Abb Turbocharging Technical information

Technical information

ABB Turbocharging

Operating turbochargers – Collection of articles
by Johan Schieman published in Turbo Magazine 1992 – 1996

2 Contents | Operating turbochargers

BBC turbochargers in operation 3
How ambient conditions affect turbocharger performance 5
Turbochargers, efficiency and the diesel engine 8
Turbochargers, load cycles and overload 12
Turbocharging systems for diesel engines 14
Turbocharger compressors – the phenomenon of surging 19
Turbocharger turbines: aspects of axial and radial turbines 23
Turbocharging over the years 28
About the author 34

Contents

Operating turbochargers | Turbo Magazine 2/1992 3

BBC turbochargers in operation

Turbochargers retain their efficiency over their service
life providing they are operated properly and kept clean
by in-service washing.

The job of a diesel engine turbocharger is to supply com­pressed air to the engine. The air, heated by the compression,
passes through a cooler which reduces its temperature and
increases its density.

The air mass is compressed in the engine cylinder to a high
pressure. Fuel is injected and burnt. It is normal for a small
portion of the cylinder lubricating oil to also be burnt during
combustion. The exhaust gases that are produced pass to the
turbocharger’s turbine, which drives the compressor.

How does the turbocharger become polluted?

Particles in the air are deposited on the compressor and
diffusor blades, in time forming a layer which reduces the
compressor’s efficiency (Fig. 1). These particles have several
origins. The air taken by the compressor via filter from the
ship’s engine room is often mixed with oil vapours. Also, the
engine room ventilation system sucks in air from outside,
and this can be laden with dust, e.g. in ports where ores are
handled. Another factor is the salt in sea air.

What to do?

To counteract the drop in efficiency a certain amount of water
is injected periodically (once a week to once daily, as neces­sary) during full-load operation (Fig. 2). This measure consid­erably lengthens the intervals between turbo charger cleanings
requiring dismantling (Figs. 3 and 4).

In time, the air flowing through the cooler also causes a layer
of dirt to form on the fins and tubes, with the result that the
pressure loss across the cooler increases and the air supplied
to the engine becomes hotter. This affects how the compres­sor behaves, in that the operating curve in the compressor
characteristics moves closer to the point where surging would
occur.

Depending on the quality of the fuel (MDO, HFO or mix) burnt
in the engine and on the constituents of the lubricating oil,
residues may be formed which pass together with the exhaust
gases into the turbine.

The composition of these residues decides whether or not a
layer of dirt is deposited on the nozzle rings and tur bine
blades. Dirt layers can be moved by periodical cleaning (once
a week to once daily, as required), allowing longer intervals
between cleanings requiring dismantling of the turbocharger.

Fig. 2: Waterwashing arrangement for cleaning a compressor.

Fig. 1: Reduction of compressor efficiency due to fouling.

␩s [-]

0.9

0.8

0.7

0.6

0.5

0.3 0.4 0.5
Air flow

[m

3

/s]

clean
dirty

4 Turbo Magazine 2/1992 | Operating turbochargers

Proven methods of cleaning are waterwashing of the turbine
during part-load operation and cleaning by blasting with
ground nutshells, etc. The combustion residues can contain
hard and chemically corrosive components which can initiate
erosion and corrosion of the turbine as well as the nozzle
and cover rings. By selecting special materials or applying
coatings to relevant parts, the service life of these compo­nents can be lengthened substantially.

The importance of cleanness

The processes described above underscore the need to keep
the turbine clean to ensure trouble-free operation. Pollution
reduces the through-flow area and turbine efficiency. Conse­quently, compressor surging can also occur in the case of
two-stroke engines.

Erosion and corrosion also enlarge the clearance between the
moving blades and the cover ring, causing a drop in turbine
efficiency as well as turbocharger speed. Providing the plant
is properly looked after, i.e. the compressor and turbine are
cleaned periodically during operation and the silencer filter is
kept clean, the efficiency will diminish only very slowly. A nor­mal figure today for the interval between two cleanings requir­ing dismantling is 8,000 to 16,000 hours, depending on the
operating conditions and unit size.

A turbocharger which has been dismantled, cleaned and had
its worn parts replaced, will have an efficiency “as good as
new”. Some 80 service stations are available worldwide to
help owners maintain high efficiencies for their turbochargers.

Turbochargers pollution and engine operation

The consequences of a drop in turbo charger efficiency for the
engine are:
– Higher exhaust gas temperatures
– Lower charging pressure
– Smaller air mass flow

The usual complaint by ship’s engineers concerns high exhaust
gas temperatures, although these are really a result of lower
pressure and smaller air mass flow. As has been said, turbo ­charger efficiency decreases only very slowly with time. Despite
this, turbocharger suppliers often hear complaints of high
exhaust gas temperatures, wrongly reported as being caused
by the turbocharger.

Often, high exhaust gas temperatures are caused by changes
in the properties of the engine’s fuel injection system over
its lifetime. With fuel injection, pressure build-up is not so fast
and combustion is slower due to wear and leaks in the fuel
pumps.

ABB Turbo Systems has investigated these factors in some
detail. It was seen that the engine itself can cause higher
exhaust gas temperatures even when turbo charger efficiency
is good. At part load, this influence is far less pronounced.
The first reaction is to run the machine with reduced output.

An often-heard request in connection with complaints of
high exhaust gas temperatures is for the turbocharger specifi­cations to be adapted to allow more air to flow. Before any
decision is taken to modify the turbocharger, it pays to first
analyze why the temperature is so high. Under certain circum­stances this can save a considerable amount of money.

The described causes of higher exhaust gas temperatures
usually appear in combinations of one or the other. In
time, the turbocharger supplies less pressure and air, and
the engine combustion starts later and lasts longer.

For older plants, it is often better to limit the output to the
maximum that is required and adapt the turbocharger
specifications to the new situation or replace the old unit
with a new, modern one with higher efficiency.

This will allow a considerable improvement in the “perfor­mance” of an installation which otherwise would not be
capable of “new” operating values.

%

Fig. 3: Example: Reduction in charging pressure with time.

Fig. 4: Compressor cleaning in service. Example: cleaning
every second day.

%

100

90

Charging pressure

80

70

Compressor
not washed

4000 8000 12000 16000

Heavy
fouling

Operating conditions

Running hours

Normal
fouling

Light
fouling

100

99

98

Charging pressure

97

96

2 4 6 8 10 12 Days

50 100 150 200 250 300

Running hours

Operating turbochargers | Turbo Magazine 1/1993 5

In an earlier article in Turbo Magazine II/1992 we explained
how the diesel engine and turbo charger interact, that the
compressor supplies air to the engine through an air cooler,
and that the engine feeds the turbine with exhaust gases to
drive the compressor. Providing the turbocharger is kept
adequately clean during operation, its performance will stay at
a high level for a long period of time. However, another factor
– the ambient conditions – also plays a role and could cause
an unexpected change in performance.

How do ambient conditions affect performance?

The properties of the air taken in by the compressor are:
– barometric pressure
– air intake temperature
– humidity

While the first and second of these have an immediate effect
on turbocharger performance, there are also some secondary
influences of an external nature which could cause the turbo ­charger performance to change. These can be, for example:
– the temperature of the cooling water used in the air cooler,

or

– increased torque requirements at lower engine speeds, e.g.

due to fouling of the ship’s hull.

The conditions that count

Barometric pressure

The barometric pressure depends on the height above sea
level, and decreases as this height increases. However, even
at sea level considerable variation is possible. (The lowest
pressure known – 0.915 bar – was measured in January 1993
in the North Sea; the highest values lie at about 1.04 bar). At
any given engine load the pressure ratio across the turbine
increases as the barometric pressure goes down. The result is
a higher turbocharger speed (Fig. 1).

Many turbochargers are in operation on ship’s propulsion
engines. For a given engine output the turbocharger speed
may differ slightly because it depends on the barometric
pressure. The pressure ratio of the compressor increases
with the turbocharger speed. The resultant charging pressure,
having built up at the lower barometric pressure, remains
slightly lower than the charging pressure at a higher baro­metric pressure. Above sea level the barometric pressure is
lower and for a given engine load the charger speed will
increase accordingly (Fig. 2).

How ambient conditions affect turbocharger performance

In addition to proper operation and washing, ambient conditions
also play a role in turbocharger efficiency.

Fig. 1

Fig. 2

[%]

110

105

100

Relative turbocharger speed

95

90

“Hurricane”

ISO

“Normal”

Engine load constant

Intake temperature constant

1.00.9

Barometric pressure p

“Fine weather”

o

Height above sea level

3000

[m]

2000

1000

0

[%]

Relative turbocharger speed

n

TL

[bar]

120

110

100

Constant engine output

0.8 0.9

Barometric pressure p

o

ISO

[bar]

6 Turbo Magazine 1/1993 | Operating turbochargers

Turbochargers on diesel locomotives in mountainous regions
have to take account of these conditions. When the maximum
permissible operational speed has been reached, the engine
output has to be reduced to prevent the charger from over­speeding. It may also be necessary to use other measures
(e.g. an exhaust gas waste gate) to prevent overspeed.
Usually, a turbocharger will be specified for International
Standards Organization conditions, i.e. 1 bar ambient pres­sure. If deviations are considerable, the engine may have to
be derated.

Air intake temperature

The energy needed to compress the air is directly proportional
to the intake temperature. Therefore, at a given turbocharger
speed, the pressure ratio of the compressor decreases with
increasing intake temperature, and vice versa.

Fig. 3 shows that the influence is considerable. If we sail with
a ship from a moderate to a tropical region (from 25 to 45°C
intake temperature) the absolute charging pressure decreases
by 6 % (from 1.0 to 0.94 in Fig. 3).

The consequences of this for a given engine load are:
– less air throughput
– lower air-fuel ratio
– higher exhaust gas temperatures

If we now sail to Arctic waters, where the intake temperature
may be –10 °C, the absolute charging pressure increases by
10 %. We now have to be careful that the maximum firing
pressure in the engine is not exceeded.

The stability of the compressor may be impaired as these
conditions move the working point on the compressor map
nearer to the surge line (Fig. 4).

Fishing vessels operating in Arctic waters are confronted with
this problem now and again. A good remedy is to prevent out­side air from entering the engine room near the charger air
intake.

Other applications in tropical climates may be taken care of
by carefully specify ing the turbocharger. In most cases the
specification is based on ISO air intake conditions of 1 bar
and 25 °C. Depending on the location and application, ISO­recommended derating may be applied to prevent thermal
and/or mechanical overload of the engine and turbocharger.

Air humidity

The performance of the turbocharger is not affected by the
humidity. However, compressed humid air carries a consider­able amount of water as it passes through the air cooler and
this may condense into the air receiver. This may be overcome
by either frequent or continuous venting, though with a subse­quent loss of some air pressure.

Fig. 3

p

Fig. 4

[-]

2

p

1

Relative pressure ratio

1.5

1.0

0.5

0.0

0– 50 50

Intake temperature t

ISO

1

[°C]

5.0

4.5

4.0

3.5

3.0

2.5

2.0

1.5

1.0

Air intake temperature

t1 = –10 °C

t

= 25 °C

1

t

= 45 °C

1

340

330

320

300

280

260

240

220

210

2.0 3.0 4.0

350

360

370

380

390

400

410

420

0,82

430

440

0,65

n

300

-1

s

450

460

V

␩

300

470

*

sV

[m3/s]

Operating turbochargers | Turbo Magazine 1/1993 7

The impact of secondary influences

Temperatures of cooling water used by air cooler

In the case of turbocharged 4-stroke engines the temperature
of the charge air has a significant influence on the air throug­put. If, for example under ISO conditions, the charge air
temperature is 50°C and, due to a high cooling water inlet
temperature, this temperature increases by 10°C to 60 °C,
the air throughput will decrease by 3% (corresponding to
absolute temperatures of 323 K and 333 K, respectively). This
moves the working point on the compressor map nearer
to the surge line (Fig. 5). Occasionally, this phenomenon is
found to be the reason for surging.

Torque requirements

For 4-stroke engines for marine propulsion the condition may
occur whereby a high torque is required at relatively low
engine speed. As the air throughput is mainly determined by
the engine speed, the operating line on the compressor map
moves to the surge line (Fig. 6). Such behavior could be
caused by fouling of the ship’s hull. In severe cases, this may
even lead to surging.

It goes without saying that these possibilities are all consid­ered in the turbocharger specification. However, the influence
of high compressor efficiency near the surge line and the con­flicting need to have a wide safety margin separating the
working point from the surge line call for a compromise.

Modern compressors with backswept vanes, e.g. those used
in BBC

®

VTR . .4 turbochargers, feature excellent stability in
terms of surging. As a result, the compressor’s high efficiency
is available over the whole load range of the engine.

p

Fig. 5

Fig. 6

2

p

1

p

2

p

1

5.0

4.5

4.0

3.5

3.0

2.5

2.0

1.5

1.0

Charge air

temperature 60 °C

3%

440

420

400

380

360

340

320

300

280

260

240

1.0 2.0 3.0

n

300

-1

s

540

520

500

480

460

0,65

␩

Charge air

temperature 50 °C

V

555

*

300

sV

[m3/s]

5.0

4.5

4.0

3.5

3.0
Lower engine speed

Higher torque

2.5

2.0

1.5

1.0

460

440

420

400

380

360

0,8

260

240

280

300

320

340

1.0 2.0 3.0

480

500

n

300

-1

s

520

0,65

555

540

Higher
speed

Lower
torque

␩

*

sV

V

[m3/s]

300

8 Turbo Magazine 2/1993 | Operating turbochargers

Turbochargers, efficiency and the diesel engine

The turbocharger has a decisive influence on the diesel engine’s
performance, while its efficiency plays an important role in how
the diesel engine and turbocharger interact.

To understand these relationships better, we have to distinguish
between the turbocharger efficiency and the turbocharging
system efficiency, and establish in what way they affect the
performance of the turbocharged diesel engines.

The turbocharged diesel engine

Fig. 1 depicts the various components of a diesel engine
installation. The air flowing to the engine first passes through
the silencer filter, which causes a small pressure drop. The air
is then compressed from pressure p

1

to p21.

The velocity of the air leaving the compressor is reduced in a
so-called diffuser before it enters the air cooler. This process
causes some loss of pressure. Afterwards, the air passes
through the air cooler. The air sustains another pressure drop.

A pressure ratio p

21/p1

is generated by the compressor. The

engine senses only p

R/po

. The reduction in the pressure ratio

from p

21/p1

to pR/pois due to the air section of the turbo -

charging system.

Similar conditions exist on the exhaust side. The connection
between the exhaust valve and the turbine inlet must be
shaped to prevent pressure losses. The backpressure after
the turbine, p

4

, has to be kept as low as possible in order to

keep the pressure ratio over the turbine, p

3/p4

, high.

This situation shows us that:
– The turbocharged engine senses the turbocharging accord-

ing to the pressure ratio p

R/po

, pR/p3and p3/po.

– The turbocharger has to cope with the pressure ratios

p

21/p1

and p3/p4.

poBarometric pressure
p

1

Pressure, compressor inlet

T

1

Temperature, compressor inlet

p

21

Pressure, compressor outlet

p

22

Pressure, cooler inlet

p

R

Pressure, air receiver

T

R

Temperature, air receiver

p

3

Exhaust gas pressure

T

3

Exhaust gas temperature

p

4

Exhaust gas pressure after turbine

T

4

Exhaust gas temperature after turbine
A Air intake, silencer filter
V Compressor of turbocharger
B-C Compressor outlet duct
C-D Air cooler
D-E Air receiver, connection to inlet valve
F Fuel injection nozzle
G-H Exhaust gas duct to turbine
T Turbocharger turbine
I Exhaust gas manifold after turbine
K* Exhaust gas boiler
L Exhaust silencer, spark arrester

*for bigger engines

Fig. 1

p

0

L

K

p

4T4

I

H

p

3T3

F

GE

p1T

1

p

0

A

p

B

21

p

C

22

D

p

RTR

Operating turbochargers | Turbo Magazine 2/1993 9

We can express the efficiency required from the turbocharger as

and the system efficiency as

The system efficiency is lower than the turbocharger efficiency.

The symbols represent:

␩

TC

Overall efficiency turbocharger

T

1

Compressor intake temperature

R Gas constant

␬

A

Ratio of specific heats of air

π

c

Compressor pressure ratio

m

c

Compressor air flow

T

3

Temperature before turbine

␬

g

Ratio of specific heats for exhaust gas

π

T

Pressure ratio over turbine

m

T

Turbine gas flow

p

R

Charge air pressure

p

o

Barometric pressure

p

3

Exhaust gas pressure before turbine

␩

system

Turbocharging system efficiency

Turbocharger efficiency and engine performance

Fig. 2 shows a cross-section of the VTR turbocharger with its
main components:
– Compressor
– Turbine
– Bearings

T

1

· R ·

␬

A

·(π

c

␬A-1

- 1

)

· m

c

␬A-1

␬

A

T3· R ·

␬

g

·(1 -

1

)

· m

T

␬g-1

π

T

␬g-1

␬

g

␩TC= (1)

T1· R ·

␬

A

·

(

(

p

R

)

␬A-1

- 1

)

· m

c

␬A-1 p

o

␬

A

T3· R ·

␬

g

·(1 -

1

)

· m

T

␬g-1

(

p

3

)

␬g-1

p

o

␬

g

␩

system

= (2)

These parts have an efficiency of their own:

␩

C

= Compressor efficiency

␩

T

= Turbine efficiency

␩

mec

= Mechanical efficiency (bearing friction)

To prevent exhaust gases from entering the bearing chamber
on the turbine side and to provide some cooling for the tur­bine disc, sealing air is taken from the compressed air flow.
Hence:

␩

vol

= Volumetric efficiency (sealing air)

There are cooled and uncooled gas inlet casings. The cooled
gas inlet casing reduces the exhaust gas temperature some­what before the gas reaches the nozzle ring of the turbine.
Therefore:
kW = Cooling effect (uncooled casing, kW = 1)

Without going into the details of ␩

C

and ␩T, it can be stated

that:

␩

TC

= ␩C· ␩T· ␩

mec

· ␩

vol

· kW (3)

The engine experiences ␩

TC

(1) = ␩

TC

(3)

Fig. 2

Compressor

Turbine

Bearing

Bearing

Sealing air

10 Turbo Magazine 2/1993 | Operating turbochargers

Fig. 3 compares a compressor map of the VTR321 with one
for the VTR 354, given approximately the same air flow
and pressure ratio. Fig. 4 compares the turbine maps. Fig. 5
shows the difference in performance of a 4-stroke engine
when a better charger is used.

The indicator diagram of a 4-stroke engine is shown in Fig. 6.
From formula (1) we can see that if ␩

TC

has a higher value and
all other values remain more or less constant, the turbine
pressure ratio will have a lower value. This means that the
pressure ratio over the engine, p

R/p3

, increases. It allows a

larger turbine area.

p

Fig. 3

Fig. 4 Fig. 5

150

160

[gr/PSh]

[bar]

p

me

12 14 16 18

Fuel consumption (be)

300

400

500

[ºC]

Exhaust gas temperature
before turbine (t

vT

)

Temperature of exhaust
gas leaving cylinder (t

nZ

)

2

1

3

[-]

Compressor pressure ratio (␲v)

5

6

[kg/PSh]

Specific air flow (l

e

)

VTR321

VTR354

Comparison of engine test results obtained with the VTR 321 and VTR 354

tot

2

tot

p

1

5.0

4.5

4.0

3.5

3.0

2.5

2.0

1.5

1.0

300

V

V

=

T

=

␩

240

1

260

300

T

280

1

1

300

320

n

340

300

-1

s

360

170

300

␩

300

220

200

1.0 2.0 3.0 4.0 5.0

380

0.82

390

395

␩

sV

V

VTR 354

[m3/s]

300

p

tot

2

tot

p

1

V

V

=

288

1

3.5

␩

␩

=

288

288.15
T

288.15
T

␩

= 23000 [min.1]

288

1

1

22000

21000

20000

3.0

19000

2.5

2.0

1.5

12000

14000

16000

15000

17000

18000

0.78

= 0.65

␩

VTR 321

1.0

2.0 3.0 4.0

V

288

[m3/s]

␩sT [-]

0.85

0.80

0.75

0.70

0.65

1.5 2.0 2.5 3.0 3.5

␲

T

VTR 354

VTR 321

Operating turbochargers | Turbo Magazine 2/1993 11

More positive work can be achieved during the gas exchange
period. This has, as its reward, a reduction in fuel consump­tion. It does of course assume, though, that the timing of the
inlet and exhaust valves has been correctly adjusted.

Turbocharging system efficiency

A typical example is shown in Fig. 7. By comparing the two
arrangements in terms of the flow conditions for air and gas, we
quickly see that the configuration with the fewer bends in the
ducts provides most benefits. A diffuser between the compres­sor outlet and the cooler inlet considerably reduces the pres­sure loss that would be caused by two 90° bends in the duct.

A straight pipe allowing a high-velocity flow of gas to the
turbine avoids the pressure loss due to a 90° bend. Fig. 8
shows the increasing differences in the turbocharger and

system efficiencies as the pressure losses increase on the
air-side.

Obviously, the ducts for the air flow must be of very good
aerodynamic design. And it is essential to keep the air inlet
filter and the air cooler of an installation clean so as to keep
the flow resistance low. High pressure losses can cause the
compressor to surge.

An increasing backpressure after the turbine, for instance due
to fouling of the exhaust boiler, will cause the charging pres­sure to drop, Fig. 9. This can cause the compressor to surge.

Turbocharging systems

A comparison of different turbocharging systems will be given
in a later issue of Turbo Magazine.

Fig. 6

p

Fig. 7

3.0

␲c [-]

␩

TC

␩

sys

3.1

3.2

0.65

0.64

0.63

0.62

0.61

0.60

0.59

“Normal” valves

0.05 0.10 0.15 bar

␲

compressor needed

to maintain 3 bar

␩

system

␩

TC

Charging pressure
PR = 3.0 bar

⌬p losses

Fig. 8

2.7

2.8

2.9

3.0

3.1

␲c [-]

“Normal” valves

0.05 0.10 bar

Backpressure

Fig. 9

p

R

Exh. closes

+

p

3

p

3

p

R

p

3

Inlet opens

TDC BDC V

+

Exh. opens

Inlet closes

+

Actual situation

Turbine
inlet

Situation could be

Compressor
outlet

Diffuser

Cooler

MPC

Pressure losses
down by at least 50%

Cooler

MPC

12 Turbo Magazine 1/1994 | Operating turbochargers

Turbochargers, load cycles and overload

The way fishing vessels are operated determines the duty cycle
of their diesel engine and turbocharger. Some breakdowns can
be avoided by understanding how these machines react to the
ships’ working schedules.

The working schedule for fishing vessels operating in the
North Sea usually begins on Monday, when they leave for the
fishing grounds, and ends the following Friday, when they
return to port at full speed. The engine power output and
speed therefore have to satisfy two requirements: one is the
need for the ship to arrive as quickly as possible at the fishing
grounds, and later to return to port, with the ship “steaming”
freely; the other involves the actual fishing operations, when
the nets are out and being dragged through the water, possi­bly over the sea bottom.

Often, government regulations restrict the length of time a
fishing vessel may stay at sea, i.e. it has a limited number of
“seadays”. These seadays include the time needed to reach
the fishing grounds. It goes without saying that the ship will
be run at its maximum possible speed in order to reach the
fishing grounds and return to harbour as fast as it can. During
the fishing, when the nets are out, the engine is run at the
maximum possible torque, and every time the nets are pulled
in the load and speed are reduced. This is a procedure that
repeats itself at least every hour.

How the machines respond

How do the engine and turbocharger react to this type of
operation? In Fig. 1, the turbocharger speed is plotted against
the engine output. The 100 percent engine output is the limit
which should never be exceeded. This is ensured by blocking
the fuel pump rack on the engine. Usually, the blocking shim
is secured by a lead seal.

The turbocharger speed at full engine load is determined dur­ing the acceptance testing of the engine. Its operational relia­bility for applications on board fishing boats is also checked.
As shown above, this application involves many hours of full­speed running to and from the fishing grounds as well as large
numbers of wide variations in engine load and speed, with the
turbocharger speed varying accordingly. An example of the
time-related load is shown in Fig. 2, the number of load cycles
over a period of time in Fig. 3. The values have been compiled
from measurements on board fishing vessels.

Load cycles show the variation in engine load from low to high
and high to low. This can cover “idling/full load /idling”,
“half load/full load/ half load”, “part load/higher part load /part
load” or “idling/part load/ idling” for the engine. The turbo ­charger speed varies as shown in Fig. 3.

Given the very heavy duty expected of the engine and its
turbocharger, it is clear that if the load level is increased
beyond the output for which the engine has been built the
lifetime of some parts of the turbocharger may be reduced,
resulting in a breakdown.

The large number of load cycles, especially at an increased
load level, may cause a reduction in the lifetime of turbocharger
parts. Turbocharger overspeed is a possibility when the ship
is “steaming” freely and the output increases rapidly with the
engine speed.

Fig. 1

[%]

100

80

60

40

Turbocharger speed

20

0

10 20 30 40 50 60 70 80 90 100 110[%]

Engine output

Operating turbochargers | Turbo Magazine 1/1994 13

“Readjusting” for disaster

ABB Turbo Systems has been confronted several times with
serious turbocharger breakdowns caused by “readjustment”
of the engine fuel pump racks with the objective of running
the engine at a higher output than the one it was intended for.
This is a very dangerous thing to do, and the consequences
can be disastrous. Although a lot of effort has been put into
containment of the turbocharger (by designing it so that if
a breakdown does occur, any broken parts remain within
the casings), there have been situations where broken parts
did fly out of the charger. This was brought about by the
extreme overload conditions in which the installation had
been operated for a prolonged period of time.

Consult the engine builder first

It should be clear from this that the engine should never be
“readjusted” for a higher output without first consulting
the engine builder. The engine builder will make sure that the
turbocharger is adapted to match the new engine output
requirements. In this way, severe breakdowns can be largely
prevented and the risk to personnel minimized.

Time-related engine load

Time-related load

Fig. 2: Fishing vessel operation. Measured engine load and related turbocharger speed.

Fig. 3: Fishing vessel operation. Measured load cycles. 100% corresponds to 53,000 load cycles in 10,000 hours.

[%]

80

60

40

Engine load

20

0

10 20 30 40 50 60 70 80 90

Operating time

[%]

[%]

80

60

40

Turbocharger speed

20

0

10 20 30 40 50 60 70 80 90

Operating time

[%]

[%]

100

80

60

Engine load

40

20

0

10 20 30 40 50 60 70 9080

Load cycles

[%]

[%]

100

80

60

40

Turbocharger speed

20

0

10 20 30 40 50 60 70 9080

Load cycles

[%]

14 Turbo Magazine 2/1994 | Operating turbochargers

Turbocharging systems for diesel engines

In another article for our readers with a technical leaning, Johan
Schieman describes how the turbocharging system influences
diesel engine performance and discusses the pros and cons of
different turbocharging systems.

Turbocharging systems

ABB turbochargers are in use on diesel engines rated at
500 kW and more. In the following, we shall consider only
those engines on which ABB’s turbochargers can be used.
Fig. 1 shows a turbocharged diesel engine. The exhaust
gases drive the turbocharger’s turbine, which in turn drives
the air compressor, both of which are mounted on the
same shaft. The compressed air is passed through an air
cooler to the engine’s air receiver.

The name given to the turbocharging system comes from the
arrangement of the exhaust ducts between the engine cylin­ders and the turbocharger gas inlets. Fig. 2 shows the different
arrangements. These are in use on 4-stroke diesel engines.
Pulse and constant pressure systems are also possible for
2-stroke engines. Pulse systems use narrow pipes with small
volumes between the cylinders and the turbine gas inlets.

The available turbocharger efficiency plays a key role in the
choice of turbo charging system. As part of the search for
better efficiency for the turbocharging system, research was
also carried out into so-called two-stage turbocharging.

Through the development of turbo chargers for higher pres­sure ratios and higher efficiencies, it has been possible over
the years to meet practically all the requirements. With today’s
compressor designs, pressure ratios of approximately

5 are

possible. It can be seen from Fig. 3 how the turbocharger

effi-

ciencies have developed over the years.

Pulse systems versus constant pressure systems

The following may help the interested reader to understand
the gas exchange process of the 4-stroke engine better. Fig. 4
shows the valve timing of the inlet and exhaust with the over­lap for scavenging the top dead center cylinder space with
fresh clean air. A pressure difference is necessary to allow the
scavenging air to flow from the air receiver through the top
dead center cylinder space into the exhaust duct. This pres­sure difference is determined by the turbocharging system
and the turbo charger efficiency. When the turbocharger effi­ciency is low, the exhaust gas pressure is only slightly lower
than the charge air pressure.

Turbine

Engine cylinder

Piston

Compressor

Aircooler

Charge air
receiver

Exhaust gas

Fig.1: Cutaway drawing of a turbocharged diesel engine.

Fig. 2: Arrangements of exhaust ducts.

1 2 3 4

2-Pulse

123456

3-Pulse

1 2 3 45678

Pulse converter

1 2 3 45678

Multi-pulse

1 2 3 45678

Modular-pulse converter

1 2 3 45678

Constant pressure

Operating turbochargers | Turbo Magazine 2/1994 15

From Fig. 5 it is seen that if the turbo charger efficiency is low
the pulse system offers advantages by allowing an optimal
clean cylinder charge of fresh air. Due to the large pressure
variations in the exhaust pipe after the cylinder during the
exhaust period, the difference between the charge air pres­sure and the exhaust pipe pressure is also quite large during
the inlet-exhaust overlap (Fig. 4). This allows ample scaveng­ing. The constant pressure system permits hardly any pres­sure difference over the cylinder, and scavenging of the top
dead center cylinder space is very limited. As turbocharger
efficiencies have become higher, this disadvantage has
become less important.

The engine firing order is chosen so as to ensure that stresses
in the crankshaft as a result of torsional vibration are as low
as possible. Using a pulse system and the given firing order,
combinations of cylinder exhausts on one pipe to one turbo ­charger gas inlet must be chosen in such a way that the
exhaust pulse of one cylinder cannot disturb the scavenging
period for another cylinder on the same pipe.

Three-cylinder combinations with a firing interval of 240°
crank angle (CA) are in use on 6-, 9-, 12- and 18-cylinder
engines, while for 8- and 16-cylinder engines the firing inter­vals used are 360°CA or 270 /450 ° CA. Pulse systems have
also been built for 5- and 7-cylinder engines.

The constant pressure system has the same large exhaust
gas receiver irrespective of how many cylinders the engine
has. The same applies to single pipe exhaust systems, where
the exhaust receiver volume is reduced in comparison with
constant pressure systems.

The efficiency with which the turbocharger turbine can oper­ate depends on how it is supplied with the exhaust gas. A
continuous gas mass flow (full flow) to the turbine supports
optimal utilization of the energy in the exhaust gas. 3-pulse
cylinder combinations, constant pressure and single pipe sys­tems assure this mode of gas supply, whereas 2-pulse cylin­der combinations provide an intermittent supply of gas to the
turbine.

In Fig. 6 the exhaust gas pressure, charging pressure and
cylinder pressure are plotted over °CA for the 3-cylinder com­bination. Fig. 7 shows the same quantities for the constant
pressure system. Fig. 8 illustrates, for the 2-cylinder combina­tion, the periods when there is no pressure in the exhaust
gas duct, i.e. no gas supply to the respective gas inlet of the
turbine. The turbine, driven by gas from the other gas inlets,
drags the blades through stagnant gas, which obviously
reduces the efficiency of the system.

[bar] [bar] [bar] [bar]

BDC TDC BDCBDC TDC BDC

Pulse

Higher TC efficiency

Constant pressure

Low TC efficiency

BDC TDC BDC

p

exh.

p

R

p

exh.

p

R

Scavenging period

p

R

p

exh.

BDC TDC BDC

Scavenging period

Pulse

Low TC efficiency

Constant pressure

Low TC efficiency

p

exh.

p

R

Fig. 5: Influence of turbocharger efficiency on the scavenging process for pulse and constant pressure systems.

␩TC [%]

Fig. 3: Development of the turbocharger efficiencies over the years.

Fig. 4: Valve timing of inlet and exhaust.

70

65

60

55

VTR..4E

VTR..4

VTR..1

VTR..0

[%]

Valves are open

100

50

Exhaust

Inlet

Scavenging
overlap

50

1970 1975 1980 1985 1990

0

TDC

0

BDC

180

TDC

360

BDC

540

TDC

°CA

16 Turbo Magazine 2/1994 | Operating turbochargers

Solutions

To overcome the problems mentioned, various systems have
been introduced to achieve a full flow gas supply to the turbine
for those cylinder numbers that cannot be divided by 3, e.g.:
– Pulse converters
– Multi-pulse converters
– Single pipe exhaust gas ducts with a relatively small gas

flow area

– Constant pressure exhaust gas receivers

In the full-load operation range of the engines, these solutions
have brought improved performance compared with the two­pulse arrangements. However, in some applications the part­load performance deteriorates in comparison with “straight”
pulse operation. This applies particularly to ships’ propulsion
systems and the traction sector (diesel electric locomotives),
where the engine output is proportional to the engine speed
to the power of 3 (Pe = n

3

).

How the system functions

We shall now use the example of an 8-cylinder engine with
pulse converters to explain how such a system achieves full
flow exhaust gas to the turbine. Fig. 9 shows the exhaust
strokes of all the cylinders over °CA of cylinder 1, the crank
rows with firing order, and the exhaust duct arrangement for
four gas inlets. By combining the ducts 1 – 8 and 2 – 7 with a
pulse converter on one turbine gas inlet, as shown in Fig. 10,
exhaust gas is supplied with an interval of 180 °CA on one gas
inlet instead of with an interval of 360 °CA on two gas inlets.

Full flow is achieved. However, it is seen from the diagram
that the high pressure pulses of 2 – 7 are exactly there where
1 and 8 have their scavenging period. Fortunately, it takes
some time for a disturbing pulse from no. 2 cylinder to reach
the turbine and for its reflection to reach the scavenging
cylinder (no. 1), with the result that the disturbance arrives at
the end of the scavenging period. By correctly dimensioning
the pulse converter areas, the intensity of the reflected pulse
can be reduced to a value lower than the charging pressure.
In this way, the scavenging process is hardly impaired when
pulse converters are used.

The “straight” pulse system uses narrow pipes having a small
volume. This enhances the build-up of intensive pulses and
turbocharger response to demand for higher engine load. By
combining two ducts via a pulse converter, the duct volume
connected to one gas inlet increases and the build-up of the
exhaust pulse becomes less intensive. The maximum pulse
pressures are lower than with a “straight” pulse system. Thus,
the turbo charger responds more slowly to an increase in
engine load. The multi-pulse converter has an even larger pipe
volume and is even slower to react. The constant pressure
system, with no pulses at all, is slowest here. Single-pipe sys­tems are comparable with the latter two systems.

For ship’s propulsion systems with a fixed pitch propeller, the
pulse system offers considerable advantages in terms of
engine part-load performance over systems with large exhaust
duct volumes. Fig. 11 shows the reason for the favourable
part-load performance. Looking at the three-pulse system, it
can be assumed that there is a constant pressure region with
superposed pulses. At full load the pulses contain only a small
part of the energy supplied to the turbine. At part load, how­ever, the pulses contain a relatively large proportion of the
energy, this proportion increasing as the loads become lower.
Similar considerations are valid for two-pulse arrangements.
At part load, the pulse system can therefore supply more air
at a higher charging pressure than systems with larger exhaust
duct volumes. Fig. 12 shows the relative differences in achiev­able part-load charging pressures.

The question can be asked, given the pulse systems’ clear
advantages in terms of performance, why many engines have
a single pipe or a constant pressure turbocharging system?
Let us consider the pros and cons of the latter systems.

p [bar]

Fig. 6: Pressures of a 3-pulse system plotted over degrees crank
angle (° CA).

Fig. 7: Pressures of constant pressure system plotted over °CA.

Fig. 8: Pressures of 2-pulse system plotted over °CA.

p

cyl1

p

exh.

p

R

240 240 240TDC cyl 1

180

Scavenging period

360

540 ° CA

p [bar]

p

cyl1

p

R

p

exh.

180 360 540 °CA

Scavenging period

TDC

p [bar]

p

cyl1

p

exh.

p

R

180 360 540 °CA

Scavenging period

TDC1360 °

Operating turbochargers | Turbo Magazine 2/1994 17

Advantages:

Simplicity
Suitability for any number of cylinders
Easy standardization of many identical parts
Constant engine speed application, few problems
One turbine gas inlet, possibly next smallest charger size

Disadvantages:

Large expansion compensators
Slow load take-up response
Low part-load charging pressure for propeller law/traction

The low part-load charging pressure leads to a low air/ fuel
ratio for the combustion and high thermal loading of the parts
of the combustion chamber.

Several methods exist with which the low part-load charging
pressure can be remedied.
– The charger is specified for an adequate part-load charging

pressure. An air waste-gate for the full-load range can be
used to limit the otherwise too high charging pressure.

– The charger is specified as before, but an exhaust gas

waste-gate is used.

– A bypass duct guides air from before the air cooler into the

turbine inlet to increase the charging pressure at part load,
allowing the turbocharger compressor to be specified for
higher efficiency at full load.

At first sight the constant pressure/ single pipe systems appear
to be more favourable in terms of ease of manufacture. How­ever, the auxiliary measures which have to be taken to meet
requirements such as fast take-up of load, satisfactory part­load performance for fixed pitch propeller drives, etc., add to
the costs.

Fig. 13 compares the performance of constant pressure and
3-pulse systems. Note the difference in ␭

v

, a measure of the

air/fuel ratio for combustion.

Trends and preferences for some applications

We can distinguish between three main fields of application
for ABB turbo chargers:
– Ships’ propulsion systems
– Generator drives (power stations, ships’ auxiliary engines)
– Locomotive drives (traction)

Fig. 11: Pulse system. Good part load performance because
of relatively high pulse energy at low load.

90

1, 8

2, 7

3, 6 4, 5

180 270 360 450 540 630 ° CA

TDC 4S
TDC 5F

TDC 2S
TDC 7F

TDC 2F
TDC 7S

TDC 3S
TDC 6F

TDC 1S
TDC 8F

TDC 1F
TDC 8S

TDC 3F
TDC 6S

TDC 4F
TDC 5S

Exh. 5

Exh. 7 Exh. 2

Exh. 3 Exh. 6

Exh. 8Exh. 1

Exh. 4

Firing order: 1-3-2-4-8-6-7-5-1
TDC.F: firing
TDC.S: scavenging

12345678

Fig. 9: Exhaust gas pressure sequence of an 8-cylinder engine
with 2-pulse system. 4 exhaust ducts, 4 turbine gas inlets.

TDC 1S

Fig. 10: Pulse converter arrangement for two 2-pulse exhaust ducts.

Fig. 12:

Relative difference in achievable part load charging pressure pR.

Exh. 2 Exh. 7

Exh. 8 Exh. 1

TDC 8S

TDC 2S 360

1

278

I

II

III

Pulse converter

Pipe area II < area I, pipe area III = 2 x area II

TDC 7S

°CA

p [bar] p [bar]

Exh. pulse

50 100 % BDC TDC

Load

Full load

p

Average exh. pressure

charging pressure

R

pR [%]

100

50

0

nM %

100

50

Load

100 %

Pulse system

Pulse converter

Multi pulse converter

Constant pressure

Engine speed

18 Turbo Magazine 2/1994 | Operating turbochargers

Occasionally, there are also applications where full torque is
required at variable speed.

More recently, we have seen an increase in applications with
pulse systems. There are even signs of 2-cylinder combina­tions staging a comeback – a trend also applying to the highly
turbocharged engines. This is due to the fact that although
the turbocharging system itself is not very efficient, the effi­ciency of the turbo chargers now available compensates in
part for the low system value. The negative influence on fuel
consumption is partly compensated for by modern fuel injec­tion apparatus, assuring the best possible specific fuel con­sumption. Two-pulse systems are found on modern 4-cylinder
auxiliary engines for ships.

As regards ships’ propulsion systems and traction applica­tions, three-pulse arrangements are frequently used on 6-, 9­and 12-cylinder engines. This is also the case for generator
drives with fast load take-up requirements.

Single pipe or constant pressure turbo charging is frequently
used on big engines for power generation, especially when
these are used for base-load operation. This type of turbo ­charging may also be found on locomotives, e.g. with
16-cylinder engines and possibly equipped with ancillary
apparatus such as a waste-gate or bypass.

So, from the point of view of the turbo charger, we can con­clude that:
– The three-pulse turbocharging system is the best in terms

of system efficiency and achievable specific fuel consump­tion, and does not require ancillary apparatus.

– The high efficiencies offered by ABB turbochargers make

the use of two-pulse systems possible up to the
higher engine outputs. The turbo charger efficiency largely
compensates for the low system efficiency.

– Single pipe and constant pressure systems are generally

used on big engines (“V” engines). Ancillary apparatus is
fitted for certain applications.

Traction curve – 3-pulse Traction curve – constant pressure

Fig. 13: Comparison of the performance of 3-pulse and constant pressure systems. Note the part load difference ␭

v

.

p

[bar abs]

[bar]

p

max

4.0

L

[bar]

p

max

p

[bar abs]

L

4.0

180

140

100

60

TL [°C]

65

55

4545

2.0

1.8

1.6

1.4

TvT [°C]

600

550

500

450

400

3.0

2.0

me

[bar]

1.0

I

[kg/kWh]

e

8.0

7.0

6.0

⌬be [%]

+ 5

+ 2.5

0

– 2.5

– 5

p

max

T

L

␭

v

␭

v

T

vT

10 15 20 p

Full-load opt.

Part-load opt.

180

140

100

60

TL [°C]

65

55

4545

2.0

1.8

1.6

1.4

TvT [°C]

600

550

500

450

400

p

p

max

T

L

␭

v

␭

v

T

vT

10 15 20 p

Full-load opt.

Part-load opt.

I

⌬b

3.0

L

2.0

1.0

I

[kg/kWh]

e

8.0

7.0

6.0

e

⌬be [%]

e

+ 5

+ 2.5

0

– 2.5

– 5

[bar]

me

20 40 60

Power

80 [%]

20 40 60

Power

80 [%]

Operating turbochargers | Turbo Magazine 1/1995 19

Turbocharger compressors – the phenomenon of surging

The many different applications of turbocharged engines
make turbocharger performance and reliability key issues,
which turns the spotlight on the compressor’s performance.
Johan Schieman explains how a radial compressor works
and why surging occurs, and looks at some problems he has
experienced.

Fig. 1: Partly disassembled VTR 354-11 turbocharger. 1 = Diffusor, 2 = Air inlet casing, 3 = Silencer filter, 4 = Compressor wheel (in ducer,
impeller), 5 = Air outlet casing.

The working of a centrifugal compressor

The main parts of the turbocharger compressor are the com­pressor wheel (inducer and impeller), diffusor, air inlet and air
outlet casing. These parts can be seen in Fig. 1, which shows
a turbocharger partly dismantled. It is usual for a silencer-filter
or a suction branch to be fitted to the air inlet casing.

Fig. 2 shows a schematic of a compressor wheel. It will be
assumed that the wheel stands still. The space between
the blades is filled with air at ambient pressure and tempera­ture. We shall consider a small air mass volume at a radius r.

Fig. 2: Compressor wheel. Airmass volume at radius r.

1

2

3

4

5

r

r

20 Turbo Magazine 1/1995 | Operating turbochargers

For a given design operation range the blades of the diffusor
must be arranged with the correct angle of incidence and a
profile that reduces to a minimum the losses caused by colli­sion. This will give the highest compressor efficiency. The gen­eral performance is positively influenced by low air friction
losses on the various flow surfaces of the compressor wheel
and the diffusor.

Fig. 4 shows the characteristic of a modern compressor. At
a given speed and with increasing volume, the achievable
pressure ratio is lower due to the lower efficiency. We find the
explanation of this in Fig. 6. Theoretically, a compressor with
backswept vanes and no losses due to incidence or to friction
will exhibit a decreasing pressure ratio with increasing volume
(line a). But there is friction, and it increases with the volume
(line b). To the left and right of design point A on the constant
speed line c, the angle of incidence of the flow into the diffu­sor is not optimal. The hatched areas show the magnitude of
the losses.

The surging phenomenon

The compressor runs at constant speed and supplies air to
the air receiver of the engine, where a required pressure must
be maintained. Fig. 7 shows the line of constant speed and
the operating line of the engine. The intersection of the two
lines is the working point A. If a slight increase in air volume
occurs, more pressure is required on the operating line and
the pressure becomes lower on the constant speed line. The
volume has to decrease again to the point of equilibrium A.
If, at the same charger speed, a slight reduction in airflow
occurs, the pressure will increase although less pressure is
required on the working line. Equilibrium is then once more at
point A. The working point A is stable on the part of the con­stant speed line inclined downwards with increasing volume.

1.5

1.0

2.0

2.5

3.0

3.5

4.0

4.5

5.0

5.5

V

300

[m3/s]

␲

c

[-]

0.76

0.78

1.00

0.89

0.78

0.67

0.56

VTR 304P

1.0 2.0 3.0 4.0

n
n

0

Surge line

High efficiency

␲

C1

␲

C2

Low efficiency

0.65

Line of constant

efficiency

Line of

constant speed

V

300

[m3/s]

Volume flow

Compressor pressure ratio

0.81

0.80

0.72

1.11

Fig. 4: Compressor characteristic.

␩

TC

Turbocharger efficiency

m

.

C

Air mass flow

T

1

Air intake temperature

R Gas constant

␲

C

Compressor pressure ratio
␬ Ratio of specific heats for air
m

.

T

Exhaust gas flow
T

3

Exhaust gas temperature before turbine

␲

T

Turbine pressure ratio

␬* Ratio of specific heats for exhaust gas

␩

TC

=

m

.

C

⫻ T1⫻ R ⫻␬/(␬ – 1) ⫻ (␲C^((␬ –1)/␬)–1)

m

.

T

⫻ T3⫻ R ⫻␬*/(␬* – 1) ⫻ (1 – ␲T^((1 – ␬*)/ ␬*))

Fig. 5

C

V

Fig. 3: Rotating compressor wheel.

We start the compressor wheel rotating (Fig. 3) with a circum­ferential speed U. At radius r, the speed is V. The small air
mass volume is subjected to a radial acceleration, V

2

⁄r, which
causes it to move radially outwards. All the air experiences
this influence and air begins to flow into the air inlet casing,
through the compressor wheel and into the diffusor and
air outlet casing. The air leaves the compressor wheel at its
circumference with an absolute velocity C.

W

U

V

r

Compressor wheel
(inducer, impeller)

U = Circumferential speed

= Speed of airmass particle
C = Absolute speed of air leaving the compressor wheel,
W = Speed of air relative to blades

Diffusor

Operating turbochargers | Turbo Magazine 1/1995 21

If a slight decrease in volume occurs at point B (at the same
pressure as A), then the pressure on the constant speed
line decreases. The compressor cannot maintain the required
pressure, the volume continues to decrease and the compres­sor surges. Point B is not stable on that part of the constant
speed line that is inclined upwards (with increasing volume).

Theoretically, instability of the compressor starts where the
line of constant speed is level. Since we combine a turbo­machine, which supplies a continuous airflow, with an engine
taking air intermittently, there will always be a pressure fluctu­ation which influences the point where instability starts. It
goes without saying that this overview of the functioning of a
centrifugal compressor has been greatly simplified.

A case where surging occurred

Before we look at this case, we shall first consider how a
turbo charger specification is determined. Generally, it is based
on so-called ISO (International Standards Organisation) condi­tions, namely ambient pressure of 1 bar and ambient temper­ature of 25°C. Margins must be incorporated in the specifica­tion to take certain deviations into account.

The deviations can be due to:
– The engine application, e.g. for a ship’s propulsion with

fixed pitch propeller, for a generator drive or dredger pump
drive (full torque, variable speed).

– The ambient conditions, e.g. variations in the atmospheric

pressure or in the intake temperature.

Where there is considerable deviation from the ISO conditions,
it may be necessary to adjust the output of the engine to the
prevailing conditions. For stationary engines, a turbocharger
specification may be adapted to the site conditions.

Fig. 7: Surging cycle.

␲c [-]

Fig. 6: Compressor characteristic with backswept vanes.

Given the knowledge available nowadays, the specifications
virtually exclude the possibility of modern compressors surging.
Why then, does surging occur and especially after several
years of operation? In some of the cases that we have encoun­tered the reasons have had something to do with the way
the turbochargers have been maintained. On a 2-stroke
engine with constant pressure turbo charging, two or more
turbochargers supply air in parallel to a common air receiver.
The compressors and turbines are kept clean by waterwash­ing during operation. Sometimes, “dry cleaning” is used for
the turbines. The intervals between overhauls that include dis­mantling of the turbo chargers are long. And on ships running
a tight time schedule, not all the turbochargers can be over­hauled at the same time. The state of wear and cleanliness
changes with time. When the differences become considerable,
the turbocharger efficiency will also change. The charger in
the worst condition may start surging. The turbochargers on
the engine have practically the same intake and exhaust gas
temperature as well as the same pressure ratios over the
compressors and over the turbines (Fig. 5).

The combined performance of the turbochargers yields an
average efficiency. The turbocharger with the lower efficiency
must achieve the pressure ratio prescribed by the other turbo ­charger(s) and will react by reducing its air mass flow. This
may result in surging. The influence of turbine wear can be
seen in Fig. 8. The turbocharger with the worn turbine works
in parallel with turbochargers in a better condition. It must
achieve the pressure ratio prescribed by the others even though
it tends to run at a somewhat lower speed. A considerable
difference in the state of turbine wear may also result in surg­ing, starting from the “worst” charger.

␲c = Pressure ratio

V = Intake volume
a = Characteristic without losses
b = Characteristic with friction losses
c = Characteristic with friction and incidence losses
A = Design point

␲c [-]

Surge line

Operating line

A

a

b

c

V

Line of

constant

speed

0–+V

A

B

22 Turbo Magazine 1/1995 | Operating turbochargers

When surging begins, it is usual to first reduce the charge air
pressure by reducing the engine output slightly or blowing off
some air from the air receiver. Such situations, which arise
after many years of operation, have been met with in practice.
The time schedules of the ships being all-important, it may
be difficult to plan maintenance so that more than one turbo ­charger can be overhauled at the same time. In such cases,
the charger specification can be adapted so as to cope better
with the prevailing situation.

Modern turbochargers having compressors with backswept
vanes perform very satisfactorily on 4-stroke engines and
generally exhibit good stability. The reasons for surging occur­ring are usually due to a reduction in airflow, caused by
obstruction of the air flowpath. This may be due to:
– A dirty filter in the silencer-filter
– Increased backpressure after the turbine

(dirty exhaust boiler, dirty ex haust silencer)
– A dirty aircooler (airside)
– A high aircooler cooling water temperature
– A dirty, contaminated nozzle ring/turbine
– The ship’s hull being dirty, causing the engine to run at full

torque/reduced speed

A schematic of the flow path for the charge air and the
exhaust gas is given in Fig. 9. The various flow obstructions
generally develop simultaneously, although they will not be
equally severe. Margins for deviations from the layout data are
incorporated in the turbocharger specification. Thus, only
extreme cases of airflow obstruction are likely to make the
compressor begin to surge.

Of course, there are many more reasons (e.g. concerning the
engine and its adjustment) for the turbocharger running out­side of its design range and for surging starting. However, it
would go beyond the scope of this article to consider them. A
reasonable maintenance schedule aimed at keeping the
engine and charger(s) in good condition is the practical way to
prevent unexpected anomalies from occurring.

[%]

4-stroke engine

2-stroke engine

Fig. 8: Influence of wear on turbine blades and cover ring.

Fig. 9: Flowpath of the air and exhaust gas.

100

95

90

85

.001 .002 .003 .004 .005 .006 S / D

T

n

TC

⌬p = f (nTC)

S = Clearance

D

= Turbine diameter

T

n

= Turbocharger speed

TC

⌬p = Pressure difference

S

D

T

new worn

Air cooler Nozzle ring Turbine

Air outlet casing

From
silencer-filter

Diffusor

Compressor

To exhaust
gas boiler

Air outlet casing

From
silencer-filter

Compressor

Air cooler Nozzle ring

Diffusor

Turbine

To exhaust
gas boiler

Operating turbochargers | Turbo Magazine 2/1995 23

Turbocharger turbines: aspects of axial and radial turbines

Exhaust gas turbines of two completely different designs –

axial flow and radial flow – are used to drive turbocharger

compressors. The author considers the pros and cons of

these turbine types for various turbocharger applications.

The axial turbine

We have seen from a look at the compressors (TM 1/1995)
that air can be compressed by spinning a wheel with appro­priate blading in a suitable housing. In a turbine, the opposite
happens. Compressed air, or in the case of the turbocharger
turbine exhaust gas, is blown against suitably shaped blading
to make the turbine rotor spin and thereby generate power
with which to drive the compressor.

These ideas are not new. Fig. 1 shows a very basic example
of the axial turbine, a windmill. This particular type was in use
as early as the 13

th

century. For turbocharger turbines, the air,
or rather the gas flow has to be used much more efficiently.
The turbine is therefore equipped with a large number of blades
and the gas is guided to these blades through carefully
directed nozzles. The cut away drawing of a turbocharger in
Fig. 2 shows the arrangement of the nozzles and the turbine.

Fig. 1: The windmill – ancestor of the axial turbine. This type has been
in use since the 13

th

century.

Fig. 2: VTR..4E turbocharger: Arrangement of nozzle ring and
turbine. 1 = Diffusor after turbine, 2 = Gas outlet casing, 3 = Nozzle
ring, 4 = Turbine, 5 = Gas inlet casing.

1

2

3

4

5

Fig. 4: Enthalpy-entropy diagram for the nozzle ring and turbine assembly.

24 Turbo Magazine 2/1995 | Operating turbochargers

The flow path of the gas through the turbine is shown sche­matically in Fig. 3. Exhaust gas from the engine enters the gas
inlet casing and flows to the nozzles. The nozzles direct the
gas jets at the turbine blades, where the energy is transferred
from the gas to the turbine. As the gas passes through the
nozzle ring, its velocity increases considerably, but no power
is extracted here. The pressure between the nozzle ring and
the turbine blades, p

int

, decreases due to the gas expanding.
Further expansion over the turbine blades, down to the back
pressure after the turbine, allows power to be extracted from
the exhaust gas.

In the field of thermodynamics, it is usual to show these
processes in so-called enthalpy-entropy diagrams. The en­thalpy is the specific energy content per unit of gas mass
(Joule/kg) while the entropy is a variable of state, expressed
in Joule/kg °K.

The turbocharger has a single-stage turbine which allows the
combination of the nozzle ring and turbine to be considered
as one flow area, thereby eliminating the intermediate pressure
and temperature, which are very difficult to measure due
to the high gas velocities between the nozzle ring and turbine
blades.

Nozzle ring Turbine

r

m

.

Gas mass flow, kg/s

p

vT

Pressure before turbine, bar

T

vT

Temperature before turbine, K (°C + 273)

p

int

Pressure between nozzle ring and turbine, bar

T

int

Temperature between nozzle ring and turbine, K

p

nT

Pressure after turbine, bar

T

nT

Temperature after turbine, K

Fig. 3: Flowpath of the exhaust gas through the nozzle ring and turbine.

h [J/kg]

p

totvT

p

statvT

p

totnT

p

statnT

⌬h

s

⌬h

s [J/kg K]

␩

s

= ⌬h/⌬h

s

⌬hs Isentropic enthalpy drop, J/ kg

⌬h Usable enthalpy drop, J/kg

␩

s

Isentropic efficiency

h [J/kg]

p

totvT

p

statvT

p

totint

p

statint

p

totnT

p

statnT

s [J/kg K]

h, s, ⌬hs, ⌬h, see Fig. 4

C

o

Gas velocity corresponding to isentropic

enthalpy drop ⌬hs, m/s

C

E

Gas velocity at nozzle ring inlet, m/ s

C

in

Gas velocity at turbine blade inlet

(equal to velocity at nozzle ring outlet), m/s

C

out

Gas velocity at turbine outlet, m/ s

⌬h

blades

Enthalpy drop over turbine blades

⌬h

blades

/⌬hs Degree of reaction

⌬h

s

= Co^2/2

⌬h

C

in

^2/2

C

E

^2/2

C

out

^2/2

⌬h

blades

Fig. 5: Enthalpy-entropy diagram showing absolute velocities.

Fig. 6: Isentropic turbine efficiency versus pressure ratio ␲

TtotvT-statnT

.

VTR 254-11 WG08EF..

exhaust boile

Exhaust gas

from engine

p

Exhaust gas to

p

vT

T

vT

p

int

nT

T

T

int

nT

silencer

spark retainer

ambient

m

m

m

␩'

st

EF09

EF16

.85

EF23

.80

.75

EF09 Smallest nozzle ring

.70

EF16 Medium-size nozzle ring

EF23 Biggest nozzle ring

.65

1.5 2.0 2.5 3.0 3.5

␲

Ttot-stat

Operating turbochargers | Turbo Magazine 2/1995 25

Fig. 7: Relative isentropic efficiency versus U/ C

O

Looking at Fig. 4, it is seen that for a given pressure ratio
p

totvT/pstatnT

, ⌬hsdepicts the specific energy difference
(enthalpy drop) without losses to the surroundings. No in­crease in entropy takes place. However, the technical proces­ses we apply always incur losses. The processes are said
to be irreversible. The entropy has to increase. ⌬h depicts the
enthalpy drop that can be used. The ratio ⌬h/⌬h

s

is referred

to as the isentropic efficiency.

The gases have a certain velocity on arriving at the nozzle
ring, where the gas expands, increasing the velocity. The gas
expands further over the turbine blades, which run a certain
speed, and exits the blading at a relatively low velocity. The
gas velocity-enthalpy relationship at the different points can
be seen in Fig. 5.

The enthalpy drop between the nozzle ring inlet and the turbine
exhaust can be divided into several parts – the gas velocity
at the nozzle ring inlet, the turbine blade inlet (equal to the
nozzle ring outlet velocity) and the turbine blade outlet. The
gas velocity at the turbine blade inlet is converted into usable
output with the best possible efficiency that can be achieved.
Obviously, after the gas has left the turbine blades it cannot
perform any more work.

If the entire enthalpy drop (expansion pressure ratio) were
used to increase the gas speed in the nozzle ring, we would
speak of an impulse turbine. But if only a part is used and the
remaining enthalpy (pressure ratio p

int/pnT

) expands over the
turbine blades, we speak of a reaction turbine. The degree of
reaction is the ratio of enthalpy drop over the turbine blades
to the total enthalpy drop.

From Fig. 3 we have seen that there is an intermediate pressure,
p

int

, between the nozzle ring and the turbine blades. Different
turbine blade heights can be applied for a given turbocharger
frame size. Nozzle rings with a gas flow area ranging from small
to large may be chosen for every blade height. A small nozzle
ring area combined with a turbine of fixed area will cause the
intermediate pressure to be low. The degree of reaction will be
low. A large nozzle ring area for the same turbine will result in
p

int

being higher, as will the degree of reaction.

The degree of reaction influences the turbine efficiency. In
Fig. 6, the turbine isentropic efficiencies have been plotted
over the pressure ratios ␲

T(ptotvT/pstatnT

). The figure also
shows clearly that, depending on the engine application requi­rements, a combination of nozzle ring and turbine can be
chosen for higher turbine efficiency at full load or part-load
operation of the engine.

Turbochargers are frequently fitted to engines with a pulse
turbocharging system. The pressure ratio over the turbine is
then highly variable. To determine the influence of the turbine
efficiency on this type of operation, it is convenient to plot
the efficiency over the so-called blade speed ratio, U/C

o

. U is

the mean turbine blade speed and C

o

is the gas velocity
corresponding to the isentropic enthalpy drop over the nozzle
ring and turbine assembly. Fig. 7 shows the main details.

Fig. 8 depicts the exhaust gas pressure during the gas
exchange period of a 4-stroke engine. The mean turbine
blade speed, U, is constant, while C

o

varies with the gas

pressure and temperature. U/C

o

changes accordingly, and
with it the efficiency. Fortunately, at low pressures, when the
efficiency is low, only a small enthalpy drop and a reduced
gas mass are involved, so that not much energy is converted
with low efficiency.

Fig. 8: Period of gas exchange in a 4-stroke engine with
3-pulse turbocharging.

a

b

␩s/␩

smax

1

.75

.50

.25

0

.5 1

U/C

O

l

r

␻

c

h

⌬h

s

p

p

totvT

statnT

U

a

Relative isentropic efficiency
plotted over U/C

U Mean blade speed

b

r Mean blade radius
␻ Angular velocity
l Blade length

⌬h

Isentropic enthalpy drop

c

s

s

CO Gas velocity, CO=1⁄2 √⌬h

o

Temp.

[K]

Press.

[bar]

U/Co optimal

Exhaust Scavenging Inlet

U/Co higher

° Crank angle

T

exh. pipe

p

exh. pipe

s

26 Turbo Magazine 2/1995 | Operating turbochargers

Fig. 9: Turbine blades with profile twisted from root to tip to take
account of the increase in blade speed with the blade length.

Fig. 11: RR..1 turbocharger. 1 = Volute outlet around the turbine,
2 = Turbine, 3 = Gas inlet casing.

The nozzle ring blades have a fixed outlet angle. The shape of
a turbine blade is shown in Fig. 9. Its profile is twisted from
root to tip to take into account the fact that the circumferential
speed U increases with the radius. It can be seen in Fig. 10
that the twisted turbine blade is necessary for opti mal flow
conditions over the blade length.

To make best possible use of the energy contained in the
exhaust gases, correctly shaped gas-inlet casings must be
provided to keep pressure losses due to the flow to a mini­mum. For this reason, channels are used in which the gas
velocity gradually accelerates from the inlet to the nozzle ring
(see Fig. 2). In addition, the gas passage after the turbine
has also been configured to limit pressure losses. The diffusor
has a shape which allows the low static pressure after the
turbine to be utilized, thereby increasing the available pressure
ratio p

tot/pstat

over the turbine assembly and the correspon-

ding enthalpy drop. This layout is used in VTR. .4E turbochar­gers, where very high overall efficiencies are required.

Turbochargers with axial turbines are used on medium-speed
and slow-speed engines. They are perfectly capable of accep­ting the exhaust gas from engines running on heavy fuel oil.
The turbine retains its high efficiency over a very long period
of time, the more so when reasonable maintenance is provided.
The axial turbine is able to supply an adequate output with
good efficiency to drive the compressor from low pressure
ratios upwards, thus assuring good part-load performance of
the engine. The latter is especially important for fixed-pitch
propeller drives. Turbochargers with axial turbines are found
on ships, in diesel power stations and on diesel locomotives,
dredgers, etc.

The radial turbine

A turbocharger fitted with a radial turbine of the so-called
mixed flow type is shown in Fig. 11. This particular design, a
type RR..1, has no nozzle ring. The gas inlet casing is shaped
in such a way that the required exhaust gas velocity at the
turbine inlet is obtained by correct dimensioning of the circular
outlet of the volute around the turbine. For different applications
different gas inlet casings are needed. This design is used
especially for large series of turbochargers having the same
turbine specification. A construction with nozzle ring is
advantageous for easy adaptation.

In the enthalpy-entropy diagram, the thermodynamic process
of the radial turbine can be shown in the same way as for the
axial turbine (Fig. 5). Unlike the axial turbine, the radial turbine
has an inlet with a diameter which is larger than that of the
outlet (Fig. 13). The energy transfer to the rotor can be written
as shown in Fig. 12.

Fig. 10: Velocity diagrams for blade profiles at the blade middle,
blade root and blade tip. Comparison of profile center lines.

1

2

3

⌬h = 1⁄2·((Uin^2 – U

out

^2)– (Win^2 – W

out

^2)+ (Cin^2 – C

out

^2))

Fig. 12

VTR..4 VTR..1

Nozzle ring (stator)

W

U

U

W

in

C

in

U

root

Parallel W

C

E

CE Gas inlet speed, nozzle ring

U

in

C

in

U

C

out

W

out

Inlet turbine blade

root

Outlet turbine blade

C

out

W

out

in

Turbine
(rotor)

U

W

in

C

U

Parallel W

in

U Blade speed
C

Absolute blade inlet speed,

in

assumed constant over
blade length (equal to nozzle
ring gas outlet velocity)

W

Gas velocity relative to blade

in

C

Absolute blade outlet speed,

out

assumed axial and constant
over blade length

W

tip

tip

Gas velocity relative to blade

out

C

out

W

out

Center lines of blade profile
Blade root
Blade middle

out

Blade tip

Operating turbochargers | Turbo Magazine 2/1995 27

U

in

r

in

W

in

W

out

U

in

U

out

C

out

C

in

r

out

U

out

rin Rotor inlet radius

r

out

Rotor outlet radius

U

in

Blade speed at inlet

U

out

Blade speed at outlet

W

in

Gas speed relative to blade

W

out

Gas speed relative to blade

C

in

Absolute gas speed at blade inlet

C

out

Absolute gas speed at blade outlet

Fig. 13: Schematic of turbine rotor showing inlet and outlet radius
and speed.

Fig. 14: Isentropic turbine efficiency plotted over expansion ratio
p

totvT

/p

statnT

.

The difference between the circumferential speeds adds con­siderably to the turbine work. If care is taken to see that the
gas is accelerated while expanding in the rotor passages
(relative velocity W

out

> Win), this will also make a positive con­tribution to the turbine work. It is clear that the absolute
velocity C

out

cannot do any work and should be as low as
possible. Considering this situation, it would appear that for
small (automotive) turbo chargers the radial turbine is a
must. An axial turbine with small dimensions is not feasible in
practice. The small-size radial turbine does not suffer from
efficiency loss due to clearance leakages as much as a small
axial turbine would.

Turbochargers with radial turbines can be produced more
economically. This is the reason for the overlap of turbo­charger sizes where both axial and radial turbines are availa­ble. Their application then depends on the engine require­ments.

The efficiency of a typical RR..1 turbine is plotted over its
pressure ratio in Fig. 14. At low pressure ratios there is a wide
gap between the efficiency with a large and with a small
volute exit area. The large volute exit area causes the degree
of reaction to be relatively high, so that the influence of
centrifugal force working in a direction opposite to the gas
flow has less influence than when the volute exit area is
small. At high pressure ratios the influence of the centrifugal
force on the gas particles is not so significant due to the
much higher energy content of the gas. The small volute exit
area exhibits high efficiency for a relatively low degree of
reaction, the large volute exit area exhibiting low efficiency.

Under adverse conditions – gases with a low energy content,
high turbocharger speed (e.g. a 4-stroke engine at full
speed with no load) – the turbine may tend to act as a centri­fugal compressor due to the centrifugal force acting on
the gas particles.

Turbocharging of automotive engines is already under way, so
that small turbochargers with radial turbines can be said
to have found their place in the engine industry. Their use on
high-speed industrial engines is also widely accepted. How­ever, for bigger engines running on poor-quality fuels the axial
turbine is generally the preferred type. ABB supplies turbo­chargers with radial turbines for engines rated at up to about
2000 kW.

The above overview is just that – an overview. Details of the
performance and operation of the turbocharger turbines
would go beyond the scope of this article. If you have a ques­tion that the article does not answer, please send it to us. We
will do our best to answer it.

RR 151 WG04AF..

␩'

sT

.85

.80

.75

Gas inlet casing

AF06 Small volute exit area

AF10 Medium volute exit area

AF15 Large volute exit area

1.5 2.0 2.5 3.0 ␲

AF10

AF15

AF06

T

28 Turbo Magazine 1/1996 | Operating turbochargers

Turbocharging over the years

Turbocharging has played a vital role in the development of
the diesel engine. This article takes us back to its beginnings
and charts its progress to the present day.

The idea of supplying air under pressure to a diesel engine
was voiced by none other than Dr. Rudolf Diesel as early as

1896. The use of a turbocharger for this purpose, however,
was the result of work by a Swiss, Alfred Büchi, who patented
his so-called “pulse system” in 1925. This system feeds the
exhaust gases of the engine through narrow pipes to the
turbocharger turbine, thus driving the compressor. The pres­sure variation in the small-volume pipes allows overlapping
of the inlet and exhaust, permitting scavenging of the com­pression space of the engine cylinders with clean air. Cylinders
that do not disturb each other’s scavenging process can be
connected to one pipe (turbine gas inlet) in accordance with
the firing order of the engine (see Fig. 1). This pulse system
was the foundation for the future success of turbocharging.

In December 1928, Alfred Büchi gave a lecture at the Royal
Institute of Engineers at The Hague in the Netherlands. From
this lecture, we learn that already then it was known that the
thermal load of a diesel engine does not essentially increase
when turbocharged (Fig. 2). An improvement in fuel consump­tion due to the better mechanical efficiency could also be
shown. And it was seen from a comparison of four turbo ­charged 10-cylinder single-acting 4-stroke engines and four
turbocharged 6-cylinder double-acting 4-stroke engines used
for a ship’s propulsion installation rated at 36,000 bhp that the

double-acting engines offered several advantages. Neither
single-acting nor double-acting 2-stroke engines could rea­sonably compete. Thought was given very early on to turbo ­charging the engine-driven scavenging pumps of the 2-stroke
engine. However, since the turbocharger efficiency necessary
for this was not available, it took many more years for this
goal to be reached.

With the market introduction of the VTR ..0 turbocharger, it
became possible to turbocharge 2-stroke engines with
engine-driven scavenging air pumps. 2-stroke engine builders
began to make use of this possibility in a big way after 1945.
However, to eliminate the scavenging pumps and reduce fuel
consumption, the turbocharging system has to be a pulse
system. This feature was introduced successfully in 1951 on a
B&W marine propulsion engine.

Fig. 1: Diagram showing the pressure p

cyl

in the cylinder, p

exh

in the

exhaust pipe and p

r

in the air receiver of an 8-cylinder engine. The
scavenging period, where the inlet and exhaust are simultaneously
open, is not disturbed by the exhaust pulses of other cylinders.

Fig. 2: Some engine test data from the lecture given by Alfred Büchi
in 1928.

p [bar]

p

cyl1

p

exh.

p

R

180 360 540 °CA

Scavenging period

TDC1360 °

1271

850

Output bhp

185

178

Specific fuel cons. g/ bhph

380 380

Exhaust gas temperature °C

565,000

550,000

Heat in the cooling water kcal/ h

Exhaust gas

Engine friction

Air compressor

fuel injection

Cooling water Cooling water

Effective output Effective output

Heat distribution in %

100

50

Turbocharged diesel engineNormal diesel engine

Exhaust gas
Engine friction
Air compressor
fuel injection

Exhaust gas
turbine output

0

Operating turbochargers | Turbo Magazine 1/1996 29

Use of the VTR-type turbochargers on 2-stroke and 4-stroke
engines increased rapidly. Through continuous development,
improvements were made to the achievable compressor pres­sure ratios, diesel engine designs (for higher permissible firing
pressures) and fuel injection systems. These improvements and
broad utilization of the thermodynamic properties of the diesel
cycle have led to the current state of the art of turbocharging.

2-stroke engines

It is interesting to look at some of the development stages
through which turbocharging of 2-stroke engines has passed.
For the sake of simplicity, longitudinally scavenged engines
(i.e. with air inlet ports and exhaust valves) will be considered.

The air supply of a 2-stroke engine without turbochargers is
provided by scavenging pumps. The VTR. .0 turbochargers
were used in two lines of engine development, one retaining
the scavenging pumps and one omitting them but using the
pulse system to obtain sufficient energy to drive the turbo ­charger(s) over the entire load and speed range of the engine.
The typical arrangements of the turbochargers are shown in
Fig. 3. Fig. 4 compares the timing of the inlet and exhaust for
the constant pressure scavenging pump system with that of
the pulse system without scavenging pumps. Fig. 5 shows a
Götaverken engine with an 850 mm bore, a stroke of 1700 mm
and 11 cylinders, each cylinder having two double-acting
scavenging pumps with a 480 mm bore, turbocharged with
three VTR 630 units. A Burmeister & Wain engine of type
10 K98FF with four VTR 750 turbochargers in pulse operation
is shown in Fig. 6. The photographs were taken around 1970.

Arrangement for 2-stroke engine
with pulse turbocharging
(B&W 1970)

Arrangement for 2-stroke engine with constant pressure turbocharging
and scavenging pumps in series with the turbocharger compressor
(Götaverken 1970)

Fig. 3: Diagrams of 2-stroke engines featuring pulse and constant pressure turbo charging.

Turbine

Compressor

Aircooler

Airflow

Exhaust gas

Engine

Compressor

Exhaust gas receiver

Turbine

g/bhph

Engine

Aircooler

Engine driven

scavenging

air pump

30 Turbo Magazine 1/1996 | Operating turbochargers

constant pressure turbocharging on the B& W engines came
in due course, and the expected reduction in fuel consump­tion with it (Fig. 7). An electrical auxiliary blower had to be
used to start the engine and for low part-loads.

The higher pressure ratios and higher efficiencies with the
VTR . .4 turbochargers allowed further increases in output and
a reduction in the specific fuel consumption. A better under­standing of the thermodynamics of the diesel cycle and the
properties of the air and gas flows has resulted in only the
longitudinally scavenged 2-stroke diesel engines being able
to hold their own in the marketplace. VTR . .4D turbochargers
(combining high pressure ratio and high efficiency) meet
the present-day requirements of the engine manufacturers
(actual status: pressure ratio 3.7:1, mean effective pressure
18 – 19 bar).

Cylinder expansion line

bar

cm

2

Degr. crank angle

Inlet port
area

Pulse timing

Constant pressure timing

EO IO BDC IS ESEO

CP

ES

CP

Timing diagram

Exhaust valve
area

The advantage of the pulse system was that no mechanical
system was necessary at any point in the engine load range
to guarantee the air supply. This was also due to the very high
mechanical efficiency of the VTR turbochargers. The exhaust
gas energy needed was obtained by opening the exhaust
valve early (Fig. 4). However, this reduced the effective expan­sion stroke of the engine.

The engine with scavenging pumps and featuring constant
pressure turbocharging can make do with a much later open­ing point for the exhaust valve, but takes part of the energy
for the air supply from the crankshaft. The engine has to drive
its own scavenging pumps. Thus, the fuel consumption of
these types of engine was not essentially different, amounting
to 150 – 155 g/bhph (204 – 211 g/kWh).

From the above, it is obvious that the combination of constant
pressure turbo charging and no scavenging pumps would
improve the fuel consumption. This became possible with the
introduction of turbochargers with improved efficiency. The
VTR . .1 was introduced in 1971. Changeover from pulse to

Fig. 4: Timing of inlet and exhaust for pulse and constant pressure
turbocharging.

Fig. 5: Götaverken engine of type VGS11U, rated 26,400 bhp,
19,400 kW at 119 rev/min. Three VTR 630 turbochargers are installed.

Fig. 6: Eriksberg B&W engine of type 10K98FF, rated 38,000 bhp,
27,950 kW at 103 rev/min, fitted with four VTR750 units.

Operating turbochargers | Turbo Magazine 1/1996 31

The next turbocharger generation, the TPL series, is in the
process of being introduced and will allow further improve­ments in engine output and fuel consumption.

4-stroke engines

The turbocharging of 4-stroke engines has, of course, also
been strongly influenced by the availability of adequate turbo ­chargers. While the 4-stroke engine is basically self-aspirating,
it is not so de pendent on high turbocharger efficiency. What is
it then, that makes turbo charging so attractive?

The turbocharger allows:
– an increase in engine output for a given speed which

is approximately proportional to the absolute charging
pressure supplied by the turbocharger,

– a reduction in required space for the installation of a

specific engine output,

– more favorable fuel consumption.

g/bhph

Fig. 7: Improvement in fuel consumption due to later opening of the
exhaust valve.

Fig. 8: Increase in compressor pressure ratio over the years.

The increase in compressor pressure ratio over the years is
shown in Fig. 8. Twenty years were needed to increase it from
about 1.3 to 2 in 1970. Initially, a given 4-stroke engine would
have a compression pressure of, say, 35 bar, and a maximum
firing pressure of 50 bar. To achieve a charging pressure of

0.3 bar gauge, the compression ratio of the engine would be
reduced from approximately 13.4 to 11.1. The compression
and maximum firing pressure would then remain the same,
but the output would increase by some 30 percent. This
would eventually have a negative influence on starting the
engine.

Encouraged by the results achieved, the engine manufacturers
designed for higher pressures. The introduction of air-coolers
to reduce the temperature of the compressed air entering the
engine cylinders helped to increase the engine output at more
moderate charging pressure levels (e.g. a mean effective pres­sure

of 10 bar, a charging pressure of 0.6 bar gauge with air­cooler and 0.85 bar without air-cooler, for the same charge of
air (kg) in the cylinders; without an air-cooler the exhaust gas
temperature would be some 50°C higher).

␲

c

12

9

6

3

Improvement in fuel consumption

0

95 90 85 80

Exhaust open before bottom dead center

⌬be

° crank angle

5

4

3

2

Compressor pressure ratio

1

70 72 74 76 78 80 82 84 86 88 90 94 9692

Range for most applications

max

medium

Year

32 Turbo Magazine 1/1996 | Operating turbochargers

Fig. 9 shows the difference in the cylinder diagram of a
4-stroke engine with and without turbocharging. It is easily
seen that the turbocharged engine has considerable advan­tages. In the T-S diagram the turbocharged engine shows a
smaller increase in entropy (S) than the non-turbocharged
engine. This results, thermodynamically, in better efficiency.
In this diagram, the slope of the lines of constant pressure
becomes steeper with increasing pressure. From the p-V
diagram, it is seen that the area representing positive work
increases considerably. The turbocharged engine shows posi­tive work in the area representing the gas exchange. For the
non-turbocharged engine, only negative work is shown. The
area representing positive work includes the output required
to overcome the friction of the engine. This friction is largely
dependent on the speed of the engine and to a lesser extent
the pressures. This is the reason for the mechanical efficiency
of an engine steadily increasing with the degree of turbo ­charging. Together, these factors mean that, essentially, a
turbocharged engine will exhibit decreasing specific fuel con­sumption with an increasing degree of turbocharging. In
addition, deeper knowledge of the thermodynamic properties
of the diesel process and a better understanding of the fuel
injection systems have helped to substantially reduce the spe­cific fuel consumption of diesel engines.

With the steady increase in the mean effective pressures, the
exhaust gas temperatures have remained approximately
constant due to the higher charging pressures. This and the
cooling of the air increases the density of the air entering the
cylinders. In this way, the ratio of air-to-fuel (kg/sec /kg/ sec)
could be kept more or less constant. The distribution of the
heat supplied with the fuel, however, changes considerably

Fig. 10: Distribution of the heat introduced with the fuel for an
increasing level of turbocharging.

p

Fig. 9: Comparison of the cylinder diagrams of a 4-stroke engine with and without turbocharging.

T

p

charging pressure

maxpmax

p

before turbine

p

t

after compressor

t

after air-cooler

ambient

p

with turbocharging

without turbocharging

p

charging pressure

p

before turbine

p

ambient

S

Heat introduced

%

with the fuel

80

Heat into effective
output

60

40

20

400

Heat into the exhaust gases after
engine cylinders (supplied to the
turbocharger turbine for further
utilization)

Heat into cooling water,

0

%

lube oil, radiation

300

200

Heat introduced
with the fuel

100

0

100 200 300 400 [%]

without with with with turbocharging

Output

V

Operating turbochargers | Turbo Magazine 1/1996 33

with the increasing level of turbocharging. This is illustrated in
Fig. 10. The heat in the lubricating oil represents roughly the
output needed to overcome friction, i.e. the mechanical effi­ciency of the engine. The heat in the cooling water is roughly
equivalent to the heat load of the combustion chamber parts.
With turbocharging, we succeed in keeping the heat flow to
the cooling water, in absolute terms, approximately constant,
this being largely independent of the level of turbocharging.
This is also one of the most important successes of turbo ­charging. Alfred Büchi mentioned this in his lecture of 1928,
and it still holds true today. The very simplified explanation is
shown in Fig. 11. The heat transmission coefficient from the
gas to the wall depends on the pressure (公p

gas

). The heat
transmission coefficients through the wall and from the wall
to the cooling water have hardly changed. The process tem­peratures still being at the same level as without turbocharging
(fuel-air ratio) and the surface areas still the same, the heat
flow to the cooling water could not change very much.

Further advantages of turbocharging

The state of the art for modern engines is an output 4 times
(400 percent) as high as the non-turbocharged engine for the
same speed and dimensions. From Fig. 10, it can be seen
that the heat introduced with the fuel at 400 percent output
has increased to only 320 percent. This means that the spe­cific fuel oil consumption (sfoc) of modern engines amounts to
only 80 percent of the sfoc of the non-turbocharged engine.
Thus, turbo charged engines are a must in terms of environ­mental protection. Fuel that is not burnt neither causes air
pollution nor produces CO

2

. Also, the modern turbocharger
combining high pressure ratios with high efficiency is an
important factor when the engine process has to be adjusted

to achieve a low NO

x

emission level through low process
temperatures (high air-to-fuel ratio, high charging pressure).
Put in general terms, the materials and energy required to
produce a turbo charged engine for a given output and speed
are considerably less than for a non-turbocharged engine for
the same output and speed.

To complete this overview, Fig. 12 shows how the turbocharger
efficiencies have developed up until the present day. As
high efficiencies allow a large turbine area and this in turn
reduces the work necessary to expel the exhaust gas from the
4-stroke engine cylinders, the full-load fuel consumption is
influenced favourably. The 2-stroke engine may further reduce
the timing for the inlet and exhaust, or allow a power turbine
to be installed to convert excess exhaust gas energy into use­ful output.

Looking to the future, ABB is preparing to introduce its new
TPS and TPL turbocharger designs combining compactness
with high pressure ratios and high efficiencies. Since non­turbocharged engines will hardly be built anymore, the future
looks bright for our turbochargers.

␩

Fig. 11: Heat transfer from the hot-gas side to the cooling water
through the wall of the cylinder liner.

Fig. 12: Increase in turbocharger efficiency as a result of continuous
development over the years.

T

gas

T

cooling water

TC

[%]

70

65

60

55

VTR..4E

VTR..4

VTR..1

VTR..0

Cylinder liner

50

1970 1975 1980 1985 1990 1995

Year

34 About the author | Operating turbochargers

About the author

From 1992 to 1996 Turbo Magazine carried a series of articles
by Johan Schieman on the subject of oper ating turbochargers.
The series came to an end with Johan’s death in April 1996.
This collection of his arti cles is both a response to the many
requests received from readers and a modest memorial to
the author.

Johan Schieman’s insights into the workings of turbochargers
made the articles an instant and eminent success. Highly
instructive, they were much appreciated, especially by ship’s
operators, who kept additional copies on board their vessels
as reference for their engine crew. Schools used the articles
for teaching and commercial trade journals asked for permis­sion to reprint them.

In 1956 Johan Schieman joined ABB (then Brown Boveri) as
a sales and application engineer at our Rotterdam offices. His
many visits to Baden made him a familiar face at headquar­ters, and in 1967 he was persuaded to stay “for three years”.
He remained in Baden for the rest of his life. During 38 years
with ABB Turbochargers Johan became a reputed specialist in
the turbo charging of 2-stroke diesel engines and an experi­enced China hand to boot. From 1985 until 1987 he was the
company’s liaison to licensees and engine builders in China,
based in Hong Kong. In the fol lowing years he visited China
two or three times a year. His interest in foreign cultures went
as far as to study the Chinese language, along with other
tongues like Spanish and Swed ish. The latter was the result of
having worked in Scandinavia, where he was instrumental in
the introduction of ABB turbo charging technology for 2-stroke
diesel engines at Götaverken and B&W.

Johan Schieman had that rare talent of being able to motivate
others to take a genuine interest in his work. Through his
ability to communicate, the sincerity of his conviction and the
perseverance with which he pursued his goals, he became
the mentor of many younger application engineers in our
company and abroad.

CHTUS-1310-1012-EN PDF © 2010 ABB Turbo Systems Ltd, Baden/Switzerland

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