Liquid piston engines by Gupta, Aman Narayan, Sunny Sharma, Shubham

Combustion Based Noise 195Transfer function

0.5
0.4
0.3
0.2
0.10 2k 4k 6k 8k 10k

F (Hz)
Figure 8.16 Transfer function obtained by explosive charge.

Different functions for various designs of combustion chambers and dif-
ferent explosion charges were compared.

All the above discussed methods use expensive and time consuming
methodology to analyze the transfer function of combustion noise, conse-
quently an alternative method of analysis has been studied which evolves
use of which include Cepstrum analysis.

Cepstrum analysis is an important method of signal processing which
has wide applications in source separation [41] ]. This methodology has
also been used for psychoacoustic analysis of noise emissions from S.I.
engine [42]. Acoustic emissions from engine have been used to reconstruct
in cylinder pressure using Cepstrum analysis [43]. This methodology has
proved effective an way for fault detection in gears [44] and condition
monitoring of engines [45].

Mathematically it can be defined as spectrum of logarithmic power
spectrum [41]. i.e.

Ca(q) = |F[log Gx(f)]| (8.15)

where q is frequency in millisecond & F as well as Gx denotes the Fourier
transformation of function.

Since auto power spectrum density function is even, both its inverse
Fourier transformations & Fourier transformations are equal. i.e.

Cx(q) = |F[log Gx(f)]| =F−1[log Gx(f)] (8.16)

when a noise source x(t) reaches measuring point as a output signal y(t)
after passing through a system h(t) the system can be represented by
equation:

y(t) x(t) h(t) x( )h(t )dt (8.17)

196 Liquid Piston Engines

101

2000 RPM
100 1600 RPM

Amplitude 10–1

10–2

10–3 102 103
101 Frequency-Hz

Figure 8.17 Structural attenuation function (Motored).

Taking Fourier transformation we have:

Gy(f) = Gx(f) Gh(f) (8.18)

Taking logarithm and Fourier transformations on both sides we have:

log(Gy(f)) = log(Gx(f)) + log(Gh(f)) (8.19)

F[log(Gy(f))] = F[log(Gx(f))]+ F[log(Gh(f))] (8.20)

Or

F-1[log(Gy(f))] = F–1[log(Gx(f))]+F–1[log(Gh(f))] (8.21)

Cy(q) = Cx(q) + Ch(q) (8.22)

Structural response function for motored condition was evaluated tak-
ing acoustic emissions as output signal and in cylinder pressure as input.
For the case of firing conditions, the rate of heat release was taken as input
parameter. Figures 8.17–8.19 shows the plots of transfer function obtained
by cepstrum analysis method.

It is clear from plots that engine noise transfer function for various
testing conditions show same trends and the engine shows a higher value
of structural attenuation above 1 kHz range. At low frequency ranges,
engine parts have high rigidity and radiation efficiency is very low. In

Amplitude Combustion Based Noise 197

101
100% load

100 50% load
10–1

10–2 102 103
101 Frequency

Figure 8.18 Structural attenuation function (1600 RPM).

101

Amplitude 100 100% load
50% load

10–1

10–2 102 103
101 Frequency

Figure 8.19 Structural attenuation function (2000 RPM).

mid frequency ranges longitude mode of vibrations in piston, connect-
ing rod and crank shaft dominates which gradually increases the struc-
tural response of engine. In high frequency ranges the radiation efficiency
increases due to Cast Iron parts in engine head. Several engines use same
materials for making of various parts, hence engines of different make but
having same size show same variations in structural response function.
When compared with structural attenuation function of AVL noise meter,
the curve obtained by Cepstrum analysis shows same variations in low
frequency range with monotonic rise up to 50 Hz followed by a step fall.
Significant differences can be seen in high frequency ranges above 1 kHz.
These variations may be attributed to differences in the design of the cylin-
der head, engine block and covers which also play a vital role in determina-
tion of structural attenuation function.

198 Liquid Piston Engines

103

Combustion noise-dB 102

101 50% load
100% load

100 102 103
101 Frequency

Figure 8.20 Combustion Noise – 1600 RPM.

103

Combustion noise-dB 102

101 50% load
100% load

100 102 103
101 Frequency

Figure 8.21 Combustion Noise – 2000 RPM.

Further neglected flow induced noise, the overall noise emissions (ON)
from engine can be expressed as sum of direct combustion noise (CN) and
mechanical noise (speed dependent). i.e.

ON = CN(H1) + MN (8.23)

where H1 is structural attenuation factor of combustion noise
Assuming that mechanical noise levels (motored conditions) do not

change significantly, the combustion noise levels for the given testing con-
ditions were evaluated using transfer functions previously described as
seen in Figure 8.20, 8.21.

From these plots it clear that operational speed of engine has significant
effect on combustion noise levels, however increase of engine load caused

Combustion Based Noise 199

SOI map

Speed SOI Engine
load Vibration
Target
MBF50 Target
MBF50

PI controller

Actual Estimated
MBF50 MBF50

Figure 8.22 Use of vibration signals as a feedback for estimation of MBF50.

an increase as fuel was injected closer to TDC position. These plots are
characterized by peaks in high frequency ranges which may be attributed
to resonance of engine structure.

8.13 Summary

Value of MBF50 affects the cycle efficiency, emissions peak pressure and
temperature achieved in an engine cycle. Hence it can be used as a feed-
back parameter for a closed loop control system to device optimal timings
of combustion process as depicted in Figure 8.22 [46].

In this part of work, vibration and noise signals of diesel engine were
analyzed using Cepstrum method. As evident from these plots the trends
in transfer function remains same in spite of variations in engine running
conditions. This function was further used to calculate combustion noise
levels from engine. The value of this function is also dependent upon com-
bustion chamber resonance frequencies, hence temperature variations
inside chamber also need to be taken into account for analyzing variations
in structural attenuation factor.

Further suitable changes in engine structure have shown a reduction of
10 d B up to reduction in engine block vibrations [47].

References and Bibliography

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springer, 2009.

2. Rakopoulos, Giakoumis, “Experimental study of combustion noise radiation
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2011.

200 Liquid Piston Engines

3. Anderton, D., Baker, J., “Influence of operating cycle on noise of diesel engine”,
SAE paper no. 730241, 1973.

4. Carlucci, P., Ficarrela, A., Laforgia, D., “Study of the influence of the injec-
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14. Arnone, L., Manelli, S., Chiatti, G., and Chiavola, O., “In-Cylinder Pressure
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15. Arnone, L., Manelli, S., Chiatti, G., and Chiavola, O., “Engine Block Vibration
Measures for Time detection of Diesel Combustion Phases, SAE Technical
Paper 2009-24-0035, 2009. doi:10.4271/2009-24-0035.

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01-0168, 2010. doi:10.4271/2010-01-0168. doi:10.4271/2009-24-0035

17. Arnone, L., Boni, M., Manelli, S., Chiavola, O., Conforto, S., Recco, E., “Diesel
engine combustion monitoring through engine block vibration signal analy-
sis”, SAE Technical Paper 2009-01-0765, 2010. doi:10.4271/2009-01-0765

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18. Chiavola, O., Chiatti, G., Recco, E., “Accelerometer measurements to
optimize injection strategy”, SAE Technical Paper 2012-01-1341, 2012.
doi:10.4271/2012-01-1341

19. Chiatti, G., Recco, E., Chiavola, O., “Vibration processing to optimize pres-
sure development in CR diesel engine”, SAE Technical Paper 2011-01-1560,
2011. doi:10.4271/2011-01-1560

20. Gang, Sheng., “Vehicle Noise, Vibration, and Sound Quality”, SAE
International, 2012. ISBN 978-0-7680-7513-7

21. Anderton, D., “Noise source identification techniques”, ISVR course notes,
2003.

22. Anderton, D., “Relation between combustion system and noise”, SAE paper
no. 790270, 1979.

23. Anderton, D., “Basic origins of automotive noise”, ISVR course notes, 1990.
24. Schaberg, P., Priede, T., Dutkiewicz, R., “Effects of a rapid pressure rise on

engine vibration and noise”, SAE paper no 900013, 1990.
25. Priede, T., Grover, E., Anderton, D., “Combustion induced noise in diesel

engines”, Proceedings of diesel engines users association congress, London, 1968.
26. Russell, M., “Reduction of noise emissions form diesel engine surface”, SAE

paper no 720135, 1972.
27. Russell, M., Haworth, R., “Combustion noise from high speed direct injection

diesel engines”, SAE paper no 850973, 1985.
28. Russell, M., Cavanagh, E., “Establishing a target for control of diesel combus-

tion noise”, SAE paper no 790271, 1979.
29. Lyn, W., “Calculation of the effect of rate of heat release on the shape of cylin-

der pressure and cycle efficiency”, proceedings of I mech conference on auto-
mobiles, pp. 34–62, 1961.
30. Wolschendorf, J., Durnholz, M., and Schmillen, k., “The IDI diesel engine and
its combustion noise variations, SAE paper no 910228, 1991.
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ment in direct injection diesel engines by means of in cylinder pressure compo-
nents”, Measurements science and technology, Vol 18, pp. 2131–2142, 2007.
32. Priede, T., Grover, E., “Noise from industrial diesel engines”, proceedings of
symposium on noise from power plant equipment, Southampton, 1966.
33. Torregrosa, A., Broatch, A., Maratin, J., Monelletta, L., “Combustion level
assessment in direct injection diesel engines by means of in cylinder pressure
components”, Measurements science and technology, Vol 18, pp. 2131–2142,
2007.
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open chamber diesel engines”, JASA, Vol 65, pp. 1474–1479, 1979.
35. Payri, F., Broatch, A., Tormos, B., and Marant, V., “New methodology for in
cylinder pressure analysis on DI diesel engines-application to combustion
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diesel engines”, Phd Thesis, University polytechnic velencia, 2010.

202 Liquid Piston Engines

37. Nguyen, T., Kai, Y., Miami, M., “Study on combustion noise from a run-
ning diesel engine based on transient combustion noise generation model”,
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38. Austen., Priede, “Origins of diesel engine noise proceedings of symposium on
engine noise and noise suppression”, IMech, London, pp. 19–32, 1985.

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Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

9

Effects of Turbo Charging
in S.I. Engines

9.1 Abstract

We all know that one of the prime objectives of any innovation is to achieve maxi-
mum output with minimum input. Automobiles also have innovations that aim
at achieving maximum mechanical efficiency & fuel-economy, both simultane-
ously. The following project-paper focuses on one of the most important topics of
present day automotive industry “Turbo charging in S.I engines”.

The principle objective of turbo charging is to increase the power output per
volume and cost of engine. A fact that a turbocharger increases the mass of air in
the cylinder and consequently allows more fuels to be burnt, improves the volu-
metric efficiency of the engine and simultaneously improves engine efficiency by
a small but worthwhile amount.

Turbochargers are have commonly been used on diesel engines for many years.
In contrast only a few petrol engines have been turbocharged until recently and
it is unlikely that a large fraction of the world’s petrol engine will be so equipped.
Most of the vehicles in the Indian automobile market continue to utilize tur-
bocharged diesel engines compared to gasoline engines. However, experts are of
the opinion that turbo growth, in the future would not be confined to diesel sector

203

204 Liquid Piston Engines

alone. There would be a tremendous growth in demand of Gasoline downsizing,
in next few years.

The whole industry is drifting away from naturally aspirated engines & com-
panies like Mercedes & Volkswagen are planning to come up with commercial
turbo-charged, downsized gasoline engines. Thus according to experts if automo-
tive sector, as a whole, is growing at 2%, then turbocharger industry is growing
at about 10%, with maximum growth rates coming from turbos incorporated
in gasoline engines.

The present project aims at analyzing the various benefits associated with
“Turbo charging in SI engines” and designing it for future automotive applications.

9.2 Fundamentals

A. An I.C engine is a device where-in combustion of fuel takes place within
the cylinder & corresponding hot gases are used to drive the output shaft.
Every such engine consists of an engine cycle comprising of 4 stages:

a. Suction
b. Compression
c. Expansion
d. Exhaust

There are two types of engines which are classified on the basis of their
combustion mechanism as follows:

A. SPARK IGNITION ENGINE: refers to the category of engine that
makes use of gasoline as fuel. It can be of two types i.e. four stroke or two
stroke, depending on whether the one engine cycle is completed in four
steps or two steps, respectively .The specialty of this engine is we make use
of a carburetor to mix the fuel and air as per the stoichiometric ratio,
which is then supplied through pumps into the cylinder (intake) at high
pressure. This mixture is then compressed and combusted with in the
cylinder using a spark plug which supplies the spark to ignite the mixture
(as per specific firing order and after a given period of time).After this the
hot gases generated as a result of fuel ignition expand (thereby rotating
crankshaft and producing output power) and finally get exhausted from
the cylinder.

B. COMPRESSION IGNITION ENGINE: This category of engines make
use of diesel as a fuel and are employed both in heavy as well as small vehi-
cles i.e. four stroke and two stroke, respectively. The specially of engine

Effects of Turbo Charging in S.I. Engines 205

DE Turbo charger Diesel engine

Air cooler

AC generator

A B
Figure 9.1 Turbocharged C.I engine.
C

Table 9.1 Comparison: Spark-ignition and diesel engine.

Engine type Compression Relative Manifold Max. Part-load
SI Engine ratio F/A absolute Bmep efficiency
CI Engine pressure (Bar)
9.5–11.5 1.0 (%)
(Bar) 12
16–20 0.2–0.7 20–25
0.5 18
30 (aprox)
1 bar

is it only makes use of fuel injector and atomizer to atomize and spray
the fuel into the cylinder that contains highly compressed air at very high
temperatures. Its advantages include high compression ratios resulting in
high engine net output lesser knock tendency, cheaper fuel.

9.3 Turbochargers

A turbocharger basically consists of a compressor and a turbine coupled
on a common shaft. The exhaust gases from the engine are directed by the
turbine inlet casing on to the blades of the turbine and subsequently dis-
charged to atmosphere through a turbine outlet casing. The exhaust gases
are utilized in the turbine to drive the compressor, which compresses the
air and directs it to the engine induction manifold, to supply the engine
cylinders with air of higher density than is available to a naturally aspirated
engine. The higher value of air-pressure achieved using a turbo-unit is

206 Liquid Piston Engines

called Boost-pressure. There exist a number of different types of compres-
sors and turbines, but few of these are ideally suitable to form the basis of
an exhaust gas driven supercharging system. The combination of a single
stage centrifugal compressor and a single stage axial flow or radial flow tur-
bine is almost universally used in turbochargers. The former type is used
for medium and large size engines, while the latter type is used for small
engines of automotive type.

Advantages:

1. Increase the fuel volumetric efficiency by about 30% to 40%.
2. Increase the number of power stroke i.e. increase the final

output.

Disadvantages:

1. A disadvantage of turbo charger is its resistance to high tem-
perature at high load, which imposes to increase equivalent
ratio (enrichment).

2. The main disadvantage in the turbine inertia and the cor-
responding long response time needed to obtain the super-
charging pressure.

9.4 Turbocharging in Diesel Engines

Turbochargers are now widely used for truck engines, their output
approaching passenger car engine values of 45 KW at one end of the power
range and 600 KW at the other for special purpose vehicles. The use of
turbochargers on vehicle engine is relatively recent, but the rapid increase
in power output demanded by the use of larger trucks and minimum
power to weight ratio legislation, has speeded up their introduction even
in passenger cars.

The factors that limit turbocharged diesel engine performance are com-
pletely different to those that limit turbocharged SI engines. The output of
naturally aspirated diesel engines is limited by the maximum tolerable
smoke emission levels, which occurs at relative A/F ratio values of about
0.7 to 0.8. It is usually constrained by stress levels in critical mechanical
components. This limits the maximum cylinder pressure which can be tol-
erated under continuous operations, though the thermal loading of critical
components can become limiting too.

Effects of Turbo Charging in S.I. Engines 207

Turbo charging in CI engines is free from knocking and self - ignition
troubles, enabling it to utilize high compression ratios, multi-port direct
fuel injections and high cetane number and cheaper exhaust control
systems which make it easier and more economical to design.

9.5 Turbocharging of Gasoline Engines

In SI engines, fuel and air are pre-mixed before the air enters the cylinder
of petrol engine. Whether a carburetor or manifold petrol injection sys-
tem is used, cylinder comprises of homogeneous air and fuel mixture, the
proportion of fuel being carefully controlled. The homogeneous mixture is
ignited by the spark plug. Unlike diesel engine the rate at which combus-
tion proceeds is governed by heat and mass transfer from an area that is
burning to an area that is not, and temperature increase due to continued
compression thus flame advances across the combustion chamber, from
the spark plug until all the fuel is burned.

Self ignition is avoided by low compression ratios, enough to hold the
temperature of mixture below the self ignition point of the fuel, and by
using a fuel having high self ignition temperature. The rate at which the
flame progresses is governed by local turbulence, heat transfer between
burning and unburned region, compression heating of the unburned
gas due to piston motion and expansion of burning mixture, air/fuel
ratio and heat transfer to the surrounding walls. Since the unburnt gas
that is removed from the advancing flame front is heated by compression

Air filter

Intercooler Air flow meter
Compressor
Throttle Turbine shaft
Turbine
Intake Engine Waste
manifold gate Catalyst

Exhaust
manifold

Figure 9.2 Sketch of a turbocharged SI-engine.

208 Liquid Piston Engines

and to some extent by radiation etc. This gas can reach its self ignition
temperature before the flame front arrives, thus increasing the chances of
knocking in the end gas region. This extremely rapid combustion gener-
ates a high rate of pressure rise in the cylinder, the impulse of force causing
the bearing to knock, generally referred to as detonation.

9.6 Turbocharging

Turbocharging applied further to S.I engines. As we know, turbo charging
is one of the most economic and effective methods used to improve both
the volumetric efficiency and thereby improve the overall efficiency of any
engine.

However, this principle is not very widely used in case of SI engines pri-
marily due to the difference in combustion systems of the two engines (i.e.
SI and CI). This difference has been explained in the previous section. The
important point is that any measures that might increase the temperature
of the mixture towards the end of the compression strokes are undesirable.
Unfortunately, turbo charging does just this. By raising the inlet manifold
pressure and temperature, the pressure and temperature of the mixture in
the cylinder will be raised throughout the compression stroke. Hence, it is
possible only to mildly supercharge the engine without inducing knock or
self-ignition. Thus a certain margin clear of knock must be maintained by
incorporating the following:

1. Low compression ratio
2. Retarded ignition timing or charged air cooling
3. Usage of high octane fuel, having high self- ignition tem-

perature etc. to offset the effect of temperature rise in the
compressor.

9.7 Components of Turbocharged SI Engines

A. Inlet and Exhaust Manifolds

The design of the inlet system may be substantially different from that used
in naturally aspirated engine; particularly if carburetor is placed before
the compressor. The primary objectives will be to minimize the volume
between carburetor and cylinder head (to improve response), encour-
age good fuel droplet break-up and mixing, avoid fuel condensation,

Table 9.2 Comparison of naturaly aspirated & turbo-charged BMP C.I engine (15.9 Litres at ambient temperature 28 °C & humidity
61%) Ref:V.R.D.E. MAGAZINE.

S.No Engine Load (N-M) Consumption Power in KW % % % Reduction Effects of Turbo Charging in S.I. Engines 209
R.P.M Of 200 G of fuel Increase Increase in fuel
1. N.A T.CH N.A T.CH in load
2. 2600 627 765 (Sec) 232 279 in consumption
3. 2400 658 785 224 264 22%
4. 2200 700 843 N.A T.CH 219 260 Power 19% 5%
200 731 849 208 236 20% 13%
13.6 12.9 20% 16% 14%
15.3 13.5 17% 32%
15.8 13.8 18%
17.3 13.1 13%

210 Liquid Piston Engines

encourage even mixture distribution between cylinders, eliminate swirl-
ing air movement and keep the temperature of the mixture down. These
objectives are some times achieved by accepting a pressure loss that would
not be tolerated on a naturally aspirated engine. The reason is due to the
most embarrassingly large amount of energy available in the exhaust system
at full speed, enabling boost pressures far above the knock-limited value to
be generated. Some flow loss, at full throttle and speed, reduces boost pres-
sure, loss being low when boost pressure and hence mass flow rate are also
small. A perforated plate can be used at the entry to the inlet manifold to
reduce the swirl and aid mixture distribution, but this method is not to be
encouraged since restrictor reduces pressure, but not temperature.

Although the potential boost available from the turbocharger is exces-
sive at full engine speed and throttle, the opposite is so at low speeds.
Thus the exhaust manifold is to be designed to ensure the maximum
utilization of exhaust gas energy at relatively low engines speeds. Thus
the pulse turbo charging system should be adopted, with short, narrow
exhaust pipes and if possible no more than 3 cylinders connected to each
turbine entry.

Long pipes have been used to allow cooling of the exhaust gas before
it reaches the turbine, but pressure wave action can cause problems. The
energy available for expansion through the turbine is reduced at low speeds,
engine compartment temperature increases and turbocharger response
suffers. Usually a large-bore exhaust pipe will be used from the turbine
exhaust onwards, together with low-loss silencer. A small bore system will
create flow restriction and back pressure at the turbine exit. Although this
need not be a problem because turbocharger can be matched accord-
ingly, it can be dangerous if the exhaust system develops a major leak.
The back pressure will be reduced and turbocharger work will increase,
possibly resulting excessive boost pressure, combustion knock and engine
damage.

B. Turbocharger Boost Pressure Control System

The need for boost pressure control system in a turbocharger unit includes:

1. Petrol engine works over a wide speed range, typically 5000
revolutions per minute.

2. Air/fuel mixture must be keep relatively closed to the stoi-
chiometric value rather than varying over a very wide range.

3. Boost pressure must be limited to avoid knock. Problem
involved in the matching the turbocharger over the normal

Effects of Turbo Charging in S.I. Engines 211

speed range becomes more severe over the wider speed
range of petrol engine.

The various ways by which these objectives achieved are:

C. Matching the Turbo Charger for the Desired Maximum
Boost Pressure

One of the simplest methods to achieve specified boost pressure at
maximum speed and load is to match the turbo charger accordingly.
The boost pressure can be limited by fitting a large turbine; however
the turbine area will be governed by mass flow through it at full power.
Since the petrol engine operates over a very wide speed range, mass
flow will be very much less at low speeds. Under these conditions, the
turbine area will be relatively large and hence the energy available for
expansion will be correspondingly small resulting in little or no boost
pressure being developed. Aggravating the situation will be low turbo
charger efficiency due to operation far from the design point conditions
of the compressor.

The overall pressure will be a boost pressure and engine torque curve
rising rapidly at high speed, with low boost at low speed. Such a curve,
with no torque back up is totally unacceptable for the automotive applica-
tion since continual gear changing is required.

D. Exhaust Waste-Gate systems

It is one of the most extensively used pressure control system. It consists
of a valve allowing exhaust gas to by pass the turbine. It is an excellent
method to control boost pressure, since the exhaust gas energy of the turbo
charged engine is excessive at full speed and load. It consists of a diaphragm

Inlet Outlet
Throttle Wastegate

Engine

Figure 9.3 Waste-Gate Control System.

212 Liquid Piston Engines

on which the boost pressure acts, opening the waste gate when the boost
pressure reaches the pre –determined value.

The various advantages of this system are:

1. Since not all the exhaust gas passes through the turbine, and
no more than air requirement of the engine passes through
the compressor, then a smaller turbo charger can be used.
This small turbo charger is best able to provide sufficient
boost at low speed (when waste gate is closed), and reduces
turbo charger lag due to its low inertia, especially if the
waste-gate is closed during acceleration.

2. With careful design, the boost curve can be tailored to pro-
duce an optimum torque cure within the constraints of
knock limit.

E. Selection of Fuel - Supply Systems

There are two types of combustion systems used in SI engines, classified on
the basis of fuel- preparation systems

1. Spark plug based system: they make use of carburetors to prepare the
fuel/air mixture in the correct stoichiometric ratio. There can be two types
of such systems based on whether turbocharger is placed before or after
the carburetor, which is called sucking-in or blow-through type systems
respectively.

The various disadvantages associated with the two systems are:

a. When turbocharger is placed after the carburetor it makes it
more difficult to provide the require the boost to the air/fuel
mixture from the carburetor, as the fuel particles are more
difficult to compress due to their higher density. Apart from
this there might also be flow-separation problems inside the
compressor.

Air cleaner

Carburetor Compressor

Turbine

Exhaust pipe

Exhaust gas

Figure 9.4 Carburetor -based turbocharger.

Effects of Turbo Charging in S.I. Engines 213

b. When turbocharger is placed before the carburetor, it has to
handle only air and hence more preferred. However this also
results in higher boost pressure losses inside the carburetor.

It might also result in excessive rise in fuel / air temperature and pressure
inside the carburetor, which both affects the carburetor and also increases
the end gas temperature in the combustion camber causing knocking and
detonation.

2. Direct-injection systems: they make use of injection nozzles to spray
the fuel/air mixture directly into the cylinder at high pressure, just dur-
ing before the compression stroke begins, such that a homogenous charge
mixture is formed which gets ignited using the spark-plug & undergoes
uniform combustion.

They are more commonly used these days, since they can be more easily
controlled for different operating conditions (like part-load, full-load etc.)
& can be used in conjunction with turbo-chargers to achieve maximum
efficiency with minimum lag at low-speed conditions.

9.8 Intercooler

When intake air is compressed by a turbocharger it is also heated, even
more so than when supercharging due to the turbo being heated by the
exhaust. Hot intake air is not good for power and will increase the chance
of detonation. An intercooler reduces the intake temperature by pushing
the air through a heat exchanger (much like a small radiator) that absorbs
some of the heat out of the charge. With less heat, less boost pressure is
needed to get the desired power and decrease the chance of detonation.
Anything that reduces the intake temperature is a big plus in a super-
charged engine.

9.9 Designing of Turbocharger

Designing a naturally aspirated SI engine of following specifications is
compared with a turbocharged SI engine of same specifications.

Specifications of a Naturally aspirated SI Engine:

D = 90 mm
L = 125 mm

214 Liquid Piston Engines

N = 2500 RPM
Rp = 7:1
K = Number of cylinders = 3

= Density of gasoline = 800 kg/m3
c.v. = 44000KJ/Kg
Maximum speed = 144km/hr
Fuel consumed = 14km/litre
Ambient pressure = 1 bar= P1
Ambient temperature = 27 °C = 300 K = T1

Calculations:

Rate of fuel consumption = mf= Maximum speed/fuel
consumed = 10.2 litre/hr

Vs = π/4 × D2 × L × N/2 × K = 2.985 m3/minute
Specific consumption of fuel = mf ˙ = mf = 8.22 kg/hr

Assuming complete combustion of gasoline, the Combustion of gaso-
line is analyzed to find the mass flow rate of air needed to burn fuel and the
mass flow rate of the exhaust formed. Thus by knowing the swept volume
the volumetric efficiency is calculated.

2C8H18+25O2 16CO2+18H2O

(GASOLINE) (AIR)

Oxygen needed for combustion of 1 kg of fuel

= 0.85(32/12) + 0.15(8) = 3.46 kg/kg of fuel

Air needed for combustion of 1 kg of fuel = 3.46/0.233 = 14.87
Hence air flow rate = m˙ a=14.87(8.22) = 122.23 kg/hr = 0.033 kg/s
Now P1V1 = m˙ a RT1
V˙air = V1˙=V2˙=. 028m3/s = 1.68 m3/minute

v (Vair /Vs ) 100 = 56%

Hence Oxygen flow rate = m˙O2
= (0.23)122.23 kg/hr = 28 kg/hr = 0.46 kg/minute

Carbon dioxide formation rate= m˙CO2=8.22(704/228) =25.38kg/hr

Effects of Turbo Charging in S.I. Engines 215

Maximum volume
Displacement volume
Clearance volume
3

Absolute pressure 2
4

0 1
0 Atmospheric pressure

0 Volume

Figure 9.5 P-V Plot of a Naturally aspirated SI Engine.

HNWoeanwtecrBefymofri˙mnedxaihtniaoguntshrtea=tpemr=es˙mHs2u˙OHr2e+Oa=mn8d˙.C2tO2e2(m3=2p40e/.r02a12tuk8r)ge/=ss1a1t .s6ta8tkegs/1h,r2,3,4 the stroke
volume and work done per cycle (area under curve) is calculated which
leads to calculation of mean effective pressure and hence theoretical power
generated. Hence the brake specific fuel consumption and brake thermal
efficiency is calculated

P2/P1 = 7 = Rp

Hence P2 = 7 Bar

T2/T1 = 7( 1/ )

Hence T2 = 300(7)0.4/1.4 = 523K
Assuming T4 = 1000 K = Exhaust gas temperature of a S.I. Engine

T3/T4 = Rp( 1/ )

T3 = 1743 K = maximum cycle temperature
For process 2–3 we have

P3/P2 T2 = T3

216 Liquid Piston Engines

P3/P2 = 1743/523 = 3.33
P3 = 7(3.33) = 23.33 Bar
P3/P4 = 7

P4 = 3.33 Bar = Exhaust pressure of gases
Now

V˙1 = V˙4 = 0.029 m3/s

P3/P4 = (V˙1/V˙3) = (V˙1/V˙2)

V˙2 = V˙3 = 7.27 10 3 m3/s

Work done per cycle w (P3V3 P4V4 ) (P2V2 P1V1 )
0.160 J 1 1

Length of indicator curve = V˙1 V˙2 = 0.02173 m3/s

IHP (IMEP)(L)(A)(N /2)(K) 36.36 KW
60,000

49.1 HP

m = mechanical efficiency = 90% = BHP/IHP
BHP = 32.96 kW = 44.1 HP
BHP = 2 pNT/60,000

T = TORQUE = 125.89 N-m

BSFC mf /BHP 0.2492 Kg/KW-hr

bth BHP/(mf cv) 32%

Case-2 Turbocharged SI engine without intercooler

Effects of Turbo Charging in S.I. Engines 217

Compressed air flow

Engine Turbocharger
cylinder oil inlet

Compressor Turbine
wheel
Ambient
air inlet Exhaust
gas
discharge

Compressor

wheel Oil outlet Wastegate

Figure 9.6 Turbocharged SI engine.

Turbine wheel

Shroud profile

Inducer diameter Backface
Exducer diameter

Tip width

Figure 9.7 Turbine wheel.

Cb B F
Cf1
A 1 = 16° 1 Cr1
Ca1 E Turbine vane
Inlet velocity Cf2 = Cr2
triangle

Cr2

2 2 = 90°

B E

Cb

Outlet velocity

triangle

Figure 9.8 Velocity diagrams of compressor vane

TURBINE
The turbine wheel is made from a high nickel superalloy investment cast-
ing. This method produces accurate turbine blade sections and forms.
Larger units are cast individually. For smaller sizes the foundry will cast
multiple wheels using a tree configuration.

218 Liquid Piston Engines

Specification of turbine wheel: Diameter = D = 270 mm = 0.27 m,
N = 50,000 RPM

Nozzle angle = 1 = 16 , Type-Radial flow turbine

Calculations for turbine wheel:
Now the power generated by turbine is calculated by analyzing the velocity
triangles for the turbine vane.

Cb = πDN/60 = 706 m/s = AB

Assuming Turbine efficiency 85%, ρ = 0.7 = Cb/ Ca1
Analyzing the inlet velocity triangle we have

Ca1 = 1009 m/s = absolute velocity at inlet
Cf1 = Ca1 Tan 16° = 289 m/s = flow velocity at inlet
AF = Ca1 Cos 16 = 969 m/s = Cw
Let transmission efficiency = 90%
Now by calculating the power required to drive the compressor and
analyzing the compressor vane, the mass flow rate of air to compressor
inlet is found
Powergeneratedbyturbine = m˙exhaust(CW)(Cb)= 6847 W = 9.17 HP
Power supplied to compressor = Pc =0.9(6847) = 6162 W = 8.25 HP

Compressor

Compressor impellers are produced using a variant of the aluminum invest-
ment casting process. A rubber former is made to replicate the impeller
around which a casting mould is created. The rubber former can then be
extracted from the mould into which the metal is poured. Accurate blade
sections and profiles are important in achieving compressor performance.
Back face profile machining optimizes impeller stress conditions. Boring
to tight tolerance and burnishing assist balancing and fatigue resistance.
The impeller is located on the shaft assembly using a threaded nut.

Main
blade

Superback Inducer dia
profile Exducer dia

Tip Shround profile
width Spliter blade

Blade
width

Figure 9.9 Comprssor wheel.

Effects of Turbo Charging in S.I. Engines 219

Compressor housings are also made in cast aluminum (cast iron for
high-pressure applications). Various grades are used to suit the application.
Both gravity die and sand casting techniques are used. Profile machining
to match the developed compressor blade shape is important to achieve
performance consistency.
Specifications of compressor
Type-Centrifugal
Outer Diameter = D2 = 150mm
Inner Diameter = D1 = 100mm
N = RPM = 50,000, P = 1.4
σ = 0.94 = Cw2/Cb2

Calculations for compressor wheel 73%

Cb2 = pD2N/60 = 392 m/s
Cw2 = 369 m/s
Pc = m˙ compressed air × Cb2 ×Cw2
m˙ compressed air = 0.0425kg/s = 2.55kg/minute
V˙ compressed air = 2.20 m3/minute = 0.036 m3/s
Assuming lean mixture formed, A: F = 18:1

m˙fuel = 0.141 kg/minute = 8.5 kg/hr
New volumetric efficiency (Vair /Vs ) 100

Now assuming compressor isoentropic efficiency as 85%, the rise in
temperature and pressure of air by compression is calculated

T01 = 27 °C = 300K
T02 = Actual Outlet temperature of air from compressor
T02˙ = Isoentropic Outlet temperature of air from compressor
hc = 85% = (T02 T01)/ (T02 T01) = isoentropic compressor efficiency
Now Cp (T02 T01) = P (s) (Cb2)2/gc J
Hence rise in air temperature in compressor = (T02 T01) = 130K
Outlet temperature of air = T02 = 430K
(T02 T01) ( / 1) = P2/P1 = Pressure rise in compressor = 2.99

Now By finding the pressure and temperatures at states 1,2,3,4,8,7,6 the
stroke volume and work done per cycle (area under curve) is calculated
which leads to calculation of new mean effective pressure and hence theo-
retical power generated. Hence the new brake specific fuel consumption
and the new brake thermal efficiency is calculated.

220 Liquid Piston Engines

Cb2

Outlet velocity triangle

Cw2 G
DE

= 3 Cf2 =2

C2 Cr2

F

Compressor vane

C

Inlet velocity C1 = Cf1
triangle Cr1

=1

Cb1
A

Figure 9.10 Velocity diagram of compressor blade.

p pch = Charging pressure
3 pa = Atmospheric pressure

2

pch 8 Ap 4

pa 7 6 V
TDC BDC

Figure 9.11 P-V curve of a turbocharged S.I.Engine.

V˙ compressed air = V˙1 = V˙4 = 0.036 m3/s

A more dense air enters the engine cylinder and hence the compression
ratio is assumed to reduce from 7 to 5.7

Now P3/P4 = (V˙4/V3˙) = (V˙1/V˙2) = 5.7 , V˙2 = V3˙ = 0.0103 m3/s
T2/T1 = 5.7 ( 1/ )
Hence T2 = 430(5.7) 0.4/1.4 = 707K

Effects of Turbo Charging in S.I. Engines 221

P7 = P6 = 1 Bar=Ambient pressure
P1 = P8 = 2.99 Bar
Hence P2 = 5.7(2.99) = 17.043 Bar
New length of indicator curve = 0.036-0.0103 = 0.0257 m3/s
T4 = 850 = Exhaust gas temperature of turbocharged S.I. engine
T3/T4 = Rp ( 1/ )

T3 = 1397 K = New maximum cycle temperature
For process 2–3 we have
P3/P2 T2 = T3
P3/P2 = 1397/707 = 1.97

P3 = 1.97(17.043) = 33.57Bar
P3/P4 = 5.7

P4 = 5.89 Bar = Exhaust pressure of gases
New work done per cycle

W= P3V3 P4V4 P2V2 P1V1 P (Stroke volume)
1 1

0.2418 J

IMEP work done per cycle
Length of indicator curve
0.2418 / 0.0257 9.4 bar

IHP (IMEP)(L)(A)(N /2)(K) 46.76 KW = 62.68 HP
60.000

m = mechanical efficiency = 90% = BHP/IHP
BHP = IHP (0.9) =42 KW=56.41 HP
BHP = 2 pNT/60,000

T = 160 N-m

New BSFC = m˙f/BHP = 0.2428 tvkg/kw-hr
NEW bth Brake thermal efficiency BHP/(mf cv) 40%

Summary of calculations

Parameter Naturally aspirated SI engine Turbocharged SI
BSFC 0.2492kg/kw-hr engine
BHP 32.96kw
0.2428kg/kw-hr
42kw

222 Liquid Piston Engines

TORQUE 125.89 N-m 160N-m
IMEP 7.3bar 9.4bar
32% 40%
bth

9.10 Operational Problems in
Turbocharging of SI Engines

The basic efficiency of the thermodynamic cycle on which the engine
operates is largely governed by the compression ratio. Thus by reducing
the compression ratio to avoid knock, the efficiency of the basic thermo-
dynamic process is reduced. As a result it is probable, but not certain, that
the overall efficiency of the engine will suffer. However, the fall in efficiency
with low compression ratio is by no means linear.

Nearly, 25% reduction in compression ratio reduces the efficiency
by only 10%. In addition to the losses occurring in the basic thermody-
namic processes, friction in all the bearings and other mechanical debits
will reduce the power output to fly wheels. Since the reduced compression
ratio partly offsets the increase in cylinder pressure due to turbo charging,
the mechanical loads may not change significantly with turbo charging.
Thus absolute power loss due to friction will remain steady.

E.g.: Consider an SI engine with compression ratio 10:1 producing the
power output of 100KW with 46% cycle efficiency and 70% mechanical
efficiency. The potential power of 312 KW of which 143 KW arrive at the
pistons and 43 KW is then lost due to mechanical inefficiency.

If the engine is turbocharged such that 1.5 times as much mixture
is trapped in the cylinder, and compression ratio is reduced to 7.5. The
potential power becomes 468 Hp and the power delivery from the piston is
196 KW. If mechanical losses remain unaltered at 43KW, the power output
at the flywheel is 153 KW and the overall efficiency is 32.7% of the normal
engine.

(Ref: turbo charging of I.C engines, Watson & Janota)

“Thus power output has increased substantially with a small gain
in efficiency”.

Effects of Turbo Charging in S.I. Engines 223

Most important is the fact that whatever modifications are made to the
engine, the charge temperature should be kept as low as possible, so that
the compression ratio can be maintained as high as possible commen-
surate with freedom from knock. There is also a strict trade-off between
boost pressure, charge air temperature ,air/fuel ratio and fuel octane
number at the knock limit, with a fixed compression ratio and an opti-
mum ignition timing. As expected octane number has a strong influence
on the permissible boost pressure and so does charged air cooling. In
addition, richer air/ fuel ratios permit higher boost pressure.

9.11 Methods to Reduce Knock in S.I Engines

The various methods used to reduce the likelihood of knock in S.I.
Engines are:

1. Low compression ratios : to hold the compression tempera-
ture rise to an acceptable limit

2. High self ignition temperature: dependent on many factors
like fuel properties, air/ fuel ratio and pressure.

3. High octane rated fuels: have higher self ignition tempera-
ture and reduce knocking.

4. Centralized spark plug: to hold down its maximum dis-
tance from the extremity of the combustion chamber.

5. High wall surface area to gas volume ratio: to keep the end
gases relatively cooler to avoid end gas detonation

9.12 Ignition Timing and Knock

Retarding combustion reduces the temperature of the end gas by delay-
ing its heating related to the TDC piston position. Thus the cylinder
volume is increasing at the critical time, reducing the compression and
temperature rise that would occur, otherwise. Retarded ignition reduces
engine efficiency by shortening the effective expansion stroke. However,
relative to reducing compression ratio, it is more flexible technique, since
it is relatively easy to retard ignition only and the boost pressure is high
enough to induce knock.

224 Liquid Piston Engines

1 –5 0
0.9
0.8
0.7
0.6
0.5
0.4
0.3
0.2
0.1–40 –35 –30 –25 –20 –15 –10

a) [deg]

To avoid an unnecessary fuel consumption penalty with retarded tim-
ing, the ignition timing technique should only be used when turbo-
charger does develop a high boost pressure. Thus at low speeds and part
load convectional timing is retained. The simplest method of achieving
this requirement is a boost pressure retarded system built into the normal
vacuum advance diaphragm of the ignition distributor.

One undesirable feature of retarded timing increase in heat rejec-
tion to the exhaust system, since the complete combustion and expansion
process is delayed. Thus the turbine inlet temperature rises. Although the
increase is small, the very high temperature of the petrol engine exhaust
gas (upto 1000 °C) is a problem for the turbine manufacturer and can cause
oxidation of lubricating oil. Furthermore the potential power increase
obtainable by turbo charging with retarded timing alone is limited.
Higher boost pressures can be used if compression ratio is also reduced.

9.13 Charge Air Cooling

The temperature rise in the compressor and its effect on the knock can
be offset by charge air cooler. Charge air cooling to an air temperature
of 45 °C enables the knock-limited ignition timing to be advance by 10°.
However the low boost pressure of the turbocharged SI engine means that
the temperature difference available between ambient air and compressed
air is small.

A large air cooler is therefore required to achieve a major reduction
in temperature and low pressure loss. However, it is doubtful whether the
extra cost, complexity and volume are warranted in passenger car appli-
cations, other than very expensive sports cars. An additional disadvantage
is deterioration in engine response due to the increase in total inlet
manifold volume and pressure loss in cooler itself.

Effects of Turbo Charging in S.I. Engines 225

9.14 Downsizing of SI Engines

Downsizing refers to the reduction of engine swept volume without
compromising on the potential engine output. It offers enhanced engine
efficiency, fuel economy and potential to meet future emission standards.
However in order to realize these benefits a number of issues faced by con-
ventional port injected down-sized engines must be resolved.

1. Compression ratio: as previously mentioned, it must be
reduced on conventional boosted engines to control knock.

2. Low speed torque: due to air flow and octane requirement
constrains, steady state torque is reduced compared to a nat-
urally aspirated engine of equivalent peak torque and power.
This adversely affects pull away from rest and performance
feel for tip-in

3. Manoeuvres at low engine speed
4. Transient response: although greatly improved with mod-

ern boosting systems, turbocharger lag is still an issue for
down sized engine. Turbocharger lag compounded by the
lack of steady-state torque is a particular issue for low speed
performance feel.
5. Economics: any downsizing technology package must be
economically justified in terms of a cost benefit analysis
compared to other technology packages.

9.15 Techniques Associated with Turbo Charging
of SI Engines Boosting Systems

A number of advanced boosting technologies aimed at addressing the
steady state low- speed torque and transient response issues are currently
in advanced engineering stages as follows:

1. Gasoline direct injection
2. Variable compression ratio
3. Variable geomety turbine
4. Exhaust gas recirculation

a. Gasoline direct injection (GDI): Direct injection is a key technol-
ogy for improving fuel economy of down sized engines. As well as the
small gain in volumetric efficiency, direct injection allows an increase in

226 Liquid Piston Engines

compression ratio of 1 to 1.5 compared with an equivalent port injected
engine as a result of charge cooling.

Swirl control
valve

Spark plug

High-pressure Swirl air
fuel injector motion

Hollow cone
spray

Piston bowl

Direct injection also allows more freedom on choice of valve overlap
for boosted engines, as starting to inject after exhaust valve closure can
prevent fuel loss. This allows improved scavenging under some conditions,
which reduces charge temperature and octane requirement.

The ability to have multiple injection events in one combustion cycle
allows both increased exhaust temperature from cold start for improve cat-
alyst light-off performance and reduced full-load octane requirement. The
increase in exhaust temperature can help to over come the increase exhaust
system thermal inertia with turbocharger. Thus the injection strategy
must be optimized for each combustion system.

b. Variable compression ratio (VCR): In principle, variable compressible
ratio is an attractive approach for down sized boosted engines. Low com-
pression ratio would be used at full-load to control knock, allowing very
high BMEP to be achieved. As load is reduced; compression ratio will be
optimized for best economy. Two variable compression approaches have
been proposed for downsized boosted engines:

a. In the other, a fixed high geometric compression ratio is used
but the effective compression ratio is varied by late closing of
inlet valve. This approach is termed as “Miller’s cycle”.

b. In one approach, the geometric compression ratio is varied
by mechanical means.

In practice, variable geometry compression ratio systems have a num-
ber of drawbacks. As well as many design and development issues such

Effects of Turbo Charging in S.I. Engines 227

as packaging, friction, transient response, durability and cost to be
resolved. The use of low compression ratio at high load will result in poor
high load fuel consumption. The resulting high full-load exhaust temper-
atures will require earlier and increased use of over-fuelling for component
protection.

The principle is to use high geometric compression ratio, (around 14:1),
giving good part load economy, and to control knock by reducing by
effective in-cylinder compression ratio by late intake valve closing. As this
reduces volumetric efficiency, higher boost pressure are required to achieve
same BMEP. However as more of the compression is done in the external
compressor and the intercooler can be used, the in-cylinder charged tem-
perature at ignition will be lower than for a conventional boosted engine,
even with a same trapped mass. This will benefit the octane requirement
while VCT would optimize cam-phasing at each speed.

E.g. Miller cycle approach was used in production by Mazda in 1980’s.
This engine featured fixed late inlet valve closure and supercharged using
lysholm compressor and inter cooled. It has the compression ratio of about
10:1. More recently, this approach has been proposed in combination with
VCT (variable cam-shaft timing) as a solution of downsized engines.

c. Variable turbine geometry (VGT): By altering the angle of the tur-
bine inlet nozzle both the effective area of the turbine (and hence energy
availability) and efficiency characteristic alter. By opening the nozzle at
full engine speed and closing them at low speed, exhaust gas energy can

Exhaust gas Speed Boost-air
sensor outlet
Sliding
nozzle Air
ring inlet
Compressor
Turbine

Pneumatic
control
cylinder

Control air

Figure 9.13 Turbocharger with VGT.

228 Liquid Piston Engines 5 13
8 2
13
2 76

4

(a) (b)

Figure 9.14 Exhaust gases recirculation system assembly for naturally aspirated (a) and
turbocharged (b) - System components: 1 humidity’ separator; 2 booster; 3 EGR valve;
4 single point injection; 5 heat exchanger: 6 compressor: 7 single point injection and
equalization box: S turbine.

adjusted to suit low and medium-speed performance while preventing the
turbo charger from over-speeding. The difficulty is one of engineering a
cheap & reliable variable geometry turbine together with associated con-
trol system.

As compression ratio increases, modern gasoline engines have exhaust
temperature higher and higher. Experts estimated it could exceed 1000°C
in the foreseeing future. Perhaps this is why VTG technology for gasoline
engines never went into mass production.

In terms of single-stage charging units, turbochargers with variable
turbine geometry are primarily used for passenger car diesel engines (in
addition to modern and economical turbochargers with boost pressure
control valves).

d. Exhaust gas recirculation(EGR):It is a technique used to cool the
flame passing through combustion chamber offsetting knock limit,
enabling optimum ignition event and more power at crank shaft
also enables addressing full load fuel consumption. Port injection
turbo used enrichment to limit peak combustion chamber temperature.
Unburned fuel cools the flame and offsets the point where the enrich-
ment starts.

EGR also enables reduction in NOx emissions, by reducing the tem-
perature within the combustion chamber and thereby avoiding any chemi-
cal reaction of nitrogen present in the air. To do this, a small portion of the
exhaust is diverted back into the intake manifold using special flaps in the

Effects of Turbo Charging in S.I. Engines 229

exhaust manifold tubes, which has a cooling effect over the homogenous
charge mixture with in the cylinder.

Generally, EGR is induced by making use of long and narrow exhaust
pipes which reduce the temperature of the exhaust and in association with
VGT to get high exhaust velocity to run the turbine and thereby high boost
pressures and minimum turbo lag.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

10

Emissions Control by Turbo
Charged SI Engines

Emission is one of the important binding factors involved in the design
of any engine. The National emissions ceiling directives set binding emis-
sions ceilings to SO2, NOx volatile organic compounds, ammonia etc. In
the recent years, NOx emissions have reduce dramatically, some of the
credit goes to 3-way catalyst for gasoline engine vehicles and NOx storage
catalyst for GDI cars. But increases in the number of diesel cars, makes it
difficult to reduce it further.

Diesel engine actually accounts for an increase and there has been little
development in the truck sector. 2020 objective includes reduction in the
emission with 82% SO2 Reduction, 60% NOx, 51% VOC(Volatile organic
compounds), 27% ammonia, 59% particulates with respect to 2000 stan-
dards. NOx is either NO or NO2 depending upon the temperature within
the cylinder. NOx production is mainly due to NO reduction area of high
ozone concentration produce more NO2 by reaction with NO, which is
harmful to humans. All diesel engines have oxidation catalysts which con-
vert Hydro-Carbons and CO into CO2 and H2O and simultaneously oxidize
NO to NO2. Diesel particle filters use some of the NO2 for regeneration,

231

232 Liquid Piston Engines

but only occasionally. Some have constant regeneration but won’t use all
NO2. Modern cars use DENOx catalyst to adsorb NO2 and NO. Gases are
treated and become N2 ad H2O by using H2 in exhaust gas stream. Some
NO2 will escape and some released during regeneration. NOx treatment
in diesels will be the main focus of EURO 6, although specific targets are
not set.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

11

Scope of Turbo Charging
in SI Engines

Today, turbo charging is most commonly used on two types of engines:
Gasoline engines in high-performance automobiles and diesel engines in
transportation and other industrial equipment. Small cars in particular
benefit from this technology, as there is often little room to fit a larger-
output (and physically larger) engine. The Porsche 944 utilized a turbo unit
in the 944 Turbo (Porsche internal model number 951), to great advantage,
bringing its 0–100 km/h (0-60 mph) times very close to its contemporary
non-turbo “big brother”, the Porsche 928. In the 1980s, turbocharged cars
were difficult to handle. The tuned engines fitted to the cars, and the often
primitive turbocharger technology meant that power delivery was unpre-
dictable and the engine often suddenly delivered a huge boost in power at
certain speeds. As turbocharger technology improved, it became possible
to produce turbocharged engines with a smoother, more predictable but
just as effective power delivery.

In future, to meet U.S. emission regulations, injector systems will have
to be optimized with increased functionalities, multiple injection strate-
gies and increase pressures (up to 2000 bars or more). However improved

233

234 Liquid Piston Engines

combustion systems will have to be supplemented with after – treatment
techniques like selective catalytic reduction (SCR) for NOx reduction and
state of art particulate filters. Turbo charging would be adopted much
faster in downsized engines for entry level vehicles, on account of their
low cost. Turbo charging in GDI offer most potential, but at a cost. Forced
induction of traditional port injected (MPFI) engines is an attractive and
economical compromise. While turbo charged GDI can use 15% to 20%
less fuel than naturally aspirated variants, turbo MPFI uses 5% to 15%
less. Now Volkswagen is taking turbo GDI into mass market with 1.4 liters
Turbo charged SI (TSI) engines in the Golf and soon to be launched Polo.
But Renault has opted for turbo MPFI for 1.2 liter TCE 100 engine used in
Clio and Twingo. Fiat is also taking the same approach with 1.4 liters T-JET
engines for Bravo.

Twin-staged turbos with two turbo chargers and extra valves is an attrac-
tive in long engine with a cross flow design like in V-8 engines and even in
sportier and middle sized cars. Turbo chargers with twin scrolls, not only
improves scavenging but also turbo efficiency at low speed and inertia i.e.
improve transient response. Twin scroll turbine housing, ball bearing and
VGT will improves the boost pressures and quicker change from one driv-
ing mode to another.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

12

Summary

1. The biggest opportunity for improving the spark ignition
engine is boosting and downsizing.

2. Stoichiometric operation enables very low air pollutant
emissions.

3. Many other design variables could contribute:
E.g. increase compression ratio, variable valve control, and
lower friction etc.

4. The major challenge is controlling knock.
5. 20–30% higher part-load efficiency plausible.

235

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

13

Conclusions and Future Work

13.1 Conclusions

The present work establishes benefits of using various intrusive as well as
non-intrusive methods to analyze noise and vibrations from a dual cyl-
inder diesel engine. Various indices described can serve as good indict-
ors for condition monitoring of engines. However some of these may be
sensitive towards interference of background noise or chamber resonance.
This drawback may be surpassed using suitable filters. Various Frequency
bands in which combustion process or piston slap dominates have been
investigated.

13.2 Contributions

1. Investigation of effects of location of various transducers
towards signals acquired.

2. Use of Cepstrum analysis to study combustion noise.
3. The application of COMSOL -7 software in conjunction

with FEA for analysis of secondary motion of skirt.
237

238 Liquid Piston Engines

13.3 Future Recommendations

The noise, vibration and harness analysis of diesel engines has been an
active area of research during past few years. This work has tried to deal
with some aspects of this issue. There are many areas in which current fur-
ther work can be done in future [1]. Some of these include:

1. Quantification of noise emissions.
a. Subjective approach
Some of indices used for this purpose include:
i. Ranking-Various subjects may be asked to rank sound
emissions according to annoyance in a scale of 1 to
10. However number of samples must be kept low to
avoid complexity [2–4].
ii. Comparison in pairs-In this method various subjects
may be asked to evaluate relative judgments on the
basis of pairs, however this method can be exhaustive
as number of pairs can be large [5, 6].
b. Objective approach-Various psychoacoustic indices that
can be used for evaluation include:
i. Loudness-It is parameter for intensity evaluation and
has unit of phon or sone. Loudness level is SPL of a
pure tone plane wave of 1 kHz frequency as perceived
by human ears in frontal direction [7].
ii. Sharpness-A 60 dB sound wave of 1 kHz frequency
has sharpness of 1 acum. Sharpness of a soundwave
can be lowered by either adding low frequency com-
ponents or by decreasing high frequency compo-
nents [8].
iii. Roughness-This parameter takes into account mod-
ulation of waves. Its unit is aper.1 asper is rough-
ness of a tone of 1 kHz at 60 dB which is modulated
by 70  Hz frequency with degree of modulation
unity [9].
iv. Impulsiveness-This parameter represents impulsive-
ness of sound pressure level. It represents the ampli-
tude and frequency of occurrence of peaks. Its unit is
Kurt and is most significant during ideal running of
engines [8].

2. Motion of gudgeon pin inside pin hole needs to be taken
into account.

Conclusions and Future Work 239

3. Piston pin is held inside hole either by a full floating system
or a semi floating one. For case of full floating system both
pin and connecting rod are made of steel, whereas in case of
semi floating system piston is made of aluminum alloy and
pin is made of steel. Hence a semi floating system is subjected
to more noise due to differences in thermal expansion coef-
ficients of different materials used. It has been observed that
pin rotates counter clockwise inside hole before strikes the
wall of piston vertically in crank angle duration 20° BTDC-
30° BTDC [10]. Further movement of oil inside pin hole can
be visualized by particle tracking velocimeter(PVT).

4. Use of gap sensors/Telemeter device to study piston second-
ary motion using different skirt profiles.
Frictional power losses for different skirt profiles can be
evaluated for various engine strokes. Skirt profile having recess
at top and bottom part of skirt has shown minimum frictional
forces as it has better lubrication load bearing surface [11].

5. Use of AVL EXITE for modelling of piston motion.
This takes into account thermal distortions of liner using

GUID (Piston-liner guidance) and EPIL (Elastic piston liner
contact) approaches [12]. Surface velocities can be analyzed
both in time and frequency domains towards thrust as well
as anti-thrust side. At higher speeds, in conjunction with
higher inertial forces, piston secondary motion has been
found to fall. Hence both approaches have shown almost
same results [13].
6. Investigation into effects of bubble formations, mist and
cavitation of lubrication oil during analysis of secondary
motion of piston [14].
7. Use the discussed methodology by varying type of injection
or use of EGR or turbo charging.
a. Effects of post injection-specific consumption of fuel

NOX and emissions can be reduced by increasing the
amount of post injected fuel and advanced injection tim-
ings.However smoke emissions were found to remain
unaffected by post injection [15].
b. Effects of EGR-EGR has been found to reduce combus-
tion noise above 300 Hz range, however excessive use of
EGR causes lowering of thermal efficiency and increase
in emissions [16].
c. Effects of turbocharging-[17]

240 Liquid Piston Engines

8. Use of Blind Source Separation (BSS) and Independent
Component Analysis(ICA)methods for effective noise
source separation.

References and Bibliography

1. Monelletta, L., “Contribution to the study of combustion noise of automotive
diesel engines”, Phd Thesis, University polytechnic velencia, 2010.

2. Otto, N.C., Amman, S., Eaton, C., Lake, S., “Guidelines for jury evaluations of
automotive sounds”, SAE Technical paper-1999-01-1822, 1999.

3. Guski, R., “Psychosocial methods for evaluating sound quality and assessing
acoustic information”, Acta Acustica, 1997.

4. Bisping, R., and Giehl, S., “Psychological analysis of the sound quality of
vehicle interior noise: field and laboratory experiments”, proceedings of AVL
conference on engine and environment, pp. 65–86, 1996.

5. Hussain, M., Golles, J., Ronacher, A., and Schiffbanker, H., “Statistical evalu-
ation of an annoyance index for engine noise recordings”, SAE paper no
911080, 1991.

6. Kahn, M., Johansson, O., Lindberg, W., and Sundback, U., “Development of
an annoyance index for heavy duty diesel engine noise using multivariate
analysis”, NCEJ, vol 4, pp. 45, 1997.

7. Zwicker, E., and Fastl, H., “Psychoacoustic-Facts and models, II edition”
Springer, 1999.

8. Schiffbanket, H., Brandl, F., and Thien, G., “Development and application of
an evaluation technique to assess the subjective character of engine noise”,
SAE paper no 911081, 1991.

9. Peluger, M., Holdrich, R, Brandl, F., and Biermayer, W., “Subjective assess-
ment of roughness as a basis for objective vehicle interior noise quality evalu-
ation”, SAE paper no 1999-01-1850, 1999.

10. Kondo, T., and Ohbayashi, H., “Visualization of Oil behavior when piston pin
noise occurs”, Honda R&D Technical Review, 2011.

11. Kim, S., Shah, P., “A study of friction and lubrication behavior for gasoline
piston skirt profile concepts”, SAE Technical Paper no 09PFL-1163, 2009.

12. Kocaoglu, C., Tabak, M., “Comparison of two modeling techniques for pis-
ton-liner interaction in terms of piston secondary motion using AVL exite”,
OTEKON 2014, Bursa, Turkey, 2014.

13. AVL EXITE Power unit users guide, Vol 3, pp. 404, 2009.
14. www.SAE.com
15. Park, Y., Bae, C., “Effects of single and double post injection on Diesel PCCI

combustion”, SAE Technical Paper 2013-01-0010, doi:10.4271/2013-01-0010,
2013.

Conclusions and Future Work 241

16. Shibata, G., Ushijima, H., Ogawa, H., and Shibaike, Y., “Combustion Noise
Analysis of Premixed Diesel Engine by Engine Tests and Simulations,” SAE
Technical Paper 2014-01-1293, doi:10.4271/2014-01-1293, 2014.

17. Rakopoulos, Giakoumis, “Experimental study of combustion noise radiation
during transient turbocharged diesel engine operation”, Vol 36, pp. 4983–4995,
2011.

18. Ochiai, K., and Yokota, K., “Light-weight, quiet automotive DI diesel engine
oriented design method”, SAE paper 820434, 1982.

19. Maetani, Y., Niikura, T., Suzuki, S., Arai, S., and Okamura, H., “Analysis and
reduction of engine front noise induced by the vibration of the crankshaft
system”, SAE paper 931336, 1993.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

Glossary

TDC Top dead center

BDC Bottom dead center

Isothermal A process in which temperature remains constant

Isobaric A process in which pressure remains constant

Isentropic A process in which entropy is constant

Isometric A process in which volume is constant, also known as

isovolumetric

LTD Low temperature difference

HTD High temperature difference

249

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

List of Important Terms

1. Stoichiometric fuel/air ratio: A mixture that contains just
enough air for complete combustion of all fuel mixture is
known as chemically correct or Stoichiometric ratio.

2. Mean effective pressure: Average pressure inside cylinders
of an IC engine based on calculated or measured power out-
put is known as mean effective pressure.

3. Twin-scroll turbochargers: A turbocharger compris-
ing a turbine, director, and compressor. The turbine may
be formed as a turbine wheel surrounded by at least two
scrolls.  The at least two scrolls may direct exhaust gases
supplied thereto toward the turbine wheel to cause rotation
thereof. The director may control distribution of the exhaust
gases between the at least two scrolls to optimize circumfer-
ential velocity in the scroll or volute, and thus impingement
velocity on the turbine. The compressor may be driven by
the turbine.

4. Turbo Lag: A turbocharger uses a centrifugal compressor,
which needs rpm to make boost, and it is driven off the

243

244 Liquid Piston Engines

exhaust pressure, so it cannot make instant boost. It is espe-
cially hard to make boost at low rpm. The turbo takes time
to accelerate before full boost comes in; it is this delay that is
known as turbo lag. To limit lag, it is important to make the
rotating parts of the turbocharger as light as possible. Larger
turbos for high boost applications will also have more lag
that smaller turbos, due to the increase in centrifugal mass.
Impeller design and the whole engine combo also have a
large effect on the amount of lag. Turbo lag is often confused
with the term boost threshold, but they are not the same
thing, lag is nothing more the delay from when the throttle
is opened to the time noticeable boost is achieved.
5. Turbo Boost: Usually measured in pounds per square inch,
it is the pressure the turbocharger makes in the intake mani-
fold. One of the ways to increase airflow through a passage
is to increase the pressure differential across the passage.
By boosting the intake manifold pressure, airflow into the
engine will increase, making more power potential. Boost is
also measured in Bar. One Bar equals 14.7 psi.
6. Boost Threshold: Unlike turbo lag, which is the delay of
boost, boost threshold is the lowest possible rpm at which
there can be noticeable boost. A low boost threshold is
important when accelerating from very low rpm, but at
higher rpm, lag is the delay that you feel when you go from
light to hard throttle settings.
7. Waste gate: The waste gate is a valve that allows the exhaust
gasses to bypass the turbine. The waste gate relies on boost
pressure to open it. Spliced into the waste gate pressure feed
there must be some form of pressure bleed. By bleeding pres-
sure to the waste gate, it is possible to control the amount of
boost by reducing the pressure at the waste gate.
8. Turbo Cool Down: A turbocharger is cooled by engine oil,
and in many cases, engine coolant as well. Turbos get very
hot when making boost, when you shut the engine down the
oil and coolant stop flowing. If you shut the engine down
when the turbo is hot, the oil can burn and build up in the
unit (known as “coking”) and eventually cause it to leak oil
(this is the most common turbocharger problem). It is a
good idea to let the engine idle for at least 2 minutes after
any time you ran under boost. This will cool the turbo down
and help prevent coking.

List of Important Terms 245

9. Multi port fuel injection (MPFI): In the multi point fuel
injection system an injector is located in the intake manifold
passage. The fuel is supplied to the injectors via a fuel rail
in the case of top fed fuel injectors and via a fuel galley in
the intake manifold in the case of bottom fed fuel injectors.
MPFI systems provide better performance and fuel econ-
omy as compared to TBI. Most of the MPFI systems use one
injector per cylinder but in certain applications up to two
injectors per cylinder is used to supply the required fuel for
the engine.

10. Central multi-port fuel injection (CMFI): This is a varia-
tion of MPFI system but in this case the injectors (usually
one per cylinder) are located in a plastic molded pod and the
fuel is distributed to the intake ports via a polymeric hose.
To avoid fuel distribution variations a fuel pressure activated
poppet valve is installed at the end of the hose. The injectors
are activated via the ECU in a similar fashion as in the MPFI
fuel systems.

11. Tuned port injection (TPI): A TPI is a fuel/air management
system that has a tuned induction system to optimize airflow
to each cylinder. This system was developed to obtain the
broadest possible torque curve. A single throttle body and
one injector per cylinder are used in this configuration. The
intake manifold incorporates long runners whose length is
tuned to the desired torque curve. For low and mid range
torque longer runners are utilized in this application.

12. Direct fuel injection (DFI): In a direct fuel injection system
one injector is located in the cylinder head for each cylinder.
The high-pressure fuel (single fluid) or low-pressure air/fuel
mixture (dual fluid) is metered directly into the combustion
chamber when the electromagnetic valve is activated by the
ECU. This fuel injection system offers the latest in engine
management systems and offers the best in engine perfor-
mance, low exhaust emissions and fuel economy.

Liquid Piston Engines. Aman Gupta, Shubham Sharma, and Sunny Narayan.
© 2017 Scrivener Publishing LLC. Published 2017 by John Wiley & Sons, Inc.

Bibliography

Watson and Janota, 1984, “Turbo charging of I.C. engines,” Macmillan press, New
York, 1983.

William Harry Crouse, Donald L. Anglin, “Automotive engines,” Glencoe, 1994,
ISBN 9780028010991.

J.B. Heywood, 1998, “Internal Combustion Engine Fundamentals,” McGraw-Hill
Education.

V. Ganesan 1994, “I.C. engines,” Tata Mcgraw hill, New Delhi, ISBN
978-1-25-900619-7.

Vasandani, Kumar, 1979, “Heat Engineering: In MKS and SI Units,” Metropolitan .
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Kant, K., Pati, A., Viswanath, B., and Thiyagarajan, R., “Cyclic Irregularities in Idle

and Fuel Delivery Variation of a Rotary Fuel Injection Pump,” SAE Technical
Paper 2004-32-0056, 2004, doi:10.4271/2004-32-0056.
Han, Z., Henein, N., and Bryzik, W., “A New Ignition Delay Formulation Applied
to Predict Misfiring During Cold Starting of Diesel Engines,” SAE Technical
Paper 2000-01-1184, 2000, doi:10.4271/2000-01-1184.
Brunt, M. and Platts, K., “Calculation of Heat Release in Direct Injection Diesel
Engines,” SAE Technical Paper 1999-01-0187, 1999, doi:10.4271/1999-01-0187
AUTO ENGINEER MAGAZINE (EDITION – AUG, 2007) http://ae-plus.com/
AUTO-CAR MAGAZINE www.magazineexchange.co.uk/Autocar-2007

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