific consumption curves

Operating characteristics and performance maps


Power

Fuel specific consumption

Torque

This chart shows the performance map of a

naturally aspirated 4S SI engine

(displacement 2525 cm3).

The brake mean effective pressure curve has

the same shape of the torque curve, because:

Curves do not start from zero power output

but from a minimum speed, below which

excessive vibrations and irregular operating

conditions would arise.

The operating range goes from the minimum

speed up to a maximum speed slightly higher

than the maximum power operating point.


Power, torque and brake mean effective pressure curves of a SI engine



Engine characteristics:

Turbocharged

Direct injection

Displacement = 4134 cm3

The torque rises very fast at low speeds,

then it is maintained constant over a wide

range of speed, and finally decreases at high

speeds. This behavior depends on:

the turbocharger characteristics;

the limits on the fuel/air equivalence ratio.


Performance map of a CI engine



T

n [rpm]

The chart on top shows the performance map of two

similar SI 4S naturally aspirated engines of different

displacement.

Displacement does not affect bmep

(there is only a marginal influence on the volumetric

efficiency related to valve area).

Therefore, torque and power output increase linearly

with the increase in displacement, on the whole

operating range.

The bottom chart shows the performance map of two

CI 4S engines, same displacement ( = 2500 cm3),

different number of valves.

The volumetric efficiency increases significantly in the

4-valve engine: this results in a higher torque output.

The increase in power output is even more

pronounced because the 4-valve engine has a wider

speed range (the valves are smaller and lighter, so can

withstand higher velocities).


Influence of displacement and number of intake valves





SI automotive engine performances

Design characteristics and performance of some SI 4S supercharged engines for automotive applications, built by European companies from 2001 to 2007


1 2 3 4 5 6 7

Year 2001 2006 2002 2002 2005 2006 2007

[cm3] 1998 1598 1796 1796 1390 1798 2979

Supercharging Turbo Turbo Mech. Mech. Mech. + Turbo Turbo 2 Turbo in //

Injection Indirect Indirect Indirect Direct Direct Direct Direct

8,8 8,8 8,7 10,5 10,0 9,4 10,2

[bar] 120 120 150 200

[kW] 140 132 141 125 125 118 225

/ [kW/l] 70 82,6 78,5 69,6 89,9 65,6 75,5

[rpm] 5400 5500 5800 5300 6000 5000 5800

[Nm] 250 230 260 250 240 250 400

/ [Nm/l] 125,1 144 144,8 139,1 172,7 139 134,3

(max) [bar] 15,7 18,1 18,2 17,5 21,7 17,4 16,8

(min) [g/kWh] 239 < 250 < 235



CI automotive engine performances

Design characteristics and performance of some CI 4S turbocharged engines for automotive applications, built by European companies from 1999 to 2006


1 2 3 4 5 6 7 8 9 10 11 12 13

Year 1999 2001 2002 2003 2004 2004 2004 2004 2005 2005 2005 2005 2006

[cm3] 3900 1995 2148 3936 2460 2967 2497 1991 1493 2993 2987 4134 3996

Turbocharging

2

Turbo

in //

Turbo Turbo

2

Turbo

in //

Turbo Turbo Turbo Turbo Turbo

2

Turbo,

series

Turbo

2

Turbo

in //

Turbo

Injection CR

I gen.

CR

II gen.

CR

II gen.

CR

II gen.

CR

II gen.

CR

II gen.

CR

II gen.

CR

II gen. CR

CR

III gen.

CR

III gen.

CR

III gen.

18 17 18 17,3 18 17,1 17 18 18 16,5 18 16,5 17

[bar] 1350 1600 1600 1600 1600 1600 1600 1600 1600 1600 1600

n. of injections 1 pilot up to

5 2 pilot 2 pilot > 1 2 pilot 2 pilot 2+1+2 5

[kW] 175 110 110 202 128 171 130 103 70 200 165 240 231

/ [kW/l] 44,9 55,1 51,2 51,3 52,0 57,6 52,1 51,7 46,9 66,8 55,2 58,1 57,8

() [rpm] 4000 4000 4200 3750 3500 4000 4200 4000 4400 3800 3750 3600

[Nm] 560 330 340 650 400 450 400 300 210 560 510 650 730

/ [Nm/l] 143,6 165,4 158,3 165,1 162,6 151,7 160,2 150,7 140,7 187,1 170,7 157,2 182,7

() [rpm] 1750-

2500 2000 2000

1800-

2500 2000

1400-

3250 2000

1600-

3000

1800-

2800 2000

1600-

2800

1600-

3500 2200

[bar] 18,0 20,8 19,9 20,8 20,4 19,1 20,1 18,9 17,7 23,5 21,5 19,8 23,0

[g/kWh] 207 202 205 198 202 202 206 208 202



Performance of 4S, 4-cylinder automotive engines

* Direct control system of valve opening, with no throttle valve.

Design characteristics and performance of the Alfa Romeo Giulietta engines.

Source: http://www.alfaromeo.it/it/Documents/schede-tecniche/GiuliettaSchedaTecnica-ConsumiPrestazioniEmissioni.pdf (last retrieved December 2013).

1 2 3 4 5 6 7 8 9 10

Fuel Gasoline Gasoline Gasoline Gasoline Gasoline LPG Diesel Diesel Diesel Diesel

[cm3] 1368 1368 1368 1368 1742 1368 1598 1956 1956 1956

Turbocharging

Turbo,

inter-

cooler

Turbo,

inter-

cooler

Turbo,

inter-

cooler

Turbo,

inter-

cooler

Turbo,

inter-

cooler

Turbo,

inter-

cooler

Turbo,

inter-

cooler

Turbo,

inter-

cooler,

var.

geom.

Turbo,

inter-

cooler,

var.

geom.

Turbo,

inter-

cooler,

var.

geom.

Air inlet Multiair* Multiair*

Injection MPFI MPFI MPFI MPFI DI MPFI Multijet 2 Multijet 2 Multijet 2 Multijet 2

9,8 9,8 9,8 9,8 9,25 9,8 16,5 16,5 16,5 16,5

[bar] 1600 1600 1600 1600

[kW] 77 88 125 125 173 88 77 103 125 125

/ [kW/l] 56,3 64,3 91,4 91,4 99,3 64,3 48,2 52,7 63,9 63,9

() [rpm] 5000 5000 5500 5500 5500 5000 4000 3750 4000 4000

[Nm] 206 206 250 250 340 206 320 350 350 350

/ [Nm/l] 150,6 150,6 182,7 182,7 195,2 150,6 200,3 178,9 178,9 178,9

() [rpm] 1750 1750 2500 2500 1900 1750 1750 1500 1750 1750

[bar] 18,9 18,9 23,0 23,0 24,5 18,9 25,2 22,5 22,5 22,5


Naturally aspirated SI engine, direct injection,

displacement = Full-load performance map: see Direct vs. Indirect injection systems

Turbocharged CI engine, direct injection,

displacement = Full-load performance map: see Performance map of a SI engine


Fuel economy characteristics


Rotational speed [rpm] Rotational speed [rpm]

bmep bmep

Load matching

1. Load requiring constant rotational speed

(as in the case of an electric generator, with the rotational speed linked to the grid frequency):

as a consequence, torque output depends on the external load.

In this case, in order to contain as much as possible any speed fluctuation, power output is divided

among many cylinders and a flywheel with a large moment of inertia is used.

Furthermore, it is possible to optimize the performance, in terms of fuel specific consumption, for

the required operating conditions, taking into account possible load variations.

2. Load requiring torque output increasing as the square of the rotational speed ):

the engine is thus matched to a fluid machine such as compressors, pumps, aeronautical or

marine propellers, etc.

In this case the engine speed at design operating conditions should be as close as possible to the

optimal one for the external load, in order to reduce the size of the gearbox, or if possible avoid

altogether its use. For large marine 2S CI engines, engine speed has even been reduced down to

12 s-1, so as to couple directly the engine to the propeller.

3. Load requiring a wide range of operating conditions, both in terms of speed and torque.

It is for example the case of ground propulsion.


Typical applications


Power generation: example of state-of-art engine performance

Load matching

Source: MAN Diesel & Turbo Power Plant Programme

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'NGEVT)GP5GVJGCVTCVGCVNQCF'NGEVTKECNQRVKOKUGF9$6#.WHV M,M9J

.WDGQKNEQPUWORVKQP MIJ 0QOKPCNIGPGTCVQTGHEKGPE[/GVJCPGPQ

/#085$QTGOO5VTQMGOO 8 8'PIKPGURGGF TRO (TGSWGPE[ *\ 'NGEVT)GP5GVRQYGT M9

'NGEVT)GP5GVJGCVTCVGCVNQCF.KSWKFHWGN9$ M,M9J

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/#0865$QTGOO5VTQMGOO 1RGTCVKQPOQFG 'PIKPGURGGF TRO (TGSWGPE[ *\ 'NGEVT)GP5GVRQYGT M9

'NGEVT)GP5GVJGCVTCVGCVNQCF9QTNF$CPM M,M9J

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/#0%4$QTGOO5VTQMGOO . . 8 8 8'PIKPGURGGF TRO (TGSWGPE[ *\ 'NGEVTIGPUGVRQYGT M9

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Power generation: example of state-of-art engine performance and dimensions

Load matching


H

WA B

C

18V48/60TS engine

Performance data

Power per cylinder

Tot. engine power

Tot. el. genset power

Spec. fuel oil consumption

acc. to ISO 3046, without pumps,

mech. Power output, +5% tolerance

Heat Rate

acc. to ISO 3046, without pumps,

mech. Power output, +5% tolerance

NOx emissions (dry at 15% O2)

Mean effective pressure

Spec. lube oil consumption

Dimensions (mm)

1

1050

18,900

18,428

171

7,305

1850

23.2/22.6

0,60

B

5410

2

1100

19,800

19,305

172

7,350

1740

24.3/23.7

0,60

C

24510

3

1150

20,700

20,183

174

7,430

1580

25.4/24.7

0,60

H

9023

4

1200

21,600

21,060

177

7,560

1480

26.5/25.8

0,60

W

4694

Unit

kW

kW

kW

g/kWh

kJ/kWh

mg/Nm3

bar

g/kWh

A

9625

Operation mode

Dry mass (t) 407 407 407 407 407


Power generation: example of state-of-art engine performance and dimensions

Load matching


With generator (genset) Without generator

In-line engine L32/44CR

Engine type No. of cyl. L L1 W H Weightmm mm mm mm t

6L32/44CR 6 6,312 5,265 2,174 4,163 39.5

7L32/44CR 7 6,924 5,877 2,359 4,369 44.5

8L32/44CR 8 7,454 6,407 2,359 4,369 49.5

9L32/44CR 9 7,984 6,937 2,359 4,369 53.5

10L32/44CR 10 8,603 7,556 2,359 4,369 58.0

V-engine V32/44CR

Engine type No. of cyl. L L1 W H Weightmm mm mm mm t

12V32/44CR 12 7,195 5,795 3,100 4,039 70

14V32/44CR 14 7,970 6,425 3,100 4,262 79

16V32/44CR 16 8,600 7,055 3,100 4,262 87

18V32/44CR 18 9,230 7,685 3,100 4,262 96

20V32/44CR 20 9,860 8,315 3,100 4,262 104

All weights and dimensions are for guidance only and apply to dry engines without flywheel. Masses include built-on lube oil automatic filter, fuel oil filter and electronic equipment.Minimum centreline distance for twin engine installation: 2,500 mm (L32/44CR), 4,000 mm (V32/44CR). More information available upon request.

GenSet dimensions

A mm 7,470 8,530 7,055 8,315 9,575

B mm 4,328 4,328 4,376 4,376 4,376

C mm 11,795 12,858 11,431 12,691 13,951

W mm 2,676 2,676 4,200 4,260 4,260

H mm 4,975 4,975 5,000 5,200 5,200

Dry mass t 84 97 117 144 172

Load matching


Power generation: plant layout

Source: Wrtsil, Power Plants Solutions 2013

Load matching


Power generation

Source: extract from The largest recip-based power plants worldwide, Modern Power Systems, February 2014, pp. 18-21.

Facilities of 80 MW or more, operating or under construction

Name Location Capacity MW Fuel Year1 Configuration E/G supplier

Aratu, Salvador Brazil 1056 Diesel U/C 120 x 18V32/40 in 6 units MAN Diesel & Turbo

IPP3 Jordon 573 Tri-fuel 2014 38 x 50DF Wrtsil

Quisqueya I+II Dominican Republic 430 HFO, nat gas 2013 12 x 18V50DF + 12 x 18V50DF Wrtsil

Boyuk Shor Azerbaijan 384 Nat gas 2013 21 x SG Wrtsil

Suape II Brazil 382.5 HFO 2011 17 x 46F Wrtsil

Geramar I+II Brazil 331.8 HFO 2010 38 x 20V32 Wrtsil

Sangachal Baku, Azerbaijan 306.8 FO, nat gas 2012 18 x 16.6 MW 50DF Wrtsil

Coloane, Macau China 271.4 CC HFO, diesel 1978-97 2x24, 2x38.6, 2x53.1, +2x20 ST MAN D&T, Pter Brotherhood

Aliaga Alosbi-II2 Izmir, Turkey 270.6 CC HFO, nat gas 20074 x 18V46, 28 x 20V34SG, 2 x 13.5 MW steam

Wrtsil

Pavana III Honduras 267.2 Oil 2004 16 x 18V46 Wrtsil

Kiisa ERPP 1 & II Estonia 250 Nat gas, LFO 2013-4 27 x W20V34DF Wrtsil

Choloma Honduras 250 HFO 2004-5 14 x 18V48/60 MAN Diesel & Turbo

IPP4 Jordan 240 HFO, DFO, gas 2014 16 x 50DF Wrtsil

Bauang La Union Philippines 241.5 HFO 1994 21 x 16ZA40S Sulzer, Alstom

Plains End, Colorado3 USA 231 Nat gas 2002, 2006 20xW18V34SG, 14xW20V34SG Wrtsil

STEC Red Gate, Texas USA 225 Nat gas 2014 12 x 50SG Wrtsil

Port-Est Reunion 222 HFO 2010 12 x 18V48/60 MAN Diesel & Turbo

Port Westward Unit 2, Oregon USA 220 Nat gas 2015 12 x 50SG Wartsila

Atlas Pakistan 220 Furnace oil 2009 11 x 18V48/60 MAN Diesel & Turbo

Kribi Cameroon 216 Nat gas/LFO 2013 Wrtsil

Pearsall, Texas USA 202.5 Nat gas 2010 24 x 20V34SG Wrtsil

Linhares Brazil 204 Nat gas 2010 24 x 20V34SG Wrtsil

Pesangarran, Bali Indonesia 200 Nat gas, HFO 2014-5 12 x 50DF Wrtsil

Nishat Pakistan 200 HFO 2010 11 x 18V46 Wrtsil

Nishat Chunian Pakistan 200 HFO 2010 11 x 18V46 Wrtsil

Vasavi India 200 HFO 1998 4 x 12K90MC-S MAN D&T, Hyundai, ABB

Liberty Power Tech Pakistan 200 CC HFO 2010 11 x 18V46, CC plant (1xST) Wrtsil

Sasolburg South Africa 175 Nat gas 2012 18 x W20V34SG Wrtsil

Viana Brazil 175 HFO 2009 20 x 20V32 Wrtsil

La Paz (Baja California Sur) Mexico 173 Diesel oil 2005-13 4 units, CC plant Man Diesel & Turbo

Eklutna, Arkansas USA 171 Nat gas, LFO 2014 10 x W18V50DF Wrtsil

Cear Brazil 168 HFO 2010 8 x 20V46F Wrtsil

Clifton Pier Bahamas 165 HFO 19631 x 6 MW, 4 x 10 MW, 2 x 26.5 MW, 2 x 33 MW

Sulzer, MAN Diesel & Turbo

Campina Grande Brazil 164 HFO 2010 20 x 20V32 Wrtsil

King Salmon California 163 Nat gas, diesel 2010 10 x 18V50DF Wrtsil

Cntrl Termica Ressano Garcia Mozambique 162 Nat gas 2014 18 x 34SG Wrtsil

Planta Arizona Guatemala 160 HFO, LFO 2003 10 x 18V46 Wrtsil

Sapugaskanda Sri Lanka 160 HFO 1984 16 x 10 MW SEMT, MAN D&T, Siemens

Attock Pakistan 160 HFO 2008 9 x 18V46 Wrtsil

Load matching

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Load matching


Combined Heat and Power generation (CHP)


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Total CHP efciency 90%

Heat to heatconsumers 44.5%

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CHP systems: example of state-of-art engine performance

Load matching

GE Power & Water, J920 FleXtra Jenbacher

Installed Dimensions Length Width Height Weight

Engine 8.4 m 2.9 m 3.3 m 87 t

Generator 5.2 m 2.5 m 2.9 m 54 t

TCA Module 3 m 6.4 m 3.4 m 36 t

Key Performance DataPerformance Data J920

(50Hz / 1,000 rpm)J920 (60Hz / 900 rpm)

Electrical Output 9,500 kW 8,550 kW

Electrical Efficiency 48.7% 48.7%

Heat Rate 7,392 kJ/kWh 7,392 kJ/kWh

Thermal Output 8,100 kWth 7,300 kWth

Total Efficiency 90% 90%

Output and efficiency at generator terminals, ISO 3046, Nat. Gas MN >80, Power Factor 1.0, 500 mg/Nm3 (@ 5% O2) NOx, Efficiency at LHV

Ins

En

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CHP systems: example of state-of-art engine performance

Load matching

GE Power & Water, Jenbacher Type 6 Gas Engines

1) Electrical output based on ISO standard output and standard reference conditions according to ISO 3046/I-1991 and p.f. = 1.0 according to VDE 0530 REM with respective tolerance; minimum methane number 80 for natural gas

All data according to full load and subject to technical development and modification.

outputs and efficiencies

technical data

1) Dimensions and weights are valid for 50 Hz applications.*J624 with 2-stage turbocharging

Dimensions l x w x h (mm)1Configuration V 60Bore (mm) 190Stroke (mm) 220Displacement/cylinder (lit) 6.24Speed (rpm) 1,500 (50 Hz); 1,500 with gearbox (60 Hz)Mean piston speed (m/s) 11 (1,500 1/min)

Scope of supply Generator set, cogeneration system, containerized package

Applicable gas types Natural gas, flare gas, biogas, landfill gas, sewage gas. Special gases (e.g., coal mine gas, coke gas,

wood gas, pyrolysis gas)

Engine type J612 GS J616 GS J620 GS J624 GS*No. of cylinders 12 16 20 24Total displacement (lit) 74.9 99.8 124.8 149.7

Containerized package J612 - J620 15,000 x 6,000 x 7,300Generator set J612 GS 7,600 x 2,200 x 2,800

J616 GS 8,300 x 2,200 x 2,800J620 GS 8,900 x 2,200 x 2,800

J624 GS* 12,100 x 2,450 x 2,900

Cogeneration system J612 GS 7,600 x 2,200 x 2,800J616 GS 8,300 x 2,200 x 2,800J620 GS 8,900 x 2,200 x 2,800

J624 GS* 12,100 x 2,450 x 2,900

J612 GS J616 GS J620 GS J624 GS*Generator set 20,600 26,000 30,700 49,900Cogeneration system 21,100 26,500 31,300 49,500

Weights empty (kg)1

Natural Gas 1,500 rpm | 50 Hz 1,500 rpm | 60 Hz

NOx < Type Pel (kW)1 el (%) Pth (kW) th (%) tot (%) Pel (kW)1 el (%) Pth (kW) th (%) tot (%)

500 mg/m3N

612 2,004 44.8 1,883 42.0 86.8 1,984 44.3 1.902 42.5 86.8616 2,679 44.9 2,510 42.0 86.9 2,652 44.4 2.535 42.5 86.9620 3,352 44.9 3,110 41.7 86.6 3,319 44.5 3.141 42.1 86.6624* 4,313 46.1 3,931 41.6 87.7

250 mg/m3N

612 2,004 43.5 1,932 42.0 85.5 1,984 43.1 1.952 42.4 85.5616 2,679 43.6 2,575 41.9 85.6 2,652 43.2 2.601 42.4 85.6620 3,352 43.7 3,211 41.8 85.5 3,319 43.2 3.244 42.3 85.5624* 4,313 44.3 4,101 42.1 86.4

Biogas 1,500 rpm | 50 Hz 1,500 rpm | 60 Hz

NOx < Type Pel (kW)1 el (%) Pth (kW) th (%) tot (%) Pel (kW)1 el (%) Pth (kW) th (%) tot (%)

500 mg/m3N

612 1,818 42.8 1,787 42.1 84.9 1,800 42.4 1,805 42.5 84.8616 2,433 42.9 2,385 42.1 85.0 2,408 42.5 2,409 42.5 85.0620 3,044 43.0 2,982 42.1 85.1 3,013 42.6 3,012 42.5 85.1

250mg/m3N

612 1,818 42.3 1,805 42.0 84.3 1,800 41.9 1,823 42.4 84.3616 2,433 42.4 2,405 42.0 84.4 2,408 42.0 2,429 42.4 84.4620 3,044 42.5 3,008 42.0 84.5 3,013 42.1 3,038 42.4 84.5

*J624 with 2-stage turbocharging

In order to make the best use of the power available on the

whole operating range, the ideal performance map requires:

o constant power output

o torque decreasing (with a hyperbolic law) with the

rotational speed

At low speeds, before reaching maximum power output, the

performance map should have the following characteristics:

o constant torque (limited by tire-ground adhesion conditions)

o power linearly increasing with rotational speed

In order to approximate the ideal characteristic as much as

possible, a gear shift (and a clutch) is needed in order to

change the speed ratio between engine shaft and wheels:

the plot on the right shows the use of a discontinuous four-

speed gearbox

Force transmitted to the wheels is proportional to torque


Ground propulsion: ideal performance characteristics

Load matching

Image source: M. Ehsani, Y. Gao, S.E. Gay, A. Emadi, Modern Electric, Hybrid Electric, and Fuel Cell Vehicles, CRC Press LLC, New York, 2005

Vehicle resistance is caused fundamentally by

4 phenomena:

1. tire rolling resistance ( cos )

2. grading resistance ( sin )

3. aerodynamic drag ( 2)

4. acceleration ( d/d)

Taking into account constant cruising speed

conditions ( = 0), rolling and grading

resistance do not depend on vehicle speed

(at least if the influence of on is neglected),

while aerodynamic drag increases with the

square of the vehicle speed:

= + + = +

Engine power output thus depends on vehicle

speed according to the following equation:

= +


Ground propulsion: vehicle resistance

Load matching


az

r

rd

P

P

FMoving direction

r

P

F

Px

z

Moving direction

(a) (b)

FIGURE 2.2Tire deflection and rolling resistance on a (a) hard and (b) soft road surface


Ground propulsion: vehicle resistance

Load matching

Rolling Resistance Coefficients

Conditions Rolling resistance coefficient

Car tires on concrete or asphalt 0.013Car tires on rolled gravel 0.02Tar macadam 0.025Unpaved road 0.05Field 0.10.35Truck tires on concrete or asphalt 0.0060.01Wheels on rail 0.0010.002

Moving direction

High pressure Low pressure

Vehicle Type

Open convertible

Van body

0.50.7

0.50.7

0.40.55

0.30.4

0.20.25

0.23

0.150.20

0.81.50.60.70.30.40.60.7

Coefficient of Aerodymanic Resistance

Ponton body

Wedge-shaped body; headlampsand bumpers are integrated intothe body, covered underbody,optimized cooling air flow

Headlamp and all wheels inbody, covered underbody

K-shaped (small breakwaysection)

Optimum streamlined design

Trucks, road trainsBusesStreamlined busesMotorcycles

Aerodynamic resistance (drag)

= +

Rolling resistance =

Source: M. Ehsani, Y. Gao, S.E. Gay, A. Emadi, Modern Electric, Hybrid Electric, and Fuel Cell Vehicles, CRC Press LLC, New York, 2005

Top chart shows that with a given power

output (corresponding to a given vehicle

speed), the fuel consumption is usually

lower at low engine speed than at high

speed.

The bottom chart shows the operating

points of an engine at constant vehicle

speed, with the highest gear and the second

highest gear.

The engine has a much lower operating

efficiency in low gear than in high gear.

Therefore, the fuel economy of a vehicle

can be improved with more gear

transmission or continuous variable

transmission.


Ground propulsion: specific consumption at constant speed

Load matching


Pollutant formation and control


SI engines: formation mechanisms


Oil layers absorb HC

COMPRESSION COMBUSTION EXPANSION EXHAUST

Deposits

absorb HC

Unburned mixture

forced into

crevices

NO forms in high-temperature

burned gas

CO present at high T

or with fuel-rich mixtures

Outflow of HC

from crevices;

some HC

burns

As burned gases cool,

first NO chemistry, then CO chemistry

freezes

Deposits desorb HC

Piston

scrapes HC

off walls

Oil layers

desorb HC

Entrainment of

HC from wall into

bulk gas

Carbon monoxide (CO) increases rapidly

as the excess air decreases (rich mixtures);

it is very low for lean mixtures.

Unburned hydrocarbons (HC) too are high

for rich mixtures, and decrease as the

air/fuel ratio increases even beyond the

stoichiometric ratio, up to a threshold level

beyond which a fraction of the HC are not

oxidized during the final part of the working

cycle due to the decrease in temperature.

The formation of nitrogen oxides (NOx)

is facilitated by high temperatures and high

oxygen content: maximum emissions are

found for slightly lean mixtures ( 0,9).


SI engines: influence of air/fuel ratio



Air equivalence ratio

Air/fuel ratio

Nitro

gen o

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es [ppm

as N

O]

Unburn

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carb

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Carb

on m

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Lean

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Nitrogen

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Fuel

consumption

Unburned

hydrocarbons

Carbon monoxide


SI engines: typical emissions with no control system


Operating mode -> Idle Acceleration Constant speed Deceleration

CO2 [%] 9,5 10,5 12,5 9,5

CO [%] 2,0 2,0 0,4 2,0

HC [ppm as C1] 4000 2500 2000 20000

NOx [ppm as NO] 100 1500 1000 100

Values given as volume fractions on a dry basis.

All main pollutants are emitted in significant quantities in all operating modes.

Therefore, a pollutant control system that can operate at the same time on CO, HC, NOx is needed.

Source: G. Ferrari, Motori a combustione interna, Il Capitello, Torino, 1996

Exhaust gas recirculation (EGR)

It allows to control NOx formation by

diluting the mixture with inert gases,

thus achieving the same result given

by an increase in air/fuel ratio

without the corresponding increase

in oxygen available.

Three-way catalytic converter

It performs the oxidization of CO, HC and

the reduction of NOx at the same time,

thanks to catalysts (noble metals such as

platinum for oxidization or rhodium for

reduction) capable to promote chemical

reactions even at relatively low

temperature.


SI engines: pollutant emission control


Image source: G. Ferrari, Motori a combustione interna, Il Ca