Supercharging of Internal Combustion Engines: Additional Chapter 12 [Softcover reprint of the original 1st ed. 1981] 978-3-540-08544-7, 978-3-642-52196-6

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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 DOI 10.1007/978-3-642-52196-6

ISBN 978-3-642-52196-6 (eBook)

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 ｾ＠ 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 agreement 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 Combustion 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 insignificant except for some types for racing drives. Today nearly all large car manufacturing companies have turbocharged spark ignition 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.

Problems of turbocharging of automotive engines

12. 1

Petrol engines .

12.1.1

Measures dealing with combustion knock

3

12.1.2

Problems of thermal loading

9

12.1.3

Control problems

10

12.2

Diesel engines .

17

12.3

Advantages of supercharging motor car engines

18

12.4

Examples of current design

23

12.4.1

Petrol engines

23

12.4.2

Diesel engines

29

List of references

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 able

number of such engines, around 60

000

｣ｯｮＺＬｾｩ､･ｲﾭ

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 unsatisfactory, 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

ｰ｡ｲｾｩ｣ｵｬ＠

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 temperatures. The development of rotors with backward swept blades has led to an improvement in both operating range and efficiency.

,

......

\

\

\ \

\ \

\

,

\

\ I

mL lorry ( i engine

Fig. 12.1

pass. (ar ( i englne

pass. (ar s i. engme

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

Fig. 12.2

Rotor of a turbocharger of Kühnle, Kopp

&

Kausch Ltd. (KKK)

with radial flow turbine and compressor impeller with backward swept blades 2

Fig. 12.3

Corrparison of

3,0

I I

compressor rnaps of impellers 2,8

with radial (solid lines) (dashed lines). KKK type 26

I

0

and backward swept blades 2,6

Pl

I

'"e ｾ＠

I /--I I

'"

VI VI

P, 2,4

ｾ＠

c-

I

E

I I I I

E

2,2

r--;-./

2,0

I

I

I / I /

1,8

I

r+ I I

1,6

1,4

/

N Vfo7T,

/

min- '

i /

4' I

I 1,2 I I 1'°0

/

/

I

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 turbocharged 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 temperature

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 advancing 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 tendency 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 -

t 1,6

1,6 PI

PI 1, 5

1,5

1,4

1,4

',3

1,3

1,2

,

I

I

I

20

40

60

80

I

100

I

120

oe

charge air temperature

Fig. 12.4 Boost pressure for borderline knack at optimum ignition timing as a function of charge air tenperature, with air-fuel equivalence ratio and actane number as parameters. From /12.2/ Fig. 7

4

I

140

)(

1,2 I

I

6

7

E ___

8

Fig. 12.5 Effect of campression ratio and charge air tenperature on the permissible boost pressure for borderline knack. From /12.4/ Fig. 3

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 regard to power output and fuel consumption at high speeds because of the permissible advance of ignition timing as shown in Fig.12.6. °b.TDC 30

ｾ＠

t 26

'P 1.0.

22

'P'

ｾ＠

""ｾ＠

ｾ＠

ature

Effect of charge air temper-

300

ｾ＠

.,-V

on the permissible ignition

advance, the pow=r output and the

kW 220

Pe

ｾ＠

9/KW.h 340

Fig. 12.6

ｾ＠

ｾ＠

V

""ｾ＠

190

ｾ＠

280

fuel consurnption. From /12.8/ Fig. 3 60

70

80 90 100 charge air temperature

110°C

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 conditions. This is possible for example by means of a temperature-controlled 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 boostdriven air turbine (Fig. 12.7).

2

3

4

5

6

I- -

I

Fig. 12.7

Air-cooled charge air

I

cooler by Garret (method 4) .

I

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

ｾ＠

I I

The cooling system has to be judged according to effectiveness, size and weight, space availability, cost, reliability, maintenance 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 setting retarded 4 0 from the knock limit. Fuel consumption is quite sensitive to deviations, a rich mixture causing incomplete combustion and a lean one slow burning. 140 0/0

I

130

t

fe

120

Fig. 12.8

Fuel consumption at full

\

110

\

load as a function of the trapped air-fuel ratio for an ignition retarded 50 from border line knock.

100

From /12.2/ Fig. 9 0,6

0,8

/ 1.0

/

/

1.2

1,4

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 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. 18 I 100 km

16

t Fig. 12.9

12

Road fuel consurnption as

a function of vehicle velocity for naturally aspirated (dashed line)

B

and turbocharged engines of equal po\o.er tuned

for best consurnption (without

catalyzer,solid line) and for best

6

emissions (with catalyser, dashdotted line). From /12.6/ Fig. 7

I

40

60

80

v

100

120

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 controlled 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 ignition timing (/12.8, 12.9, 12.10/), which takes into ac count further parameters such as engine and air temperatures. A microcomputer calculates the optimum values of fuel rate and ignition timing according to the parameters measured by the probes, using

8

r

140 km/h

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 eontrol 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 temperatures 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 turboeharging, 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 turboehargers for spark ignition engines therefore operate with full admission up to now. HO\\ever, if spaee allows one to collect the exhaust 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 increase the flow are used.

12.1.3 Control Problems Unlike the diesel engine, the spark ignition engine is quantitycontrolled, 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 increasing 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 compressor. In the first case, the turbocharger runs at a higher

90000

80000

70000

t

60000

Ne

40000

30000

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

10000 2000

3000

initial speed and accelerates better. With carburetor engines, a throttle in front of the compressor has the following advantages: - 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; otherwise, 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 compressor 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

1

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 valve; 6

induction pipe; 4

charge air duct; 7

turbocharger (compressor); 5 air recirculation throttle plate; 8

injection lines; 10 exhaust manifold; 11 exhaust rnuffler;

14

intake manifold; 9

turbocharger (turbine);

12

exhaust pipe; 13

exhaust silencer; 15 waste gate;

16

boost pressure control line (waste gate); 17

ventilation; 18 boost

pressure controlline (air recirculation valve); 19 20 vacuum limiter; 21 22

interconnection; 23

switch /12.12/

12

fuel

bypass air valve;

air line to bypass air valve and vacuum limiter; vacuum controlline; 24

boost pressure control

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 6 Exhaust manifold; 7

air filter 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 acceleration and torque characteristic, and methods to meet these requirements 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 control

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. üf the various bypass arrangements,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 explained 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 displacement 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 unsymmetrical 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 integrated 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 lowpressure 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. 12.16

Exhaust waste gate can-

trolled by the engine back pressure P3 Fig. 12.15

Section of an exhaust

and additionally by the static pres-

waste gate with cooling bores,

sure P1st at the compressor inlet, for

by Garrett.

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 controlling 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.

16

12.17.

Maximum boost is thus generated only in the region of

Fig. 12.17

Boost pressure charac-

teristic of the Audi 200.

bo r

ｉＧ ｾ＠

0,8

f-

Kurbo - Europe

\

r--

0,4

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

"

1/ 0,2

'-..

1\,....

/

V o

2000

4000 NE -

maximum torque. Aboost pressure rising with falling engine speed can also be obtained by means of resonance pipes between

｣ｯｭｰｲ･ｳｾ＠

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 requirements, only one bypass valve is usually employed. The branches of the exhaust manifold are in this case combined be fore the turbine, 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 particular, 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 support. 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 turbocharging. 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 consumption 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 consumption is not always lower because of the compromise between fuel consumption and acceleration presently made. A small turbocharger or a narrow inlet scroll improve acceleration, but

30 l

100km

25

t 20

n.Q. engine

f 15

10

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

5 !

40

80

120 v_

I

160 km/h 200

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

Inner city 1 passenger top speed 50 km/h

1.4

88

17.58

19.9

25.60

17.7

Suburb 1 passenger top speed 50 km/h

1.1

100

17.92

17.9

15.91

15.9

City 1.4 trial course 1 passenger top speed 50 km/h

4.3

80

15.00

18.7

13.06

16.3

Black forest. rrax. gross weight

6.0

340

50.05

14.7

46.90

13.7

Table 12.11

Itr/1oo km

liter

ltr/1oo km

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/.

CVS

HC g/mile

22.08

15.71

NO g/mile

5.71

3.86 16.7

19.9

HC g/test

5.32

4.71

CO g/test

116.92

68.93

8.01

5.05

No g/test fuel cons. ltr/loo km

20

turbocharged 3.05

CO g/mile fuel cons.ltr/loo km Europa

naturally aspirated 4.46

Fig. 12.19

Road fuel consumption of a

14

diesel, a turbocharged diesel and tv.u

100 km

petrol engines in the same vehicle at

12

constant velocities. From/12.14/ Fig. 6 10

t

8 6

2

o

40

60

80

v

Table 12.111 indicates that the consumption-optimized supercharged 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

with without catalyser CO

92

HC NO

6.2 x

15.2

40.6

0.6

17 .8

2.6

7.7

122 8.6 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 supercharged version. The lagging of the boost pressure in the turbocharged 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

20 0

ｾ＠ /.",,/"

kmj h

V/ 100

Fig. 12.20 Acceleration of a natural-

50

ly aspirated and a turbocharged petrol engine of about equal po\\er men going through the gears. From/12.2/ Fig. 51

o

1/

J.

/

10

--

---

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

t

20

30

s

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

40

12.4 Examples of current design 12.4.1

Petrol engines

From among the many types of turbocharged engines presently manufactured or developed, only a few examples can be briefly described here, preferably those of European origin for which technical publications are available. Further details can be found in the publications mentioned. The company of Dr.h.c. Porsche Ltd. initially developed turbocharged 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 catalyser 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 cm 3 and 382 kW were developed /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 cm 3 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 turbocharging. 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 displacement, producing a peak output of 185 kW at 5200 rev/min for

24

Fig. 12.23

Four-cy1 inder

Saab engine, 900 turbo.

rrodel

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 40 0 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; exhaust tur-

7 exhaust manifold; 8

bine; 9 exhaust pipes; 10 waste gate; 11

controlline; 12

bypass pipe;

13

air recirculation valve;

14

air bypass

ＳＧＰ

ｉｲ］ 2,

］ＺｃｾｌＬ＠

standard conditions Po . 981 mbar T• • 293K

Fig. 12.26

Corrpressor map of the

turbocharger KKK K27 with super0,30

0,05 volume flow rate

26

inposed operating line of the

BMW engine /12.18/

map is shown in Fig. 12.26

the boost pressure at 185 kW is only

0.5 bar gauge. 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 overboosting control, all by cutting off the impulses to the fuel ihjection 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 additional 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 continuous 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 fivecylinder engine

Version Stroke/Bore

Europe

mn/mn

86.4/79.5

3

displacenent

2144

cm

nominal output

kW

at speed

rev/min

peak torque

Nm

at speed

rev/min

125

100

5200

5400

265

202

3200

3000 7: 1

conpression ratio max. boost

bar gauge

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

28

USA

0.82

0.38

SUper

Regular 186.4

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 injection 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 compressor 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 manifold 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

- -

bo ost pressure 0.90 bar boost pressure 0.55 bar

/

// 1/

100 80

'/ I

/

,Ir

500

o

400

-- r----. -,

300

V ｾ＠

t

o W 2000

/

//

V

/

Ｓｲｾ

ｾ＠

ｾ＠

...;.-

ｾ＠ ..... ｾ＠ ｾ＠

3000 N

4000 min- 1 5000

E

200 J

2000

-

) 1:

V//

Y

10 N·m

ß 1I

20

20

I

-{/ ｾ＠

-

ts

/

/

＠ｾ r- r--_

40

4th ｧ･｡ｲ

t

I I

/

30

/

/

--

t uned intake manifold - - resonator

/ ---

180

60

40

/

4000 NE -

,

6000 min- I

teristics of the turbocharged ALPINA

Fig. 12.31 Acceleration of the BMW ALPINA B7 turbo with tuned manifold (solid lines) and with

engine c.f. Fig. 12.29 /12.20/

resonator system (dashed lines)

Fig. 12.30 Output and torque charac-

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 crn 3 . The output of the naturally 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 comparison 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 unfavourable 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

11.0 kW

characteristics of the diesel engines

120

OM 617 (dashed), OM 617 A (dashdotted) and OM 617A Rekord (solid lines) /12.22/

/

100

t Pe

/ ....1./

80

60 .,/

40 20

.......

/v

ｾ＠

1--

.-,-

'-

1--

/ ,fo'

ｌ Ｇｾ＠

/ / r-.

V

-. -'-

T-' / ' .-/ V V

L-

V i--

1.00 N,m

/

o

ｾ ＧＭ

-- -- r- _

350

'"

'-

300

t

t

250 200

..........,

-.r- ..........

150 100

1000

2000

3000

I

J

4000

min- 1

NE -

the record engine in order to produce the greater output. The difference in output and in torque in particular is clearly shown in Fig. 12.33. In the production engine, the waste gate li31

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 NO x could possibly be reduced below that of the naturally 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 onwards.

3, 5

p,/p,

.".- ｾ＠

3, 0

,....-- P / P / V/ V V l

-

2,5 2,0

1,5

ｾ＠ e:; ?

1,0

ｾ＠

4

t J ; tL

900

oe

..... t)

,ti

/"""

ｾ＠ Fig. 12.34

Pressure ratios and

V

V

100

ternperatures at the turbocharger of the OM 617A Rekord engine. P2/P 1= compressor pressure ratio; P3/P4 = turbine: t 2 = charge air temperature behind cooler /12.22/

50

-:::::

o 1000

!--

-L

2000

/

V

./ ,./'V .........

V

--

700

t4 t, -

300

,

t'

t, 3000

4000

5000min-'

E

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 turbocharged 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

500

Department of Trade (DOT) or the German Federal Ministry of Research and Technology (BMFT)

in order to demonstrate the technical

feasability. In addition to the publications /12.14, 12.15/ already 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 turbocharging 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 NO x 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 because 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 NO x 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

Specifications of the VW-Golf experimental car with turbocharged diesel engine, fully enclosed, with exhaust gas recirculation. Frorn /12.14/, Fig. 11.7

Engine

Swirl combustion chamber, displacement 1.5 ltr, turbocharged diesel with wastegate

Output

51.5 kW at 5000 rev/min specific weight 2.52 kg/kW max. torque

125 Nm at 3000 rev/min

top speed

160 km/h

acceleration

o •.. 100 km/h in 13.55 s

Emissions

HC

CO

US g/mile

0.11

0.8

NO x 0.9

ECE g/test

0.6

2.33

2.95

Fuel consumption ltr/100 km

US canp. 4.7

Noise level

lSO-R 362

71dB(A)

ECE 6.4

SAE J

particles 0.25 0.6

DIN 6.3

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 cm 3 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 between 1995 and 3589 cm 3 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 cm 3 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 turbocharger /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 über den Einsatz aufgeladener Ottomotoren zum Antrieb von Personenwagen. Diss. TU München 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 turbocharging 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 Turbochargers, 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 für Personenwagen mit 73,5 kW Literleistung. ATZ 80 (1978) p. 141/146

/12.9/

Hartiq, G.: Digital gesteuertes Motorzündsystem. Elektronik-Heft 9/77 Francis-Verlag München

/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 für den Porsche 924 Turbo. MTZ 40 (1979) p. 107/111

/12.13/

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

/12.14/

Sturzenbecher, U. and Sator, H.: Kraftstoffverbrauchsreduzierung durch Wirkungsgradverbesserung der Motoren. Autohaus 18/1978, p. 1714/1719

36

/12.15/

Kuck, H.A. et al. Emissions- und verbrauchsgünstiger Dieselmotor für Kompaktfahrzeuge - Zusammenfassende Darstellung der erzielten Ergebnisse. BMFT-Vorhaben TV 7545: Entwicklung verbrauchs- und emissionsgünstiger Dieselmotoren für Kleinwagen

/12.16/

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

/12.17/

Auto motor sport, Sonderdruck Heft 23 (1977)

/12.18/

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

/12.19/

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

/12.20/

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

/12.21/

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

/12.22/

Scherenberg, H.: Abgasturbo-Aufladung für Personenwagen-Dieselmotoren. ATZ 79 (1977) p. 479/486

/12.23/

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

/12.24/

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

/12.25/

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

/12.26/

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

/12.27/

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

37