-
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
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by Springer-Verlag Berlin Heidelberg 1978.
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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were superseded by the gas turbine.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
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37
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