Professor Dr.-Ing. DrAng. E. h. KARL A. ZINN ER Formerly Director of Research, Diesel Engine Department M.A.N., Augsburg/Germany

Dr. GUSTAVWINKLER Lecturer at the University ofBath/England

ISBN 978-3-540-08544-7 ISBN 978-3-642-52196-6 (eBook) DOI 10.1007/978-3-642-52196-6

This work is subJecllo copyright. All righls are reserved, whelher lhe whole or pari oflhe malerial is concerned, specifically lhose oflranslalion, reprinling, re-use of illuslralions, broadcasling, reproduclion by pholocopying machine or similar means, and storage in data banks. Under ~ 54 01' the German Copyright Law. where copies are made for olher lhan privale use, a fee is payable 10 lhe publisher, lhe amount of lhe fee 10 be delermined by agree-ment wilh lhe publisher.

by Springer-Verlag Berlin Heidelberg 1978.

The use 01' registered names, trademarks. eie. in this publicalion does not imply. even in the absence 01' a speciflc statement, that such names are exempt from the relevant protective laws and regulations and therefore ffee for general use.

2362/3020 -543 210

Preface to the Supplement

The development in turbocharging of internal combustion engines

has proceeded so rapidly that the publishers and I have decided

to add a supplement to the book "Supercharging of Internal Com-

bustion Engines" published in 1978. Since the greatest progress

has been made in the field of turbocharging of automotive engines,

this supplement is confined to this area. When the first edition

of the German original "Auf ladung von Verbrennungsmotoren" was

published in 1975 turbocharging of automotive engines was insig-

nificant except for some types for racing drives. Today nearly

all large car manufacturing companies have turbocharged spark ig-

nition or Diesel engines for commercial vehicles on the market or

to be developed.

The author hopes that this supplement will be welcome as a useful

addition to the basic publication.

Stadtbergen/Augsburg February 1981

Karl Zinner

Contents

12.

12. 1

Problems of turbocharging of automotive engines

Petrol engines .

12.1.1 Measures dealing with combustion knock

12.1.2 Problems of thermal loading

12.1.3 Control problems

12.2 Diesel engines .

12.3 Advantages of supercharging motor car engines

12.4 Examples of current design

12.4.1 Petrol engines

12.4.2 Diesel engines

List of references

3

9

10

17

18

23

23

29

36

12. Problems of turbocharging of automotive engines 12.1 Petrol engines

As mentioned in section 2.3, the supercharging of spark ignition

aero-engines was already highly advanced at the time when they

were superseded by the gas turbine. During and after World War I,

the engines were mechanically supercharged, around World War 11

exhaust turbocharged. There was, however, little incentive to

supercharge automobile engines other than those in sports- and

racing cars, because it was simpler and ehe aper to obtain more

power by increasing the size of the engine.

The first exhaust-turbocharged spark ignition engines for motor

cars became commercially available in 1962. Even though a con:,~ider

able number of such engines, around 60 000 for the Chevrolet Cor-

vair alone /12.1/, were built in the USA between 1962 and 1966,

they were not a breakthrough. Apparently, acceleration was un-

satisfactory, because instead of an improvement in torque backup,

an increase in peak output from 74 to 110 kW, later to 130 kW

was preferred. As compared to aero engines, the turbocharging of

car engines poses some additional problems, while at the same

time some factors favouring the supercharging of aero engines are

lost, as mentioned already in section 11.1.

The additional requirements for car engines are:

- Torque backup with fallinq speed to overcome increased

resistance at inclines without excessive ge ar charging, or to

allow locking of transmissions with hydraulic torque converters

at low engine speeds

Fast response to an increase in power demand without noticeable

delay, i.e. quick rise in boost pressure

- Large engine speed range (see Fig. 12.1)

- Small space requirements,low weight and in par~icular low cost.

One condition for the successfull introduction of turbocharged

car engines is the availability of small inexpensive high-speed

turbochargers with high efficiencies even at small flow rates, a wide operating range and reliability at high exhaust tempera-

tures. The development of rotors with backward swept blades has

led to an improvement in both operating range and efficiency.

...... , \ \ \ \ \ \ \ \ , \ I

mL lorry ( i engine pass. (ar ( i englne pass. (ar s i. engme

Fig. 12.1 Operating charcteristics on the compressor map for lorry

and car diesel engines and for petrol engines. From /12.2/ Fig. 27.

The increase in stresses with the large tip speeds required by

higher pressure ratios has been compensated for by the use of bet-

ter materials and improved casting methods, Fig. 12.2 ,Fig. 12.3.

2

Fig. 12.2 Rotor of a turbocharger of Khnle, Kopp & Kausch Ltd. (KKK)

with radial flow turbine and compressor impeller with backward

swept blades

Fig. 12.3 Corrparison of 3,0

compressor rnaps of impellers I I

with radial (solid lines) 2,8

0 I and backward swept blades '" I 2,6 e I (dashed lines). KKK type 26 ~ Pl '" /---P, VI I VI

2,4 ~ I c- I E I E I 2,2 I / I / 2,0

r--;-. I / I / I /

1,8 I / N Vfo7T, I / r+ min- '

1,6 I I

i / 1,4 4 '

I I I

1,2 I I

1'0 0,05 0,10 0,15 0,20 0,25 0,30 V, ,/To/T,

Today, practically

all of the larger automobile manufacturers are developing turbo-

charged engines, even though commercial introduction is not rushed.

Some of the models presented at the International Automobile

Exhibition 1979 in Frankfurt were put on sale in 1980 only /12.3/.

As compared to the compression ignition engine, there are three

main obstacles to the reliable supercharging of spark ignition

engines for use in automobiles: combustion knock, exhaust temper-

ature and mixture control.

12.1.1 Measures dealing with combustion knock

The most common type of knock is caused by end gas explosion in

regionsof the combustion chamber to which the flame front has not

yet travelled. As the end gas is compressed and heated by the ad-

vancing flame front, the prereactions lead to a sudden combustion

in the end zone. Knocking combustion causes a noticeable increase

in heat transfer, which can destroy the components of the combus-

3

tion chamber which are already highly thermally loaded. The ten-

dency to knock can be reduced by retarded ignition timing, but

this reduces power and increases fuel consumption and exhaust

temperature.

Fig. 12.5 shows the relationship between compression ratio and

permissible boost pressure at the knock limit for a given engine,

indicating the necessity of a reduced compression ratio at larger

boost pressures. The resultant lowering of efficiency can only be

compensated by secondary measures such as optimization of mixture

control and ignition timing and by the relatively 10'..,rer friction

losses at higher mean effective pressures. Fig. 12.4, but even

more so Fig. 12.5 indicate the importance of charge air cooling.

1,9 1,9

bar 1,8 1,8

1,7 1,7 -

1,6 t 1,6 PI PI

1,5 1,5 )(

1,4 1,4

' ,3 1,3

1,2 , I I I I I I 1,2 20 40 60 80 100 120 oe 140 I I

charge air temperature 6 7 E ___ 8

Fig. 12.4 Boost pressure for border-

line knack at optimum ignition

timing as a function of charge air

tenperature, with air-fuel equiva-

lence ratio and actane number as

Fig. 12.5 Effect of campression ratio

and charge air tenperature on the

permissible boost pressure for border-

line knack. From /12.4/ Fig. 3

parameters. From /12.2/ Fig. 7

4

For example, at the low speed of 2500 rer/min and an absolute

boost pressure of 1,5 bar, a compression ratio of 6 is possible

at the knock limit; with charge air cooling to 60 0 C this could be

increased to 8. Charge air cooling is also advantageous with re-

gard to power output and fuel consumption at high speeds because

of the permissible advance of ignition timing as shown in Fig.12.6.

Fig. 12.6 Effect of charge air temper-

ature on the permissible ignition

advance, the pow=r output and the

fuel consurnption. From /12.8/ Fig. 3

b.TDC 30

t 26 'P 1.0.

22

.h 9/KW 340

300

280

~ 'P' ~ ~ ~

"" ~ Pe ~ ~

"" ~ ~

~ V .,-V

kW 220

190

60 70 80 90 100 110C charge air temperature

Although charge air cooling in automobiles requires extra efforts,

there is a growing awareness that these efforts pay back in the

form of increased power and component life and of reduced fuel

consumption and emissions, unless there are other restrietions.

Such restrietions could be limited space not allowing an effective

installation of the charge air cooler in a given vehicle, or the

low performance of a charge air cooler in the free air flow at low

vehicle velocity and high output requirement, for example on long

slopes with heavy trailers. Despite these restrietions, it cannot

be overemphasized that heat transferred in the charge air cooler

does no longer have to be transferred in the engine to the cooling

water; that is to say, the sum of heat transferred in charge air

cooler and engine is constant for a given output. This rule has

5

been confirmed by many experiments. In a fresh design, it may

under certain circumstances even be advantageous to increase the

charge air cooler at the expense of the engine radiator, provided

that the cooling effect is sufficient under all operating condi-

tions. This is possible for example by means of a temperature-con-

trolled electrically driven cooling fan. The heat energy already

transferred in the charge air cooler has a beneficial effect on

the knock limit and the thermal loading, and thereby on the power

output and the engine reliability.

There are four different methods available for eooling the charge

air:

1. Charge air-water cooler in the cooling water eircuit of the engine

2. Charge air-water cooler with aseparate eooling water circuit

3. Charge air-air cooler in front of or adjacent to the engine radiator

4. Charge air-air cooler with separate fan powered by a boost-driven air turbine (Fig. 12.7).

2 3 4 5

6

Fig. 12.7 Air-cooled charge air

cooler by Garret (method 4) .

1 Exhaust manifold 2 Exhaust turbocharger 3 Compressed air duct 4 Cooling fan 5 Flow of charge air 6 Flow of cooling air 7 Charge air cooler 8 Intake manifold

I--

I I

I

I ~ I

The cooling system has to be judged according to effectiveness,

size and weight, space availability, cost, reliability, mainte-

nance requirements and power requirement. The simplest system

6

8

according to 1, is also the least effective, even if the cooling

water coming from the radiator first passes through the charge

air cooler and only then through the oil cooler, if fitted, and

the engine. Investigations on a turbocharged truck engine /12.5/

have shown that method 3, is the most advantageous with regard

to cost effectiveness, followed by 4, which may have advantage of

flexibility when fitted into an already available vehicle type.

Optimization of air-fuel ratio and ignition timing are of prime

importance with regard to the complex interrelationship between

knocking combustion and ignition advance, mixture strength,

boost pressure, charge and cylinder temperature and the effect

of the variables on a power output, fuel consumption and exhaust

missions. Fig. 12.8 shows the wellknownrelationship between fuel

consumption at full power and air-fuel ratio at an ignition set-

ting retarded 40 from the knock limit. Fuel consumption is quite

sensitive to deviations, a rich mixture causing incomplete com-

bustion and a lean one slow burning.

Fig. 12.8 Fuel consumption at full

load as a function of the trapped

air-fuel ratio for an ignition retar-

ded 50 from border line knock.

From /12.2/ Fig. 9

t f e

140

0/0

130

120

110

100

0,6

I

\ / \ / /

0,8 1.0 1.2 Atr ----

Emission control laws in the USA and in Europe differ from each

other (they are more stringent in the USA, in particular as far

1,4

as NO is concerned). Exhaust emissions and fuel consumption cannot

7

be optimized simultaneously (see Fig. 12.9), and cars for the

American market therefore have to be equipped differently from

those in Europe. Usually, engines for the USA are built with

a tripie catalyser after the exhaust turbine and an oxygen probe,

a so-called lambda indicator in the exhaust gas stream /12.7/.

This probe controls the air-fuel ratio, keeping it to within

very narrow limits of stoichiometric, which allows emissions to

be reduced below the legal limits by means of the catalyser.

The US-versions are usually sold with a somewhat lower peak

output.

Fig. 12.9 Road fuel consurnption as

a function of vehicle velocity for

naturally aspirated (dashed line)

and turbocharged engines of equal po\o.er

tuned for best consurnption (without

catalyzer,solid line) and for best

emissions (with catalyser, dash-

dotted line). From /12.6/ Fig. 7

18 I

100 km

16

t 12

B

6

40 60 80 100 120 v

I r

140 km/h

Present European emission standards do not necessitate the use of

a catalyser, but here too fuel consumption and emissions have to be

optimized. As knocking combustion must not occur under any

operating conditions, the ignition timing which is usually con-

trolled only by engine load (carburetor vacuum) and engine speed

(centrifugal control) must allow for a safety margin with regard to

the knock limit. This margin can be reduced with electronic ig-

nition timing (/12.8, 12.9, 12.10/), which takes into ac count fur-

ther parameters such as engine and air temperatures. A microcom-

puter calculates the optimum values of fuel rate and ignition

timing according to the parameters measured by the probes, using

8

a pre-programmed map that had been determined experimentally.

Part load operation with lean mixtures is thus possible, resulting

in redueed fuel eonsumption and exhaust emissions. Sehemes of

this kind are also possible for naturally aspirated engines, but

they are partieularly suitable for supereharged engines with a

wider range of eonditions from part to full load operation.

Further improvements going in the direetion of a elosed-Ioop eon-

trol to aehieve an optimization of the eomplex interaetions in

the spark ignition engine are investigated /12.11/.

12.1. 2 Problems of thermal loading

The temperature stresses in the turbine rotor assoeiated with

high exhaust temperatures ean be eheeked satisfaetorily by the

use of modern heat-resistant materials. The exhaust gas tempera-

tures of spark ignition engines are eonsiderably higher than

those of diesel engines, but the exhaust eontains little oxygen

and is therefore less eorrosive. Small radial turbines usually

have no bladed stator, the eorreet ineidenee angle at the rotor

tip being aehieved by the dimensioning of the inlet serolI.

Rather, problems with high temperatures may oeeur in the turbine

easing, the exhaust manifold or the bypass valve. A eomplete

separation of the exhaust gas from individual groups of eylinders

all the way up to the turbine rotor, i.e. pure pulse turboehar-

ging, is not yet possible beeause of the high exhaust temperatures.

Inlet serolls with double entry do not yet have suffieient

durability (thermal fraetures) /12.6/. Separating walls and lips,

whieh would keep the gas flows apart, are burned off beeause

there is no possibility for eooling /12.12/. Exhaust turboehar-

gers for spark ignition engines therefore operate with full

admission up to now. HO\\ever, if spaee allows one to collect the ex-

haust from groups of eylinders with a suitable firing order into

individual manifolds, it is advantageous to keep these manifolds

separate up to some point before the turbine entry /11.2,11.19/.

Based on developments for automative turbines, ceramie compo-

nents that can withstand the high temperatures are also being

developed for turboehargers.

For example, ceramic inlet easings would probably make pulse

turbocharging possible. A eonical diffusor irnrnediately after the

turbine would inerease its effieiency /12.6/. To counter the

increased heat flow in the engine eomponents, natrium eooled

9

valves for spark ignition engines, oil cooling of the piston by

a fixed jet and, if necessary, a larger cooling water pump to in-

crease the flow are used.

12.1.3 Control Problems

Unlike the diesel engine, the spark ignition engine is quantity-

controlled, requiring a throttle plate to control the charge

quantity. The throttle plate can be located either in front of or

after the charge compressor, in either case, there are advantages

and disadvantages. The advantage of a throttle in front of the

compressor is that the compressor will not surge when the throttle

is suddenly closed during gear changes or when the vehicle pushes

the engine. When the throttle is partly closed, the pressure level

before compressor is reduced, moving the operating point in the

compressor map for a given mass flow rate to the right and in-

creasing the turbocharger speed at power equilibrium between

turbine and compressor. Fig. 12.10 illustrates the difference in

turbocharger speeds for throttles in front of and after the com-

pressor. In the first case, the turbocharger runs at a higher

Fig. 12.10 Turbocharger speeds with

throttle plates placed either before

(upper lines) or after corrpressor

(lov.er lines) as functions of engine

speed and load. From /12.2/ Fig. 40.

10

90000

80000

70000

t 60000 Ne

40000

30000

10000

2000 3000

initial speed and accelerates better. With carburetor engines,

a throttle in front of the compressor has the following advan-

tages:

- The same system as with the naturally aspirated engine

can be used

The tuning of the carburetor is easier

- The turbulence in the crnpressor homogenizes the charge

- The evaporation of the fuel reduces the charge temperature.

However, the disadvantage of a throttle plate in front of the

compressor is that the latter must be absolutely oiltight; other-

wise, oil would be sucked into the engine at part load conditions.

This would result in loss of oil, fouling and, in particular,

unburned hydrocarbons in the exhaust gas, because the oil sucked

in would not burn completely. Carbon rings can seal the compres-

sor shaft oil-tight, but have the disadvantage of increased wear

and friction losses. If the carburetor is located after the

engine, it has to be pressure tight and is usually of a two-stage

design to simplify tuning.

If the throttle plate is placed after the compressor, labyrinth

seals with practically no wear can be used in the turbocharger.

To avoid compressor surge when closing the throttle, apart-load

discharge valve (recirculation valve) is often used which allows

part of the compressed air to return to the compressor intake.

The recirculation must not affect the metering of the air mass

flow rate; see Fig.12.11 There may also be an additional vacuum

limiter to prevent the backfiring of unburned lean mixture in the

exhaust system when the engine is being pushed by the vehycle.

No vacuum limiter is necessary if the ignition timing is greatly

advanced under these conditions, e.g. by means of electronic

timing.

An air recirculation valve, which is often integrated into the

turbocharger as shown in Fig. 12.12, is,according to Hiereth,

/12.2, 12.4/ unnecessary if the exhaust bypass is controlled by

the pressure in front of the throttle plate (see Fig. 12.13).

It seems, however, that this simplification depends not only on

such a control mechanism, but also on the position of the surge

line in the compressor map and on the design of the charge air

manifold.

11

24

Fig. 12. 11 Flow scherre of air and exhaust gas of the Porsche 924 turbo

eng ine wi th throttl ing after the turbocharger. 1 air filter; 2 mixture

control; 3 induction pipe; 4 turbocharger (compressor); 5 air recirculation

valve; 6 charge air duct; 7 throttle plate; 8 intake manifold; 9 fuel

injection lines; 10 exhaust manifold; 11 turbocharger (turbine);

12 exhaust pipe; 13 exhaust rnuffler; 14 exhaust silencer; 15 waste gate;

16 boost pressure control line (waste gate); 17 ventilation; 18 boost

pressure controlline (air recirculation valve); 19 bypass air valve;

20 vacuum limiter; 21 air line to bypass air valve and vacuum limiter;

22 interconnection; 23 vacuum controlline; 24 boost pressure control

switch /12.12/

12

1

7

1

Fig. 12.12 Exhaust turbocharger by KKK, type K 26, with air recirculation

valve integrated into the corrpressor casing. 1 air intake; 2 corrpressor

rotor; 3 shaft; 4 turbine casing; 5 turbine rotor; 6 journal hearings;

7 air mtiet; 8 air recirculation valve /12.12/

Fig. 12.13 Scheme for boost pressure

control, with throttie plate located

behind the corrpressor. 1 engine ;

2 turbocharger; 3 waste gate;

4 charge air cooler; 5 air filter

6 Exhaust manifold; 7 intake manifold;

9 throttle plate.

From /12.2/ Fig. 36

Fig. 12.13 also introduces one to the control of the exhaust tur-

bine, which is absolutely necessary for automotive engines. As

13

already mentioned, the motor car puts high demands on the torque

characteristic and the acceleration of the engine, as weil as on

the speed and mass flow range of the compressor. The fundamental

problems posed by the exhaust turbocharing with regard to accele-

ration and torque characteristic, and methods to meet these re-

quirements were already discussed in sections 8.2, 8.3 and 8.4.

All motor car engines offered for sale at the moment use exhaust

bypass control; it appears that small turbochargers with variable

geometry are either unreliable or too expensive. With bypass con-

trol the turbocharger is designed for a flow rate corresponding

to an average engine speed; the compressor and in particular the

turbine are much smaller than would be sensible in order to achieve

maximum power at full engine speed. A small turbine flow area is

necessary not only with regard to the torque characteristic, but

also to obtain good acceleration. The effect of the turbine flow

area on the boost pressure is considerable /12.2/.

At full engine output, an uncontrolled turbocharger of this design

would overspeed and overboost, with all the consequences for the

loading of turbocharger and engine. f the various bypass arrange-

ments,such as discharging the air from a certain boost pressure

onwards, or bypassing the exhaust gas under the control of the

back or boost pressure, the first is no longer in use. As ex-

plained in publications such as /11.2/, the bypassing of exhaust gas

is thermodynamically advantageous, because it reduces the back

pressure at full load as compared to the discharging of air.

The back pressure, with exhaust bypass control at high engine

speeds usually higher than the boost pressure, increases the dis-

placement work of the piston and the amount of exhaust gas trapped

in the cylinder. This in turn increases the fuel consumption.

With exhaust bypass control, the turbine can be made smaller

because of the smaller flow rate, resulting in higher efficiences

and better acceleration from part-load conditions. In addition,

the turbine blading is frequently designed to give an unsymmetri-

cal efficiency characteristic with respect to u/co' slanted

towards higher efficiencies at small flow rates.

The exhaust bypass valve, also called waste gate, can be integra-

ted into the turbocharger or it can be attached to the exhaust

14

manifold. Garrett-AiResearch have developed two different rnethods

for the control of this valve (see Fig. 12.14), one for dry air

in the compressor (for diesel engines and petrol engines with

fuel injection into the intake man[old) and the other for layouts,

in which the carburetor is located in front of the compressor

/12.13/. In the first instance, the actuator of the waste gate is

integrated into the valve and located at the turbine casing; in

the second instance, the actuator is located on the compressor

casing and seperated from the valve.

Fig. 12.14 Exhaust turbocharger by Garrett AiResearch with integrated

waste gate and control pressure tap on the seroll. Dry and W8t versions

for air (left) and air-fuel mixture (right).

The diaphragm of the actuator, made of Viton, has only limited

heat resistance, and the valve, whether integrated or not,

should therefore be located in the air stream of the cooling fan

and be equipped with fins. It can also be useful to cool the

diaphragm by the means of air bled from the control line, as long

as this does not affect mixture control; the air must not contain

fuel. In the design for diesel engines shown in Fig. 12.15, the

cooling air passes through a bore in the valve shaft to the low-

pressure side and is discharged through a suitable passage together

with the exhaust gas.

The separate installation of the bypass valve offers the advantage

15

Fig. 12.15 Section of an exhaust

waste gate with cooling bores,

by Garrett.

Fig. 12.16 Exhaust waste gate can-

trolled by the engine back pressure P3

and additionally by the static pres-

sure P1st at the compressor inlet, for

the Audi 200.

of better cooling in the air flow of the fan. The valve can also

be controlled by the engine back pressure instead of the boost

pressure. In this case, a longer control line becomes necessary

to allow the exhaust gas to cool down. As the rate of flow only

caused by leakage is very small, the cooling of the exhaust gas

poses no particular problems. Fig. 12.16 gives an example of a

bypass valve controlled by the back pressure, in which the con-

trolling pressure acts in the same direction as the back pressure

against the spring force. In addition, in this design for the

5-cylinder engine of the Audi 200 /12.19/, the static pressure at the

compressor inlet acts on the upper side of the diaphragm. For a

given total pressure, i.e. the ambient pressure, the static

pressure decreases with the rate of air flow so that the boost

pressure drops again with increasing engine speed as shown in

Fig. 12.17. Maximum boost is thus generated only in the region of

16

Fig. 12.17 Boost pressure charac-

teristic of the Audi 200.

bo r

0,8

0,4

0,2

o

I'~ f-Kurbo - Europe

r-1/

/ V

2000

\ '-..

- ....... 1'. turbo - USA "

4000 NE -

1\, ....

maximum torque. Aboost pressure rising with falling engine speed

can also be obtained by means of resonance pipes between compres~

sor and engine tuned to a low engine speed in addition to the

bypass control.

12.2 Diesel engines

Because of the differences already mentioned - no knock limit,

low exhaust temperatures, quality control - the supercharging of

motor car diesel engines is much simpler than that of petrol

engines. The same rules apply as far as turbocharger matching and

control are concerned. Because of the low exhaust temperatures,

pulse turbocharging combining suitable cylinders is possible.

Again, the large range of engine speeds makes waste gating

necessary. With regard to cost, manifold layout and space re-

quirements, only one bypass valve is usually employed. The branches

of the exhaust manifold are in this case combined be fore the tur-

bine, giving full admission to the turbine. If the torque

characteristic is more important than peak power, charge air

cooling is unnecessary.

To avoid smoke at low boost pressures, the fuel injection pumps

are equipped with aboost pressure-controlled fuel limiter.

Alternatively, the fuel limiter can also be controlled by the

absolute pressure to take into account the decreasing air density

at higher altitudes.

17

12.3 Advantages of supercharging motor car engines

The advantages claimed here are the same as for supercharging in

general: Reduced weight and space requirements for a given power

output as compared to the naturally aspirated engine, lower cost

per unit output, higher efficiency for diesel engines in particu-

lar, larger operating range for a given engine, smaller radiator for a given output, reduced noise and emissions. The two factors

mentioned first initially carried the greatest weight, so that

supercharging was introduced in the engines of racing and

sports cars. A small turbocharger with speeds up to 150000 rev/min

boosting the engine output by 40 kW to 110 kW weighs about 6 kg.

Mounted directly onto the exhaust manifold, it requires no sup-

port. Despite the additional waste gate and the extra air and

exhaust manifolds, the ratio of additional power to additional

weight is much more favourable than the corresponding ratio of

naturally aspirated engine. Volkswagen /12.14/ state that the

mass-to-power ratio of a 1.5 liter naturally aspirated diesel

engine with an output of 37 kW is 3 kg/kW, whereas for the same

engine turbocharged to 55 kW it is only 2,4 kg/kW. The latter

engine has a turbocharger with integrated waste gate, but no

charge air cooler.

Statements about savings are more difficult to come by. Small

turbochargers manufactured in large numbers are by themselves

quite cheap, but the necessary alterations to the manifolds

and the engine (such as oil-jet piston cooling) and possible

alterations in the engine compartment affect manufacturing costs.

The penetration of turbocharged engines in motor vehicles has

been given aboost by new emission standards and the energy

crisis. Measures to limit the emissions of petrol engines by means

of lean mixtures and retarded ignition timing have resulted in

apower reduction that can be more than compensated with turbo-

charging. More expensive engines with larger displacements are

therefore not necessary. For economic reasons, too, operation

with rich mixtures to obtain maximum power should be avoided.

Supercharged petrol engines, too, may have a lower fuel consump-

tion than naturally aspirated engines of equal power because of the

18

reduced friction losses of a smaller engine, particularly in the

lower output range (see Fig. 12.18). At full load, fuel con-

sumption is not always lower because of the compromise between

fuel consumption and acceleration presently made. A small turbo-

charger or a narrow inlet scroll improve acceleration, but

30 l

100km

25

t 20 f

15

10

5

n.Q. engine

! I

12. 18 Road fuel consurrption of a

naturally aspirated and a turbocharged

engine of about equal po1r.Br at constant

velocities. From /12.2/ Fig. 52 40 80 120 160 km/h 200

v_

increase fuel consumption at full load. At part load, the

turbocharged engine is always more efficient; at full load, it

all depends on the design compromise.

Fuel economy is measured on the road, exhaust emission in a

standardized test. As a rule, the supercharged engine leads

in both measures; see for example Tables 12.1 and 12.11.

In Fig. 12.19, the fuel consumptions of two naturally aspirated

petrol engines of different outputs are compared with those of

two diesel engines, one supercharged, the other not, but all

mounted in the same vehicle. Clearly, consumption of the turbo

diesel is considerably reduced at higher speeds.

In F ig. 12.9, the difference in fuel consumption between opti-

mizing with respect to fuel economy or emissions had already

been illustrated. Even the emissions-optimized engine showed a

slightly reduced consumption at part loads. In addition,

19

Table 12.1 Road fuel consumption of naturally aspirated

and turbocharged engines of equal output.Fram /12.2/

Course n.a. engine t.c. engine

Description Nr. distance km liter Itr/1oo km liter ltr/1oo km

Inner city 1.4 88 17.58 19.9 25.60 17.7 1 passenger top speed 50 km/h

Suburb 1.1 100 17.92 17.9 15.91 15.9 1 passenger

top speed 50 km/h

City 1.4 4.3 80 15.00 18.7 13.06 16.3 trial course 1 passenger

top speed 50 km/h

Black forest. 6.0 340 50.05 14.7 46.90 13.7 rrax. gross weight

Table 12.11 Exhaust emissions of naturally aspirated and

turbocharged engines of equal output for the

Europa test without reactor and for the CVS test.

From /12.2/.

naturally aspirated turbocharged CVS HC g/mile 4.46 3.05

CO g/mile 22.08 15.71 NO g/mile 5.71 3.86 fuel cons.ltr/loo km 19.9 16.7

Europa HC g/test 5.32 4.71 CO g/test 116.92 68.93

No g/test 8.01 5.05 fuel cons. ltr/loo km

20

Fig. 12.19 Road fuel consumption of a 14

diesel, a turbocharged diesel and tv.u

petrol engines in the same vehicle at

constant velocities. From/12.14/ Fig. 6

100 km

12

10

t 8 6

2

o 40 60 80

v

Table 12.111 indicates that the consumption-optimized super-

charged engine has lower CO emissions even without a catalyzer.

The large amount of hydrocarbons was traced in this case to

an oil leak in the compressor. Noise emissions, too, have been

reported by several workers to be reduced with supercharging,

particularly as far as exhaust noise is concerned. The turbine

muffles the lower frequencies which are otherwise more difficult

to control. Thus, a simpler muffler with decreased flow resistance

may be used.

Table 12. III Exhaust emissions of naturally aspirated and

turbocharged engines in g/test (ECE). From /12.6/

naturally aspirated turbocharged ECE standard

CO

HC

NO x

92

6.2

with without catalyser

15.2 40.6 122

0.6 17 .8 8.6

2.6 7.7 14

As far as acceleration is concerned, the comparison has to be

based on engines of equal maximum output in the same vehicle,

21

not on the same engine in a naturally aspirated and a super-

charged version. The lagging of the boost pressure in the turbo-

charged engine cannot be avoided altogether. The pick-up of the

boost pressure depends mainly on the turbine design, i.e. its

flow area and the inertia of the rotor. According to Hiereth

/12.2/, the lag might be acceptable if the pressure pick-up was

not delayed by more than 0.5 seconds. A comparison between the

accelerations of a naturally aspirated and a turbocharged engine

is not unfavourable for the latter (see Fig. 12.20). According

Fig. 12.20 Acceleration of a natural-

ly aspirated and a turbocharged petrol

engine of about equal po\\er men going

through the gears. From/12.2/ Fig. 51

20 0

kmj h

100

50

J. 1/ /

~ -----/.",,/" V/ /

-- t.c.-petrol engi -- n.a.- petrol engi

ne ne

o 10 20 t

30 s 40

to /12.8/, the lag in boost pressure and output may even be

advantageous, because the danger of spinning of the driving wheels

is reduced for powerful cars. It is likely that the driver has

to adjust to the response of the turbocharged engine; after this

adjustment, he will not be aware of any disadvantages.

Volkswagen have shown with their experimental vehicle /12.15/,

that the turbocharged diesel engine can reach good values of fuel

consumption and exhaust and noise emissions. The combustion

process was optimized for emissions, nitrogen oxides were reduced

by means of exhaust recirculation, and the engine was fully

enclosed. The practical application of such an optimized engine

is, however, still hampered by considerably higher production

costs.

22

12.4 Examples of current design

12.4.1 Petrol engines

From among the many types of turbocharged engines presently manu-

factured or developed, only a few examples can be briefly des-

cribed here, preferably those of European origin for which tech-

nical publications are available. Further details can be found

in the publications mentioned.

The company of Dr.h.c. Porsche Ltd. initially developed turbo-

charged engines for racing cars. The model 917 was mentioned as

an example in the first edition of this book. In 1975, the model

924 was introduced with a watercooled, naturally aspirated 2-1itre

engine of 92 kW/11.12/. The demand for greater engine power lead

to the 924 Turbo, Fig. 12.21, which is produced in a European

version of 125 kW using premium gasoline and an American version

of 110 kW using regular unleaded gasoline and an exhaust cataly-

ser to meet the more demanding emissions standards. The exhaust

turbocharger by KKK type K 26 has an integrated air recirculation

valve and aseparate exhaust waste gate as shown in Fig. 12.12.

The additional output of 33 kW is obtained with an extra mass

of 29 kg (turbocharger including recirculation valve, additional

manifold, no charge air cooler) .

Fig. 12.21 Four-cylinder Porsche engine , IlDdel

924 turbo. Waste gate

on left.

23

Based on the air cooled flat-six engine model 911, so called 3

production racing cars model 935 with 2857 cm displacement and

441 kW output, and model 936 with 2142 cm3 and 382 kW were deve-

loped /12.16/. Both engines have turbochargers for each bank as

shown in Fig. 12.22, with turbines triple-pulse operated under

full admission. The model 935 has air-water charge air cooling,

the model 936 an air-air cooler.

Fig. 12.22 Six-cylinder Porsche

engine, IIDdel 935, with charge air

cooler.

The Swedish company Saab were the first to produce a passengercar

with a turbocharged petrol engine for the European market. The

models 99 turbo and 900 turbo have a four-cylinder engine of

1985 cm3 displacement and compression ratio 7.2:1, producing

106 kW at 5000 rev/min; see Fig.12.23 /12.17/. The torque peaks

at 3000 rev with 225 Nm, falling to 190 Nm at 5500 rev/min. A

Garret-AiResearch turbocharger with integrated waste gate con-

trolled by the engine back pressure is used. The four cylinders

have a common exhaust manifold for constant-pressure turbochar-

ging. Maximum boost is 0.7 bar gauge; the engine is equipped

with natrium-cooled valves and mechanical fuel injection (Bosch

K - Jetronic).

At the IAA in Frankfurt in 1979, both Bayerische Motorenwerke

(BMW) and Audi NSU Auto Union presented turbocharged models for

sale in 1980. The BMW engine has six cylinders of 3.2 litre dis-

placement, producing a peak output of 185 kW at 5200 rev/min for

24

Fig. 12.23 Four-cy 1 inder

Saab engine, rrodel

900 turbo.

the European marked without exhaust reactor; see Fig. 12.24 /12.18/.

Peak torque is 380 Nm at 2600 rev/min, falling to 340 Nm at full

speed. The exhaust manifold is shared by all six cylinders.

The KKK turbocharqer is not combined with the waste qate: for retter coolinq,

the latter is attachEd to the exhaust manifold and controlled by the

boost pressure picked up at the compressor spiral casing. The

charge air, leaving the compressor downward, passes through a

charge air cooler placed in the air flow, the throttle and the

air distributor into six resonater pipes of equal length. In order

to avoid compressor surge, an air recirculation valve is placed

in front of the throttle. Fig. 12.25 illustrates the flow scheme

of air and exhaust gas. The charge air is cooled by more than

400 C at the higher speeds. The operating point on the compressor

Fig. 12.24 3.2 ltr

BMW six-cylinder petrol

engine with turbocharger

25

Fig. 12.25 Flow scheme of air and

exhaust gas of the BMW engine /12.18/:

air filter; 2 air flowrneter;

3 compressor; 4 charge air cooler;

5 throttle; 6 intake manifold;

7 exhaust manifold; 8 exhaust tur-

bine; 9 exhaust pipes; 10 waste gate;

11 controlline; 12 bypass pipe;

13 air recirculation valve;

14 air bypass

26

3'0Ir===:C:::~L, standard conditions

2, Po . 981 mbar T 293K

0,05

volume flow rate

0,30

Fig. 12.26 Corrpressor map of the

turbocharger KKK K27 with super-

inposed operating line of the

BMW engine /12.18/

map is shown in Fig. 12.26

0.5 bar gauge.

the boost pressure at 185 kW is only

The engine is equipped with Bosch L-Jetronic injection, which

serves three additional functions on the turbocharged engine:

Overspeeding cut-off, prevention of backfiring in the exhaust

pipe during deceleration down to speeds of 1200 rev/min, and over-

boosting control, all by cutting off the impulses to the fuel ih-

jection valves. Fig. 12.27 is a sectioned isometrie drawing of the

turbocharged Audi five-cylinder engine /12.19/. With few exceptions,

the components of this 2.144 litre engine are identical with those

Fig. 12.27 Isometrie section of the turbocharged

five-cylinder engine by Audi, rrodel 200 /12.19/

of the naturally aspirated engine: The crank case requires some ad-

ditional machining in order to attach the oil jets for piston

cooling, among other things. The bowl of the piston is enlarged

to reduce the compression ratio to 7:1, and the exhaust valves are

natrium-cooled. The exhaust manifold, shown in Fig. 12.28, is cast

in one piece of austenitic modular iron with an additional flange

for the waste gate, and contains separate channels for cylinders

1, 2 + 5 and 3 + 4. Thermal expansion had to be taken into ac count

27

Fig. 12.28 Exhaust manifold of the

Audi 200 /12.19/.

in this design. The channels are separate until immediately in

front of the turbine, narrowing down to produce a pulse-converter

effect. The turbocharger by KKK, type K 26, has neither acharge

air cooler nor an air recirculation valve. The engine has con-

tinuous fuel injection (Bosch K-Jetronic).

The separate waste gate is attached to the exhaust manifold; its

design, shown in Fig. 12.16, and its control and boost characteri-

stic were already described in section 12.1.3 (Fig. 12.17). As

usual, the American version differs from the European version

because of the unleaded petrol and tighter emission standards,

having a so-called lambda transducer, a catalytic exhaust reactor

Table 12.IV Specifications of the turbocharged Audi five-

cylinder engine

Version Europe USA

Stroke/Bore mn/mn 86.4/79.5 3

displacenent cm 2144

nominal output kW 125 100

at speed rev/min 5200 5400

peak torque Nm 265 202

at speed rev/min 3200 3000

conpression ratio 7: 1

max. boost bar gauge 0.82 0.38

fuel SUper Regular

dry engine \\Bight 186.4 incl. clutch, starter etc.

28

and a lower boost limit. A comparison of the two versions is given

in table 12.IV. The only passenger car engine with so-called

combined turbocharging to become known is the re-design of the

3-litre 6-cylinder BMW engine by ALPINA Burkhard Bovensiepen KG,

Buchloe. It produces the notable output of 221 kW, or nearly 100 hp

per litre /12.8, 12.20/. The arrangement of the resonator pipes and

the receiver volume can readily be seen in Fig. 12.29.

Fig. 12.29 Arrangement of resonator system of the turbocharged ALPINA engine B7 /12.20/

An exhaust turbocharger by KKK type K 27 is used; the separate

waste gate reduces the thermal problems. In contra-distinction

to the usual arrangement, the air flow controlling the fuel in-

jection is mete red after the compressor to keep the inlet pressure

loss low. The compressor is more sensitive to throttling on the

inlet than on the outlet; also, the denser air after the compres-

sor produces larger control forces on the stagnation plate or

baffle. Fig. 12.30 illustrates the power and torque characteri-

stics of the 3 litre ALPINA engine, both with tuned intake mani-

fold and with turbocharging combined with resonator. Fig. 12.31

shows the difference in acceleration.

12.4.2 Diesel Engines

The turbocharging of passenger car diesel engines advances only

slowly, even though it causes fewer problems than that of petrol

engines and had been used early in diesel racing cars in the USA.

29

220 kW

200

180

- bo ost pressure 0.90 bar - - boost pressure 0.55 bar

/ /

40 -- tuned intake manifold -- resonator

--- / /

/ I 30

t 4th gear~ -{/

/ // 1/

100

80 I

,Ir r-

1I

60

40

20

o 2000

'/ / I /

/ I

~ r--_ --

4000 NE -

r----. -,

J ,

Nm 500

400 t 1:

300

200

6000 min- I

Fig. 12.30 Output and torque charac-

teristics of the turbocharged ALPINA

engine c.f. Fig. 12.29 /12.20/

t s -

20 -

10

)

W o 2000

V ~

//

~ ~ ~ .....

3000 N E

Y V//

V /

/

3r~~ ~ ...;.-

4000 min-1 5000

Fig. 12.31 Acceleration of the

BMW ALPINA B7 turbo with tuned

manifold (solid lines) and with

resonator system (dashed lines)

Fig. 12.32 shows the engine type OM 617 A /11.21/by Dairnler-Benz

for the Mercedes 300 SD. This is the turbocharged version of the

type OM 617, a five-cylinder engine of 91 rnrn bore and 92.4 rnrn

stroke giving a displacement of 3005 crn3 . The output of the na-

turally aspirated engine of 59 kW at 4000 rev/rnin is raised to

Pig. 12.32 Turbocharged five-cylin-

der diesel engine type OM 617A for

the Dairnler Benz nodel M=rcedes

30

85 kW at 4200 rev/min by means of turbocharging without charge air

cooling. The small size of the turbocharger stands out in compa-

rison to the air filter.

The same engine, but with charge air cooling as a "record engine",

was used in the experimental vehicle C 111 111 in order to prove

the reliability of the engine through aseries of endurance world

records /12.22, 12.23/.

The output and torque characteristics of the naturally aspirated

engine OM617, of the turbocharged version OM617A and of the

record engine are shown in Fig. 12.33. The engine has a common

exhaust manifold for all five cylinders, a number which is un-

favourable for pulse turbocharging. The production engine has a

Garret AiResearch turbocharger type T03 with integrated waste

gate, but there is no waste gate on the different turbocharger of

Fig. 12.33 OUtput and torque

characteristics of the diesel engines

OM 617 (dashed), OM 617 A (dash-

dotted) and OM 617A Rekord (solid

lines) /12.22/

11.0 kW 120

100

t 80 Pe

60

40

20

.,/ ~ L'~

/,fo' /

/v .......

/ / .-,-

~'-1./ ....-1-- '-1--

1.00 N,m

350

o // r-.

'" V 300 t /' .-/ T-'

V V L-i-- --

V

1000 2000

-. -'--- r-_

3000 NE -

the record engine in order to produce the greater output.

'- .......... , -. r- ..........

I J

4000 min- 1

The difference in output and in torque in particular is clearly

shown in Fig. 12.33. In the production engine, the waste gate li-

t

250

200

150

100

31

mits the boost pressure ratio to 1.75, but in the record engine it

can reach a value of 3.3. Without wastegating, this pressure ratio

is obtained at a back pressure ratio P3/P 4 = 2.7 (Fig. 12.34). The fuel consumption of the turbocharged version is in all tests

lower than that of the naturally aspirated engine, and the same

applies to CO and HC emissions /12.21/. The slightly increased

emission of NOx could possibly be reduced below that of the natu-

rally aspirated engine, if charge air cooling were used (see section

8.6.1 and Fig. 8.27). The engine type OM 617 A-Turbo was initially

available only in the USA, but is sold everywhere from 1980 on-

wards.

Fig. 12.34 Pressure ratios and

ternperatures at the turbocharger of

the OM 617A Rekord engine.

3, 5

3, 0

2,5

2,0

1,5

1,0

,ti

100

50

o

-

e:;

~

.".-~ p,/p,

/ ,....-- Pl/ P4 V V

/

~ V ? ~ t

..... t)

/""" !-- t 4

V V t, -

V V -- t' / , ./ ,./' V L .........

-::::: -- t,

J ; tL 900 oe 700

500

300

P2/P1= compressor pressure ratio;

P3/P4 = turbine: t 2 = charge air

temperature behind cooler /12.22/ 1000 2000 3000

E

4000 5000min- '

The 1.5 litre turbocharged diesel engine of the Volkswagen Golf

is included here, even though it is not yet in production. Reports

containing detailed and remarkable test results are already

available; as far as is known, this is presently the smallest tur-

bocharged diesel engine with the highest output relative to its

displacement. The vehicles existing with this turbocharged engine

are all research objects, developed in conjunction with the U.S.

32

Department of Trade (DOT) or the German Federal Ministry of Re-

search and Technology (BMFT) in order to demonstrate the technical

feasability. In addition to the publications /12.14, 12.15/ al-

ready mentioned, areport is also available by the DOT /12.24/

describing the results of tests regarding fuel consumption,

exhaust emissions, noise level, top speed, acceleration and so on,

of several prototype VW diesel engines with and without turbo-

charging that have been installed in vehicles of various sizes and

weights.

The four-cylinder engine with swirl combustion chamber (see Fig.

12.35) has a bore of 70 mm and a stroke of 80 mm and is equipped

with a Garret-AiResearch turbocharger type T 3 with integrated

waste gate. The output is 55 kW, an increase of 50 % over the

naturally aspirated engine with 50 hp (37 kW). Stiff emissions

regulations, for example the NOx limit of 1 g/mile valid in the

Fig. 12.35 View of the Golf turbo diesel engine by vw

USA from 1980 onwards, require a slight reduetion of output be-

cause exhaust gas recirculation is used. Table 12.V summarizes

the main spezifieations of the fully enelosed Golf engine with

exhaust gas recireulation. It remains an open question whether the

NOx limit could also be met by charge air eooling instead of

exhaust gas recirculation which has some drawbacks regarding fuel

consumption and CO emissions.

33

Table 12.V

Engine

Output

Emissions

Specifications of the VW-Golf experimental car

with turbocharged diesel engine, fully enclosed,

with exhaust gas recirculation. Frorn /12.14/, Fig. 11.7

Swirl combustion chamber, displacement 1.5 ltr,

turbocharged diesel with wastegate

51.5 kW at 5000 rev/min

specific weight 2.52 kg/kW

max. torque

top speed

acceleration

HC

US g/mile 0.11

ECE g/test 0.6

125 Nm at 3000 rev/min

160 km/h

o .. 100 km/h in 13.55 s

CO NO particles x 0.8 0.9 0.25

2.33 2.95 0.6

Fuel consumption ltr/100 km

US canp. 4.7 ECE 6.4 DIN 6.3

Noise level lSO-R 362 71dB(A) SAE J 958A 66dB(A) idling 59dB(A)

Another turbocharged diesel engine already available commercially is the XD 2S by Peugeot(Fig. 12.36). The fourcylinder engine of

Fig. 12.36 View of Peugeot turbo

diesel engine XD25 with OOost

pressure contra! system /12.25/

boost pressure control system

94 mm bore, 83 mm stroke and 2304 cm3 displacement has an output

of 59 kW and a torque of 120 Nm at 4150 rev/min. The maximum

34

torque of 180 Nm at 2000 rev/min is 50 % higher than that at full

speed /12.25/.

The engine is equipped with a Garret AiResearch turbocharger type

T03 with integrated wastegate that limits the boost to 0.6 bar

gauge. All of the Peugeot diesel engines use the swirl chamber

combustion system Ricardo-Comet Mk 5.

The Italian company Stabilimenti Meccanica VM in Cento which

originally built mainly diesel engines for industrial use, has

now developed aseries of engines with 4,5 and 6 cylinders of be-

tween 1995 and 3589 cm3 displacement which in their turbocharged

HT versions are intended for installation in passenger cars /12.26/.

The smallest four-cylinder turbocharged diesel Type 488 HT with a

bore of 88 mm and a displacement of 1995 cm3 produces an output of

63 kW at 4300 rev/min and a peak torque of 260 Nm at 2500 rev/min.

The engines have KKK turbochargers with integrated wastegates and

Bosch distributor-type injection pumps. A view of the HR 488 HT

destined for the Alfa Romeo "Alfetta" is given in Fig.12.37.

Fig. 12.37 View of VM turbo diesel

engine 488 HT with KKK K24 turbo-

charger /12.26/

Other turbocharged diesel engines for passenger cars are under

development, such as those BMW are developing: alone a 2.4 litre

engine and another one together with Steyer-Daimler-Puch and the

AVL of Prof. List /12.27/.

35

List of references

/12.1/ McInnes, H.: Turbochargers. Editor and Publisher: Bill Fisher, USA 1976

/12.2/ Hiereth, H.: Untersuchung ber den Einsatz aufgelade-ner Ottomotoren zum Antrieb von Personenwagen. Diss. TU Mnchen 1978

/12.3/ Bahr, A.: Fahrzeug-Dieselmotoren mit Abgasturboladern auf der IAA 79. MTZ 40 (1979) p.606/610

/12.4/ Hiereth, H.: Besonderheiten und probleme des Ottomotors mit Abgasturboaufladung. Automobil-Industrie 2/79, p. 19/25

/12.5/ Marion, G. und Bidault, M.: Recent evolution in turbo-charging diesel engines for truck application. Conference on Turbocharging and Turbochargers, Inst. of Mech. Engineers, London 1978

/12.6/ Spindler, W.: Matching a Turbocharger to a Passenger Car Petrol Engine. Conference on Turbocharging and Turbo-chargers, Inst. of Mech. Engineers, London 1978

/12.7/ Gorille, I. et al. Bosch electronic fuel injectors with closed loop. control. SAE-Paper Nr. 750 368

/12.8/ Indra, F.: Entwicklung eines aufgeladenen Ottomotors fr Personenwagen mit 73,5 kW Literleistung. ATZ 80 (1978) p. 141/146

/12.9/ Hartiq, G.: Digital gesteuertes Motorzndsystem. Elektronik-Heft 9/77 Francis-Verlag Mnchen

/12.10/ Gorille I.: Digital Engine Control for European Cars. SAE-Paper no. 800 165 (Febr. 1980)

/12.11/ Geiger, I. et al. Ottomotoren mit elektronischer Regelung. Automobil-Industrie 1/79, p. 44/55

/12.12/ Dorsch, H. and Weber, I.: Abgasturbo-Aufladung fr den Porsche 924 Turbo. MTZ 40 (1979) p. 107/111

/12.13/ Gantz, J.L.: Garret-Turbolader fr schnellaufende Ver-Verbrennungsmotoren. MTZ 40 (1979) p. 81/83; see also pamphlet SPA 4988, Garret-AiResearch Industrial Division

/12.14/ Sturzenbecher, U. and Sator, H.: Kraftstoffverbrauchs-reduzierung durch Wirkungsgradverbesserung der Motoren. Autohaus 18/1978, p. 1714/1719

36

/12.15/

/12.16/

/12.17/

/12.18/

/12.19/

/12.20/

/12.21/

/12.22/

/12.23/

/12.24/

/12.25/

/12.26/

/12.27/

Kuck, H.A. et al. Emissions- und verbrauchsgnstiger Dieselmotor fr Kompaktfahrzeuge - Zusammenfassende Darstellung der erzielten Ergebnisse. BMFT-Vorhaben TV 7545: Entwicklung verbrauchs- und emissionsgnsti-ger Dieselmotoren fr Kleinwagen

Mezger, H.: Turbocharging Engines for Racing and Passenger Cars. SAE-Paper No. 780 718

Auto motor sport, Sonderdruck Heft 23 (1977)

Lange, K.H. et al. Ein aufgeladener BMW-Sechszylinder-ottomotor. MTZ 40 (1979) p. 575/578

Dommes, W. and Naumann, F.: Der aufgeladene Fnfzylin-dermotor des Audi 200, ATZ 82 (1980) p 49/58

Indra, F.: Kombinierte Aufladung an einem Personenwa-gen-Ottomotor hoher Literleistung. MTZ 40 (1979) p 581/584

Oberlnder, K. et al. The Turbocharged Five Cylinder Diesel Engine for the Mercedes-Benz 300 SO. SAE-Paper No. 780 633

Scherenberg, H.: Abgasturbo-Aufladung fr Personen-wagen-Dieselmotoren. ATZ 79 (1977) p. 479/486

Liebold, H. et al. Aus der Entwicklung des C 111 111. Automobil-Industrie 2/1979 p 29/35

Data Base for light-weight Automotive Diesel Power Plants. Report No. DOT-TSC-NHTSA-77-3, I. US Depart-ments of Transportation

Kunberger, K.: Two New Auto Diesels. Diesel and Gas Turbine Progress Worldwide, June 1979, p 64/65; see also MTZ 40 (1979) p 608

Chelline, R.: Turbocharged Automotive Diesels from VM. Diesel and Gas Turbine Progress Worldwide, Sept. 1979, p 26/27; see also MTZ 40 (1979) p. 610

BMW design their own diesel. Diesel Engineering No. 800, Spring 1979, p 16/17

37

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