Turbocharger
Two stroke crosshead engines must be supplied
with air above atmospheric pressure for it to work.
Although turbochargers were developed in 1925, it
was not until 1950’s that large two stroke engines
were turbocharged. Pressurized air is needed to
“scavenge” the cylinders of the exhaust gases and
supply the charge of air for next combustion cycle
was provided by mechanically driven air
compressors (roots blower) or by using the space
under the piston as a reciprocating compressor
(under piston scavenging). This of course meant
that the engine was supplying the energy to
compress the air, which meant that the useful work
obtained from the engine was reduced by this
amount.
• About 35% of the total fuel energy goes out
in the exhaust gas. The turbocharger uses
7% of the total energy (20% of the exhaust
gas energy) to drive a single row turbine. The
turbine shaft drives a rotary compressor. Air
is drawn and compressed. Due to
compression, the air temperature rises.
Hence it is cooled in a cooler to increase its
density and then sent to the air inlet manifold
or scavenge air receiver. At full power of
diesel engine, the turbocharger may be
rotating at > 10000rpm.
• Power of a two stroke diesel engine = pm x L x A x N
x no. of cylinders
• Where pm = mean effective pressure; L = stroke
of the engine;
• A = cross sectional area of the cylinder; N =
revolution per second of the engine
• Thus to increase the power of the engine of given
swept volume i.e. the power to weight and volume
ratios of the engine we have to increase either mean
effective pressure or the revolution per second of
the engine. The approach of increasing power output
by increasing speed is unattractive, due to rapid rise
of mechanical and aerodynamic losses, and the
corresponding fall in brake thermal efficiency.
• For increasing the mean effective pressure, more
fuel has to be burnt during the cycle of the engine,
which requires higher quantity of air per cycle of the
engine. The purpose of supercharging is to increase
the mass of air trapped in the cylinders of the
engine, by raising its density. A compressor is used
to achieve the increase in air density. Two methods
of supercharging can be distinguished by the
method to drive the compressor. If the compressor
is driven from the crankshaft of the engine, the
system is called ‘mechanically driven supercharging’
or often just ‘’supercharging’. If a turbine drives the
compressor, which itself is driven by the exhaust
gas from the cylinders, the system is called
‘turbocharging’. The shaft of the turbocharger links
the compressor and the turbine, but is not
connected to the crankshaft of the engine.
• The advantage of the turbocharger, over a mechanically
driven supercharger, is that the power required to drive
the compressor is extracted from the exhaust gas energy
rather than the crankshaft. Thus turbocharging is more
efficient than mechanical supercharging. However the
turbine imposes a flow restriction in the exhaust system,
and therefore the exhaust manifold pressure will be
greater than atmospheric pressure. If sufficient energy
can be extracted from the exhaust gas, and converted
into compressor work, than the system can be designed
such that the compressor delivery pressure exceeds that
at turbine inlet, and the inlet and exhaust processes are
not adversely affected. For a compressor pressure ratio
of 5, allows to increase the specific power output of the
engine by 400%.
Turbocharging systems – principles
•
Consider now an ideal
turbocharged four-stroke
engine. Turbocharging raises
the inlet manifold pressure,
hence the inlet process (12-1) is
at pressure P1, where P1, is
above ambient pressure Pa. The
blow-down energy is
represented by area 5-8-9. The
exhaust manifold pressure (P7)
is also above the ambient
pressure Pa. The exhaust
process from the cylinder is
represented by line 5,13,11,
where 5,13 is the ‘blow-down’
period when the exhaust valve
opens and high pressure gas
expands out into the exhaust
manifold.
• Process 13, 11 represents the remainder of the exhaust
process, when the piston moves from BDC to TDC
displacing most of the gas from the cylinder to exhaust
manifold. This gas is above ambient pressure and
therefore also has the potential to expand down to
ambient pressure whilst doing useful work. The potential
work that could be done is represented by the hatched
area 13-9-10-11. This work is done by the piston but
could be recovered by the turbine in the exhaust. It will
be called the piston pumping component of the exhaust
energy.
• The maximum available energy to drive a turbine will be
sum of areas 5-8-9 and 13-9-10-11, but it is impossible
to devise a practical system that will harness all this
energy. To achieve this, the turbine inlet pressure must
instantaneously rise to P5 when the exhaust valve
opens, followed by isentropic expansion of the exhaust
gas through P7 to the ambient pressure (P8 = Pa).
During the displacement part of the exhaust process, the
turbine inlet pressure would have to be held at P7. Such
a series of process is impractical.
• Consider a simpler process that would occur if a larger
chamber were fitted between the engine and the turbine
inlet in order to damp down the pulsation in exhaust gas
flow. The turbine acts as a flow restrictor creating a
constant pressure (P7) in the exhaust manifold chamber.
The available energy at the turbine is given by area 7-810-11. This is the ideal ‘constant pressure turbocharging
system’. Next consider an alternative system, in which a
turbine wheel is placed directly downstream of the
engine, very close to the exhaust valve. The gas would
expand directly through the turbine along line 5-6-7-8,
assuming isentropic expansion and no losses in the
exhaust port. If the turbine were sufficiently large, both
cylinder and turbine inlet pressure would drop to equal
ambient pressure before the piston has moved
significantly from BDC. Thus the piston pumping work
would be zero during the ideal exhaust stroke and area
5-8-9 would represent the available energy at the
turbine. This is the ideal ‘pulse turbocharging system’.
Principles of constant pressure turbocharging
• With constant pressure turbocharging, the exhaust ports from all
cylinders are connected to a single exhaust manifold, whose volume
is sufficiently large to ensure that its pressure is virtually constant.
The unsteady exhaust flow processes at the cylinders are damped
into a steady flow at the turbine. Only one turbocharger need be
used, with a single entry from the exhaust manifold, but frequently
several smaller units are fitted so that a reasonable boost pressure
can be obtained in the event of a turbocharger failure. A major
advantage of the constant pressure system is that turbine inlet
conditions are steady and known, hence the turbine can be matched
to operate at optimum efficiency at specified engine conditions. The
main disadvantage is that the available energy entering the turbine
is low, since full advantage has not been taken of the pulse energy.
In Figure4 area 7-8-10-11 denotes energy available to the turbine,
hence the energy represented by area 5-7-13 cannot be used. This
energy is not lost, since energy loss only occurs by heat transfer, but
since no work is done during the pressure reducing process 5-7 it
represents a loss of potential turbine work.
•
•
•
Typically, a constant pressure exhaust manifold will consist of a large
diameter pipe running along the exhaust side of an engine, with each
exhaust port connected to it via a short stub pipe. On a ‘vee’ engine, the
large bore manifold will usually lie between the banks with the inlet valves
and manifolds arranged to be on outside.
The volume of the exhaust manifold should be sufficient to damp pressure
pulsations down to a low level. For guidance however, it can be stated that
the volume would normally be in the range of 1.4 to 6 times the total swept
volume of the engine.
If the exhaust manifold volume is not sufficiently large, the ‘blow-down’ or
first part of the exhaust pulse from a cylinder will raise the general pressure
in the manifold. Since all cylinders are connected to the same manifold, it is
inevitable (if the engine has more than three cylinders) that at the moment
when the blow-down pulse from one cylinder arrives in the manifold,
another cylinder is nearing the end of its exhaust process. The pressure in
the latter cylinder will be low; hence any increase in exhaust manifold
pressure will impede its exhaust process. This will be particularly important
where the cylinder has both inlet and exhaust valves or ports partially open
and are relying on a through-flow of air for scavenging the burnt combustion
products. A rise in exhaust manifold pressure at this time is virtually
inevitable in an engine with more than three-cylinders, unless the volume is
large. This will be particularly important on a two-stroke engine, since if the
exhaust pressure exceeds inlet pressure during ‘scavenging’, the engine
cannot run at all.
• From a purely practical point of view, the exhaust manifold is simple
to construct although it may be rather bulky, particularly relative to
small engines with few cylinders. However, for large engines with
many cylinders, the convenience of being able to join all cylinders to
a common exhaust manifold with a single turbocharger on top or at
either end is useful. A major disadvantage of the constant pressure
system arises from the use of an exhaust manifold having large
volume. When the engine load is suddenly increased or a rapid
engine speed increase is required, the pressure in the large volume
is slow to rise. Hence the energy available at the turbine increases
only gradually. Turbocharger, and therefore engine response, will be
poor. The poor response of the constant pressure turbocharging
system restricts it from consideration for applications where frequent
load (or speed) changes are required.
• The turbine inlet temperature (T3), the inlet manifold pressure (P2)
and exhaust manifold pressure (P3) are functions of overall turbine
efficiency () and to lesser extent, the air-fuel ratio (AFR) also plays
its part. The air-fuel ratio at full load will be governed by thermal
loading or onset of black smoke in the exhaust. The turbine inlet
temperature will also be dependent on air-fuel ratio, the amount of
cool scavenge air passing through the cylinders and heat loss from
the exhaust manifold.
Principles of pulse turbocharging
•
Here the effect of varying pipe
length on the timing of the
reflected pulse is shown. In case
1, the reflection occurs after the
exhaust valve has closed causing
no problem, but this is a rare case
since it can occur with an
exceptionally long manifold. More
common is case 4, in which the
reflection time is very short
relative to the valve opening
period. Case 2 is the serious one,
that can occur with long pipes, of
the reflected pulse raising exhaust
pressure at the valve or port,
during the scavenge period. The
turbocharger position and exhaust
pipe length must be chosen to
avoid this situation, or scavenging
will be seriously impaired.
Principles of pulse turbocharging
• On a multi-cylinder
engine, narrow pipes
from several cylinders
can be connected via a
single branch manifold to
one turbine. Consider a
three cylinder four-stroke
engine. Due to the phase
angle between cylinders
the opening periods of
the exhaust valves follow
successively every 2400
with very little overlap
between them.
• Thus a steady ‘train’ of
pressure pulses arrives at
the turbine, virtually
eliminating the long periods
of pure windage, although
the average turbine
efficiency will remain lower
than that obtained with
correctly matched constant
pressure system (operating
near the peak of the
efficiency curve). The
remaining important point to
consider is the exhaust
pressure close to the valve,
during the valve overlap
(scavenging) period.
•
•
As with the constant pressure system, a good pressure drop between inlet
and exhaust manifold during the period when both valves are open is
important in the case of a four-stroke engine with significant valve overlap
and vital for a two-stroke engine. The pressure history in the inlet and
exhaust manifolds (at a valve) is shown, the pressure drop during the period
of valve overlap being hatched. Clearly, at the running condition shown, the
pressure drop is satisfactory.
The diagrams have shown the way to increase average turbine efficiency by
reducing windage periods, whilst avoiding interference with scavenging of
one cylinder due the effect of the blow-down pulse from another. The
pressure pulse exhausting from a cylinder travels along the manifold until it
reaches a junction. At the junction it divides into two pulses (each of smaller
magnitude due to the effective area increase) one travelling down each
adjacent pipe. One pulse will travel towards the turbine; the other will arrive
at the exhaust valve of another cylinder. It is the latter pulse, from cylinder
number 3, that has arrived near cylinder 1 just at the end of the scavenge
period of cylinder 1, that could be a problem. If it had arrived earlier
(perhaps due to shorter exhaust pipes) it would have interfered with
scavenging. This type of interference due to the direct action of a pressure
wave from another cylinder is quite separate from the action of a pressure
pulse reflected from the turbine, whether the latter started from cylinder 1,2
or 3.
• Most engines have four or
more cylinders. Let us consider
a six cylinder engine. If all six
cylinders were connected to a
single entry turbine via narrow
pipes, the pressure waves
from each cylinder would
significantly interfere with the
exhaust processes of each
other during valve overlap and
the exhaust stroke, thus
increasing piston pumping
work. The effect would be poor
engine efficiency. A two-stroke
engine might not operate
under these conditions.
• The difficulty can be avoided
by simply connecting the
cylinders in two groups of
three, either to two different
turbines, or separate entries of
a single turbine. If the correct
cylinders are grouped together,
then the pressure pulse
system in each group will be
the same as that shown in
earlier figure. It may be
concluded that the six-cylinder
engine is similar to the threecylinder, from the
turbocharging point of view,
but turbine performance may
be slightly worse due to the
losses associated with the join
of two sectors of a divided
entry turbine.
• It is disadvantageous to connect more than three
cylinders to a single turbine entry. Thus for the fourcylinder engine, pairs of cylinders (1-2 and 3-4) would be
connected to a double entry turbine. On engines with
other numbers of cylinders, the general rule will be to
connect cylinders whose firing sequences are separated
by 2400 crank angle (in case of four-stroke) and 1200
(two-stroke) to a turbine inlet, and select those cylinders
whose exhaust processes are evenly spaced out.
However, this is not always possible. For example, on a
vee-form engine, the vee angle will introduce an
additional phase difference to the firing intervals between
cylinders.
• The principal advantage of the pulse over the constant
pressure system is that the energy available for
conversion to useful work in the turbine is greater.
However, this is of little value if the energy conversion
process is inefficient.
•
With three cylinders to a turbine entry the average turbine efficiency will be
much higher since windage is almost eliminated. The efficiency is better still
if the valve timing permits a larger overlap by having longer exhaust periods
(2900) as is the practice in medium speed diesel engines. However, turbine
efficiency, averaged over the unsteady flow cycle, will be lower than
obtained in a well matched steady flow system. If two cylinders were
connected to a turbine entry the average turbine efficiency will be lower than
would be the case with three cylinders, since (short) windage periods would
exist . Thus pulse turbocharging system is most suitable for those engines
whose exhaust manifolds may connect groups of three cylinders to a turbine
entry. However, even if this is not possible, the loss in turbine efficiency due
to partial admission and unsteady flow is usually more than offset by the
additional energy available at the turbine, hence the pulse system is by far
the more widely used. In practice, the constant pressure system is used
exclusively on very large, highly rated two-stroke engines. On these engines
the ratings are such that very large pressure pulses would be generated
with the pulse system. Since most of the exhaust pulse energy coincides
with the peak of the pulse, matching this point with high instantaneous
turbine efficiency is important. In practice it is difficult to maintain high
turbine efficiency when the pressure ratio exceeds 3:1, hence turbine
efficiency will be low if exhaust pressure pulse amplitude substantially
exceeds this value. This is what happens on very highly rated engines;
hence constant pressure systems operate with higher turbine efficiency,
more than offsetting their lower available energy.
Pulse turbocharging of two-stroke engines
• The major difference of 2stroke engine with
respect to 4-stroke
engine is that the exhaust
pressure diagram is
somewhat different, due
to the long scavenge
period of the two-stroke
engine. It tends to consist
of quite distinct ‘blowdown’ and ‘scavenge
period’.
Pulse turbocharging
Advantages
Disadvantages
1. High available energy at turbine
1. Poor turbine efficiency with one or two cylinders
per turbine entry
2. Good performance at low speed and load
2. Poor turbine efficiency at high ratings
3. Good turbocharger acceleration
3. Complex exhaust manifold with large number of
cylinders
4. Possible pressure wave reflection problems on
some engines
Constant pressure turbocharging
Disadvantages
Advantages
1. High turbine efficiency due to
steady flow
1. Low available energy at turbine.
2. Good performance at high load
2. Poor performance at low speed and
load
3. Simple exhaust manifold
3. Poor turbocharger acceleration
Compressor characteristic and the surge limit
• Centrifugal compressor
characteristics are similar to
those of centrifugal pumps. At
a constant RPM, the
characteristic would appear
similar to the figure. At
constant speed the discharge
pressure first rises as
volumetric flow increases and
then drops off rather sharply.
The compressor efficiency
curve also rises to a peak,
although at any constant this
peak is to the right of the
pressure peak. The power
consumed by the compressor
is related to the product of
discharge pressure and flow
rate.
• In the region to the right of the peak in pressure curve,
operation will be stable: in this region a momentary drop
in volumetric flow rate, for example, perhaps brought on
by a momentary reduction in engine speed, will be
countered by a rise in pressure, with little or no effect on
the turbine. In the region to the left of the pressure peak,
a momentary drop in volumetric flow rate will be
accompanied by a drop in discharge pressure and a
reduction in compressor power consumption. Operation
in the unstable area to the left of the pressure peak may
result in compressor surge. As the pressure at the
compressor discharge falls below that downstream, the
flow can reverse. The result can simply be a pulsation if
the situation is not severe or of long duration, or the
reversed flow can continue to the air intake and become
audible, ranging in volume from a soft sneezing to a very
loud backfiring sound.
• Obviously, operation in the
surge region should be
avoided; consequently,
turbocharger designers
establish a line, called a surge
limit, through the pressure
characteristics slightly to the
right of the peak. Similar data
as previous figure are obtained
at several constant speeds
covering the range of
operation, and plotted together
on the same axes. The
resulting compressor
performance map is shown.
Turbocharger matching
•
•
Inlet conditions of the
compressor P1 (after
pressure drop across air
suction filter) and T1 are
selected.
An estimate is made of the
power of the engine at a
particular engine RPM. Also,
an estimate is made of
amount of air, ma, the engine
would require at above
conditions.
ma = Vsw x a x vol x N (for
two stroke diesel engine).
The pressure in the air
manifold and temperature
are estimated to get the a.
•
•
A drop across air cooler is
assumed and added to air
manifold pressure to get the
compressor discharge
pressure P2. The
compressor pressure ratio
P2/P1 can now be
calculated.
The compressor frame size
and its diffuser can now be
selected by entering the
family of compressor maps
(Figure 20 and Figure21)
with the values of P2/P1 and
Va or (ma T0/P0). The
operating point must have
adequate margin from surge
limit i.e. it must be 10% to
20% to the right of the surge
limit at the value of P2/P1.
•
•
•
•
•
•
•
•
From the performance map of the selected compressor and
diffuser, the compressor efficiency (c) and turbocharger RPM are
read at the operating point.
The required power to drive the compressor is given by
Wc = ma Cpa T1 [ (P2/P1) (-1)/  -1]/ c
The required turbine power Wt = Wc / mech
where mech is the known characteristic of the bearings.
Again an estimate is made of the gas conditions at the turbine
inlet i.e. P3 and T3 from basic principles and empirical data
including previous performance.
The turbine outlet pressure P4 is also estimated by adding an
amount to atmospheric pressure to allow for typical losses
through exhaust gas economiser. Then expansion ratio P3/P4 is
calculated.
The mass flow rate of gas, mg, is calculated by adding mass flow
rate of fuel, mf, to mass flow rate of air, ma, which was estimated
previously.
•
In general, the selection of a
compressor wheel diameter
predetermines turbine
characteristics, which
includes wheel diameter and
blade length. With values of
P3/P4 and Vg, a turbine
blade and nozzle angle
selection curve, such as
Figure 22, can be entered for
the frame size under
consideration, to select
nozzle opening and blade
angle.
•
•
•
•
•
•
The following are calculated:
Mean tangential velocity of
blade, Um =  x Dm x RPM
where
Dm = mean diameter of
turbine wheel
Ideal gas speed at nozzle
exit Cg = [2Cpg T3 {1(P4/P3)-1/]
The turbine efficiency  t can
then be obtained from figure
23.
The available power of the
turbine Wt = m g C pg T3 [
1 – (P4 / P3 ) (-1)/  ]  t
•
•
A comparison is made between the turbine power
available and turbine power required. If available
power is greater than required power, then the
estimate of air manifold pressure, made in step 2, can
be raised and the procedure is repeated with this new
assumption. If the available power is lower, a lower
value of air manifold pressure is assumed.
This iterative process is continued till the required
turbine output matches the achieved value within a
percent or two, with the final matching to be done by
actually running the turbocharger and engine on a testbed and making final adjustments by changing
compressor diffuser vanes and turbine nozzle rings.
Cleaning Turbochargers in operation
• Periodic cleaning reduces or even prevents
contamination, allowing significantly longer intervals
between overhauls. The proposed cleaning method,
carried out periodically, will prevent a thick layer of dirt
from forming. A thick layer of dirt can cause a drop in
efficiency and increased unbalance on the compressor
side of the turbocharger, which could influence the
lifetime of the bearings.The compressor wheel of the
turbocharger can be cleaned during operation by
spraying water into the air inlet casing. The dirt layer is
removed by the impact of the injected water. Since the
liquid does not act as a solvent there is no need to add
chemicals. The use of saltwater is not allowed, as this
would cause corrosion of the aluminium compressor
wheel and the engine. Water is injected from a water
vessel that holds the required quantity of water.
Procedure
•
The best results are obtained by
injecting water during full-load
operation of the engine, i.e. when the
turbocharger is running at full speed.
The complete contents of the water
vessel should be injected within 4 to
10 seconds. Successful cleaning is
indicated by a change in the charge air
or scavenging pressure, and in most
cases by a drop in the exhaust gas
temperature. If cleaning has not
produced the desired results, it can be
repeated after 1 0 minutes. The
interval between compressor cleanings
will depend on the condition of the
turbocharger suction air. It can vary
from 1 to 3 days of operation. If a very
thick layer has built up and it cannot
be removed using the method
described, it will be necessary to
dismantle the turbocharger in order to
clean the compressor side. Since the
dirt layer is removed by the kinetic
energy of the water droplets, the
engine has to be run at full load.
Cleaning the Turbine
The combustion of heavy fuel in diesel engines causes
fouling of the turbine blades and nozzle ring. The result
of this fouling is reduced turbine efficiency and engine
performance as well as an increase in the exhaust gas
temperature, Experience has shown that the
contamination on the turbine side can be reduced by
regular cleaning in operation, and that such cleaning
allows longer intervals between turbocharger
overhauls. Some of the deposits have their origin in
soot, molten ash, scale and unburned oil, partially
burnt fuel and sodium vanadyl-vanadat. Investigations
have shown that most of the residues are caused by
the calcium in the lube oil reacting with the sulfur from
the fuel to form calcium sulfate during the combustion
process.
•
The quantity of the deposits depends on
the quality of the combustion, the fuel
used, and the lube oil consumption. The
frequency with which cleaning has to be
carried out depends on the extent of the
contamination on the turbine side. Two
cleaning methods exist:
1. Wet cleaning (water injection)
2. dry cleaning (solid particle injection)
Procedure for wet cleaning
1. The exhaust gas
temperature before the
turbine should be in the
range of 200 to 4300 C
2. The boost pressure should
be above 0.5 bar to
prevent water entering the
oil chamber on the turbine
side.
3. The quantity of injected
water will depend on the
exhaust gas temperature,
water pressure, size of the
turbo-charger and number
of gas inlets.
• Water should be injected for 5 to 10 minutes.
Check if the water has entered the turbine parts
by opening the drain of the gas outlet casing.
Water flowing out provides assurance that
enough water has passed the nozzle ring and
the turbine blades. The interval between turbine
cleanings will depend on the combustion, the
fuel used and the fuel oil consumption. It can
vary from 1 to 20 days of operation.
• Principle
• The dirt layer on the turbine components is
removed by thermal shock rather than the kinetic
energy exerted by the water droplets.
Procedure for dry cleaning
•
•
•
•
The exhaust gas temperature
before the turbine should not
exceed 5800 C.
Dry cleaning has to be carried out
more often than water cleaning as
it is only possible to remove thin
layers of deposits. A cleaning
interval of 1 to 2 days is
recommended.
To ensure effective mechanical
cleaning, granulated dry cleaning
media are best injected into the
turbine at a high turbocharger
speed.
The quantity needed will vary from
0.2 l to 3 l, depending on the size
of the turbocharger.
• Experience has shown that the best results are
achieved with crushed nut-shell or granulate.
• Principle
• The layer of deposits on the turbine components
is removed by the kinetic energy of the granulate
causing it to act as an abrasive. Experience has
shown a combination of the two to be very
effective, especially in the case of 2-stroke
engines.