International Journal of Research In Science & Engineering
Volume: 1 Issue: 1
e-ISSN: 2394-8299
p-ISSN: 2394-8280
ENERGY ANALYSIS OF DIESEL AND BIODIESEL FUELLED
C.I.ENGINE USING FIRST LAW OF THERMODYANAMIC -A REVIEW
Sagar P.Potdukhe
1
PG Student, Department of Mechanical Engineering, Govt. College of Engineering, Amravati.
Sagar.potdukhe@gmail.com
ABSTRACT
Modeling has been used to investigate the combustion performance of a single cylinder direct injection diesel
engine fuelled by biodiesel like cotton seed oil. The focus of the present study is to review the different available
model used for modeling of CI engines. The modeling of CI engine is divided into single zone, multizone and
multi-dimensional model. The model evaluation is made based on the time complexity, space complexity, and
prediction accuracy using the developed computer program like MATLAB. Our focus is on single zone model
which further subdivided in many submodel i.e. heat release rate, heat transfer, ignition delay period, droplet
evaporation, intake and exhaust flow and combustion model. The numerical simulation was performed with the
standard specification of a CI engine by using MATLAB software with least time.
,Keywords: Diesel engine modeling, IC Engine, Biodiesel, cotton seed oil methyl ester
----------------------------------------------------------------------------------------------------------------------------1. INTRODUCTION
Engine simulation has been extensively used to improve the engine performance. Compression ignition direct
injection (CIDI) diesel engines have been widely used in heavy-duty vehicle, marine transportation and now have
been increasingly being used in light duty vehicles, particularly in Europe and Japan Experimental work which is
aimed at fuel economy and low pollutants emission for IC engine requires change in input parameter which is highly
demanding in terms of money and time. So, in order to overcome this drawback, an alternative simulation of engine
performance with the help of mathematical model and powerful digital computers lowers the cost and time. [2]
In these simulation models, the effect of various design structures like design of combustion chamber input
parameters (intake pressure, injection timing, etc.) and operation changes (compression ratio, speed, etc.) can be
estimated in fast and non-expensive way provided that main mechanism are recognized and modeled perfectly to
meet the experimental results. [12]
Depending upon the various possible applications different types of models for diesel engine combustion process
have been in use. In the order of increased complexity and increased computer system requirements these can be
classified as single zone model models, quasi dimensional phenomenological models and Multidimensional
computational fluid dynamics models. These models can reduce the number of experiments. [5]
In case of single zone model cylinder temperature, pressure and mass can be obtained from ordinary differential
equations by using the first law of thermodynamics and equation of state in each process. Biodiesel (cotton seed oil)
have become an alternate to petro diesel in the view of the faster depletion of petro diesel. Understanding the aspects
of biodiesel combustion is now possible with the simulation models. The measured pressure rise in an engine is used
to tune the model and helps in calculating the rate of heat release from the engine cylinder.[6]
2. Energy analysis
In engineering, modeling a process has come to mean developing and using the appropriate combination of
assumption and equation that permit critical feature of the process to be analyzed. The modeling of engine process
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Volume: 1 Issue: 1
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continues to develop as our understanding of the physical and chemistry of the phenomena of interest steadily
expand and as the capability of computers to solve complex equation continues to increase.[1]
2.1 Types of models.
These can be two categorized.
1) Thermodynamic model.
 Zero-dimensional single zone.
 Quasi-dimensional multi-zone.
2) Fluid dynamic model.
 Multidimensional model.
1) Thermodynamic model.
Thermodynamic energy conservation based model are zero-dimensional (since in the absence of any flow
modeling, geometric feature of the fluid motion cannot be predicted), phenomenological (since additional detail
beyond the energy conservation equation is added for each phenomenon in turn), and quasi-dimensional (where
specific geometric feature, example, the spark-ignition engine flame or diesel fuel spray shapes, are added to
basic thermodynamic approach).
In single zone models, the working fluid in the engine is assumed to be a thermodynamic system, which
undergoes energy and/or mass exchange with the surroundings and the energy released during the combustion
process is obtained by applying the first law of thermodynamics to the system.[1]
2) Fluid dynamic model.
Fluid-dynamic based model are often called multi-dimensional model due to in their inherent ability to
provide detailed geometric information on the flow field based on solution of the governing flow equation. [1]
2.3 Governing Equations for Open Thermodynamic System
It is often required the model a region of the engine as an open thermodynamic system. Example are the
cylinder volume and the intake and exhaust manifolds (or portion of these volume). Such a model is appropriate
when the gas inside the open system boundary can be assumed uniform in composition vary with time due to heat
transfer, work transfer and mass flow across the boundary, and boundary displacement. Such an open system is
illustrated in fig. 1.2.1. The important equations are mass and energy conservation.
These equations for an open system, with time or crank angle as the independent variable, are the building
blocks for thermodynamic-based model.[1]
2.3.1 Conservation of Mass
The rate of change of the total mass m of an open system is equal to the sum of the mass flows into and out of
the system[1]:
̇ 𝑗 𝑚̇ 𝑗
𝑚= ∑
Mass flow into the system are taken as positive; mass flow out are taken as negative.
Fig.1.1 Open thermodynamic system
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For conservation of the fuel chemical element, it is convenient to use the fuel fraction f, which is define as 𝑚𝑓 /𝑚,
where 𝑚𝑓 denotes the mass of fuel (or fuel elements in the combustion product) in open system
𝑑
𝑚𝑓 = (𝑚𝑓 ) = ∑𝑗 𝑚̇𝑓 = ∑𝑗 𝑚̇𝑗̇ 𝑓𝑗
(1.2)
𝑑𝑡
Differential of eq.(1.2) lead to an equation for the rate of change of fuel fraction:
𝑚̇
𝑗
𝑓̇ = ∑𝑗 ( ) (𝑓𝑗 − 𝑓)
(1.3)
𝑚
The fuel/air equivalence ratio ∅ is related to f via ∅ = [(𝐹/𝐴)𝑠 (1 − 𝑓)]. Hence the rate of change of equivalence
ratio of the material in the open system is
∅̇ =
𝑓̇
1
(1.4)
(𝐹/𝐴)𝑠 (1−𝑓)2
2.3.2 Conservation of Energy
The first law of thermodynamic for open system in fig. 1.1 can be written [ref 3]:
𝐸̇ = 𝑄𝑤̇ − 𝑊̇ + ∑𝑗 𝑚̇𝑗̇ 𝑓𝑗
(1.5)
̇
𝑄𝑤 is the total heat transfer rate into the system, across the system boundary and equal the sum of the heat transfer
rate out of the system across each part of the boundary. 𝑊̇ is the work transfer the rate out of the system across the
boundary; where the piston is displaced, the work transfer rate equal to pV. Because all energies and enthalpies are
expressed relative to the same datum, it is not necessary to include the heat released by combustion in Eq. (1.5); this
is already accounted for in the energy and enthalpy terms.
The goal is the define rate of change of state of the open system in terms of 𝑇̇ and 𝑝̇ . Two approaches are
commonly used, depending on whether the thermodynamic property routine provide values for internal energy u or
enthalpy h. thus 𝐸̇ in Eq. (1.5) can be expressed
𝐸̇ =
𝑑
𝑑𝑡
(𝑚𝑢)
𝐸̇ = (𝑚ℎ) −
or
𝑑
𝑑𝑡
(𝑝𝑉)
It is assumed that the system can be characterize by T, p, and ∅; thus
u = u(T, p, ∅) ℎ = ℎ(𝑇, 𝑝, ∅)
𝜌 = 𝜌(𝑇, 𝑝, ∅)
And the rate of change of u, h, and 𝜌 can be written in the form
𝜕𝛼
𝜕𝛼
𝜕𝛼
𝛼̇ = ( ) 𝑇̇ + ( ) 𝑝̇ + ( ) ∅̇
(1.6) Where 𝛼 is u, h, and 𝜌 . Using the ideal gas law in its two
𝜕𝑇
𝜕𝑝
𝜕∅
forms, 𝑝 = 𝜌𝑅𝑇 and 𝑝𝑉 = 𝑚𝑅𝑇, and Eq. (8) for 𝜌̇ , an equation for 𝑝̇ can be derived :
𝑝̇ =
̇
𝜌
𝑉̇
1 𝜕𝜌
1 𝜕𝜌̇
𝑚̇
(− −
𝑇̇ −
∅+ )
𝜕𝜌⁄𝜕𝑝
𝑉 𝜌 𝜕𝑇
𝜌 𝜕∅
𝑚
(1.7)
Returning now to the energy conservation equation, expressing 𝐸̇ in term of 𝑢̇ 𝑜𝑟 ℎ̇ and 𝑢̇ 𝑜𝑟 ℎ̇ in term of partial
derivatives with respect to T, p, and∅, and substituting for 𝑝̇ with Eq. (1.7), one can obtain equations for:
𝑇̇ = [𝐵 −
𝑝 𝜕𝑢 𝑚̇
𝑉̇
𝐷 𝜕𝑝 𝑚
V̇
1
𝑉
( −
B = −RT +
C= 1+
V
m
T ∂R
+
𝜕𝑅
𝜕∅
∅)]⁄(
𝜕𝑢
𝜕𝑇
+
(Q ẇ + ∑j ṁ j hj − ṁu)
R ∂T
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𝐶 𝑝 𝜕𝑢
𝐷 𝑇 𝜕𝑝
)
(1.8)
(1.9)
(1.10)
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and m is mass of gas in the cylinder, 𝑚̇ 𝑖 and 𝑚̇ 𝑒 are the mass flow rates through the inlet valve and the exhaust
valve, and f is the fraction mf/m. The subscript i and e denotes the properties of the flow through the intake and the
exhaust valves respectively.[1]
Compression
m = 0 ̇ and ḟ = 0
̇
̇
v
Bm
B
V
Q
Ṫ = (− − w )
A
(1.11)
Exhaust:
ṁ
ḟ = − e (fe − f)
ṁ = −ṁ e
B
ṁ e
A
m
T = [−
h
V̇
m
C
B
V
B
(1 − ) − − φ +
1
Bm
(1.12)
(−ṁ e he − Q ẇ )]
(1.13)
Where,
he is the enthalpy of flow through the exhaust valve.
Above modelling technique have been traditionally used in two different directions [1]
 In one way, both these models have been used to predict the in-cylinder pressure as a function of crank
angle from an assumed energy release or mass burned profile (as a function of crank angle).
 Another use of these models lies in determining the energy release/mass burning rate as a function of crank
angle from experimentally obtained in-cylinder pressure data.
2.3.3 Engine modeling general equations for first-law of thermodynamics
The first-law of thermodynamics applied to the engine cylinder reads [4].
dQ 𝐿
dφ
−𝑃
dV
dφ
=
dU
dφ
− ∑𝑗
𝑑𝑚𝑗
dφ
ℎ𝑗
(1.14)
where dQL/dφ is the rate of heat loss to the cylinder walls, as described by the semi-empirical Annand or
Woschni correlations, p is the pressure of cylinder contents, dmj is the mass exchanged (positive when entering) in
the step dφ and hj is the specific enthalpy of it. Subscript j denotes fuel injection, and exchange with the exhaust
manifold, inlet manifold and crankcase .
Ideal gas relation is given by
𝑝𝑉 = 𝑚𝑅𝑠 𝑇
(1.15)
The ideal gas relation can be expressed in differential form as:
𝑃
dV
dp
dT
dT
+𝑉
= 𝑚𝑅𝑠
+ 𝑅𝑠 𝑇
dφ
dφ
dφ
dφ
(1.16)
The application of the first-law of thermodynamics to the engine cylinder (open cycle).
∑ 𝑚𝑖 𝑐vi
𝑖
𝑑𝑚𝑗
dT dQ 𝐿 𝑚𝑅𝑠 𝑇 dV
𝑑𝑚𝑖
=
−
+∑
ℎ𝑗 − ∑ 𝑢𝑖
dφ
dφ
𝑉 dφ
dφ
dφ
𝑗
𝑗
The volume V of the engine cylinder, needed in the previous equations,
𝐷2
𝑉 = 𝑉𝑐𝑙 + 𝜋 ( ) 𝑥
4
With Vcl the clearance volume and x the piston displacement from its TDC position
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𝑥 = 𝑟(1 − 𝑐𝑜𝑠𝜑) + 𝐿[1 − √1 − 𝑓 2 𝑠𝑖𝑛2 𝜑 ] (1.19)
Where L is the connecting rod length, r the crank radius, f = r/L and the crank angle 4 is measured from the bottom
dead centre position (BDC). Consequently[4]
dV 𝐴𝜋𝐷2
𝑓 cos 𝜑
=
[𝑟 sin 𝜑 (1 +
)]
dφ
4
√1 − 𝑓 2 𝑠𝑖𝑛2 𝜑
(1.20)
dφ = 6N dt
With N the engine speed expressed in rpm, for transforming the various terms from time to degree crank angle basis.
3. Engine Submodels
Submodels included in the study of the smallest region inside the engine cylinder which helps in proper
understanding of the actual phenomena of the engine. It includes the different engine submodels which also help to
study complex behaviour of the combustion.[12]
3.1 Intake and Exhaust Manifold Submodel
Intake and exhaust manifold Submodel deals with the mass flow rate inside the cylinder either the air entering
during the intake stroke, mass of fuel during the compression stroke and the amount of gases flow through the
exhaust stroke. The timing and rate of fuel injection into the chamber affects the spray dynamics and combustion
characteristics.
If the upstream pressure of the injector nozzle is known and assumed that flow through the nozzle is incompressible,
one dimensional and quasi-steady, then the mass flow rate of fuel injected inside the cylinder through the nozzle is
given by [2].
𝑚 = 𝐶𝑑 𝐴𝑛 √2𝜌𝑙 ∆𝑃
(1.21)
In this type of mass flow rate, the intake and exhaust tanks (submodel) are treated as plenums with known
temperatures and pressures. Earlier these models do not provide any information whether the flow is subcritical or
supercritical at the inlet or outlet. They can be solved with the help of conservation of mass equation: in order to
calculate the mass flow rate the equation is given by [2]
𝜑1
𝑚 = 𝐴𝑒 𝑃1 √
𝑅1 𝑇1
𝜑1 =
2𝛾
𝛾−1
2⁄𝛾
𝑃2
[(
2⁄𝛾
𝑃1
(𝛾+1)⁄𝛾
𝑃2
)−(
(𝛾+1)⁄𝛾
𝑃1
)]
(1.22)
These equations are further modified by taking into account the critical behavior of the fluid at intake manifold and
exhaust manifold. The intake process is assumed as a subcritical flow and the intake flow rate is given as [2]
𝑑𝑚𝑠
𝑑∅
=
1
6𝜂
𝜇𝑠 𝐹𝑠 √
2𝑘𝑠
𝑘𝑠 −1
2⁄𝑘𝑠
𝑃2
[(
2⁄𝑘
𝑃1 𝑠
)−(
(𝑘 +1)⁄𝑘𝑠
𝑃2 𝑠
(𝑘 +1)⁄𝑘𝑠
𝑃1 𝑠
)]
(1.23)
During the exhaust process, the flow at outlet valve is considered as a supercritical flow due to higher pressure
difference between cylinder and exhaust port. When the pressure at the outlet valve decreases, then the behavior of
flow changes to subcritical flow.
3.2 Modeling of heat release rate
For empirical zero-dimensional modeling, the heat release approach is commonly used. In this, the thermal
effects of combustion, rather than the process itself are modeled. Expressions are developed which correlate the
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gross heat release rate to the cylinder pressure trace using first law of thermodynamics. The generated heat release
data can be presented using mathematical tools such as Wiebe functions. These models consist of coefficients whose
values need to be determined experimentally. The models can be verified by comparing the predicted cylinder
pressure with the experimentally obtained cylinder pressure. The heat-release model should include[1]:
a) An empirical or analytical description of ignition delay,
b) An expression for the heat released during the pre-mixed phase,
c) An expression for the heat released during the diffusion phase.
3.3 Ignition delay modeling:Ignition delay can be defined as the time interval between the start of fuel injection and the start of combustion.
Many researchers have given correlations for predicting ignition delay in which the earliest was Wolfer (1938). The
empirical formula used in the present model is shown in equation which was given by Hardenberg. This equation
gives the value of ignition delay in terms of values of degrees of crank angle using temperatures (T) in Kelvin and
pressure (P) in bar. The equation (33) has been used to predict the ignition delay.
The ignition delay is modelled with an Arrhenius type of correlation. [1]
𝜏 = 3.45. 𝑃 −0.2 exp(2100⁄𝑇)
(1.24)
The three important variables to be taken into account for ignition delay modeling are mean temperature, mean
pressure and load. Engine load is given in terms of the percentage of maximum torque output. So modified equation
is given by
𝜏 = 𝑎. ∅−𝑏 𝑃−𝑐 exp(2100⁄𝑇)
(1.25)
Ø = load in percent
3.4 Mass Flow Model
Thermodynamic model based on the step-by-step ‘emptying and filling method described by Watson and
Janota. The conservation equations are used to evaluate the work, heat and mass transfer taking place across the
boundaries of the control volume. The equations are solved at each crank & angle interval over an engine cycle of
360 degrees for a two-stroke engine or 720 degrees for a four-stroke engine. A comparison is made between the
calculations for the previous cycle and the present cycle. If the convergence criteria have been met, the results will
be recorded and the calculations stopped; otherwise a new cycle is commenced. Convergence usually takes between
five to eight cycles.[13]
A constant manifold pressure is assumed, and the mass transfer rate across valves and ports
is calculated using the orifice analogy. For subsonic flow, this is described by
√
̇
𝑚̇ = 𝐶𝑑𝐴𝑃
𝑢 (
2𝛾
𝛾−1
)(
1
𝑅𝑢 𝑇𝑢
𝑃
) [( 𝑑)
𝑃𝑢
2⁄𝛾
𝑃
− ( 𝑑)
𝑃𝑢
(𝛾+1)⁄𝛾 (𝛾+1)⁄𝛾
]
(1.26)
For chocked flow
𝑚̇ = 𝐶𝑑𝐴𝑃𝑢 √(
1
𝑅𝑢 𝑇𝑢
)(
2
𝛾+1
)
(𝛾+1)⁄(𝛾−1)
(1.27)
Where 𝑚̇ is the mass flow rate through the valve or port, Cd is the discharge coefficient, A is the area, and γ, R, T,
and P refer to the specific heat ratios, the specific perfect gas constant, absolute temperature, and pressure,
respectively, for the fluid. The subscripts d and u denote downstream and upstream, respectively. The product Cd A
represents an effective flow area, and it is calculated at each crank angle step corresponding to the valve or port
timing input by the user. The value for Cd may be either a fixed average value, or it may be based on a set of
nondimensional flow coefficients which are stored in simulation model as functions of valve lift. In the latter case,
the user has a choice of either a high-swirl or zero-swirl inlet port and either a good or poor exhaust port. As a
means of simplifying the input data, the program contains a typical valve lift polynomial to calculate the lift at each
crank angle. The same polynomial is used for both inlet and exhaust valves.[13].
3.5 Heat Transfer Model
One-dimensional heat flux model to describe the heat transfer through the wall into the coolant at each crank
angle step. Heat fluxes through the piston, cylinder liner, and cylinder head are calculated separately based on a
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resistance network analogy approach. The heat flow path between the cylinder gases and the engine coolant can be
considered to consist of three parts: gas to wall, conduction through wall, and wall to coolant. In the model, a
composite wall is used with the option of using up to three layers with different thickness and material properties.
Wall to coolant convection heat transfer is modelled using empirical bulk surface heat transfer coefficients.[12]
𝑑𝑄𝑤
= ℎ𝑖. 𝐴𝑖(𝑇𝑤, 𝑖 − 𝑇𝑔𝑎𝑠)
(1.28)
𝑑𝑡
4. LITERATURE SURVEY
The following references were studied for analysis of experimental results by using different model and
biodiesel in different working conditions on First law of Thermodynamic
Vivek Kumar Gaba, Prerana Nashine and Shubhankar Bhowmick(2012): developed a combustion model for a
diesel compression ignition (CI) engine for constant pressure combustion process. The work analytically examines
the performance of a CI engine with the minimum use of diesel fuel as pilot fuel, and bio-diesel as secondary fuel.
The combustion model has been developed for an ideal diesel engine using blends of biodiesel ranging from 20% to
100%. Using the first law of thermodynamics and equation of state in each process, the cycle was critically
analyzed. The specification of a standard CI engine was used for numerical calculations. The variation of
temperature with different equivalence ratio were studied and reported. The thermal efficiency of a pure diesel
engine was observed to decrease exponentially from 67% to 47% as the equivalence ratio increases from 0.7 to 1.3.
It was also observed that B- 20 and B-40 formed a good mixture among all other blends and the efficiency of pure
bio-diesel was comparatively less than diesel as well as for other blends.
A sharp increase in work output was observed as the equivalence ratio increases due to more fuel injection. It was
also observed that efficiency of pure bio-diesel is comparatively lesser than all other blends, but the maximum cycle
temperature for this blend is least. This indicates that NOx emission was not present in case of pure diesel; hence
pure biodiesel can be used resulting in less pollution despite of less efficiency & work output.
Work aims to develop a combustion model for a diesel compression ignition (CI) engine for constant pressure
combustion process. The work analytically examines the performance of a CI engine with the minimum used of
diesel fuel as pilot fuel, and bio-diesel as secondary fuel. The combustion model had been developed for an ideal
diesel engine using blends of biodiesel ranging from 20% to 100%. Using the first law of thermodynamics and
equation of state in each process, the cycle had been critically analyzed. The specification of a standard CI engine
had been used for numerical calculations. The variation of temperature with different equivalence ratio is studied
and reported. The thermal efficiency of a pure diesel engine is observed to decrease exponentially from 67% to 47%
as the equivalence ratio increases from 0.7 to 1.3. It is also observed that B- 20 and B-40 formed a good mixture
among all other blends.
The work in a one-dimensional combustion model, employing constant eddy diffusivity and a one-step chemical
reaction had been developed and applied to study the flame propagation in a Spark Ignition (SI) engine at 1600 and
4200 rpm under fuel rich conditions using one and two zone thermodynamic models. The thermodynamic models
had been compared with 1-D model for average mixture temperature, the temperatures of the burned and unburned
gases and the flame surface area and indicate that the one-dimensional model predictions are very sensitive to the
eddy diffusivity and reaction rate data where as the two-zone thermodynamic model predicts, first, a monotonically
increasing flame surface area with time and, then, a monotonically decreasing surface area.
BORDET Nicolas, CAILLOL Christian, HIGELIN Pascal (2010). This paper presented a new 0D
phenomenological approach to predict the combustion process in Diesel engines operated under various running
conditions. The aim of this work was to develop a physical approach in order to improve the prediction of incylinder pressure and heat release for developing a tool for engine pre-mapping. The main contribution of this study
was the modeling of the premixed part of the Diesel combustion.
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In phenomenological Diesel combustion models, the premixed combustion phase was sometimes modeled as the
propagation of a turbulent flame front. However, experimental studies had shown that this phase of Diesel
combustion was actually a rapid combustion of part of the fuel injected and mixed with the surrounding gas. This
mixture ignites quasi instantaneously when favourable thermodynamic conditions are locally reached. A chemical
process then controls this combustion. they were worked on Spray and entrained surrounding gas model, two states
turbulence model: K-k modified model, Vaporization model, Premixed Combustion Model, Diffusion Combustion
Model, Extended Model for Multi Injection.
Results were measured engine data are compared with computed results using the developed combustion model.
Injection rates were modeled with Hermits polynomials optimized for each injection. The determined model
parameters were then kept constant for the whole range of engine operating parameters. This experimental apparent
energy rate was compared with the apparent energy rate calculated from the simulated pressure traces.
Future work should improve the model to take into account large EGR rates as well as in-cylinder temperature
distribution. To improve the existing model, interaction between sprays must be taken into account in multi-injection
cases. Furthermore, a more detailed spray model can improve general results especially in cases with multiinjection.
Zehra Sahin, Orhan Durgun (2008). In the presented study, the aim was to develop a complete cycle model and
to prepare a computer code for determining complete cycle, performance characteristics, and exhaust emissions of
diesel engines. For this purpose, a computer program had been developed. To compute diesel engine cycles, zerodimensional intake and exhaust model developed by Durgun,. Details of fuel spray formation, fuel–air mixing, swirl
and heat transfer had been taken into account in this model. Also, an emission model based on the chemical
equilibrium and kinetics of NO had been developed to calculate the pollutant concentrations within each zone and
the whole of the cylinder. Using the developed computer program, complete engine cycle, engine performance
parameters and exhaust emissions could be determined easily. The values of the cylinder pressure and engine
performance parameters predicted by the presented model matched closely with the other theoretical models and
experimental data the accuracy of a newly developed model, like the presented one, must be controlled comparing
with the experimental and theoretical values given in the relevant literature. Numerical results obtained from the
presented model were compared with the experimental results and with the theoretical models, the results of which
were accepted to be at sufficient accuracy. These indicate that cylinder pressure values obtained from the presented
model are lower than Bazari’s results. However, formation and vaporization of fuel droplets were modeled, and
spray wall impingement was also taken into account in detail in Bazari’s model, and in the presented model and in
Bazari’s models, Annand’s correlation has been used for the calculation of heat transfer.
In future studies, it is planned to take into account a detailed wall impingement model and evaporation model.
The effects of the residual gases in the cylinder could be used and a more detailed heat transfer model would also be
included in the calculations.
C.D. Rakopoulos and E.G. Giakoumis (2006). This paper surveys the publications available in the open literature
concerning diesel engine simulations under transient operating conditions. Only those models that include both full
engine thermodynamic calculations and dynamic power train modeling are taken into account, excluding those that
focus on control design and optimization. Most of the attention is concentrated to the simulations that follow the
filling and emptying modeling approach. One of the main purposes of this paper is to summarize basic equations
and modeling aspects concerning in-cylinder calculations, friction, turbocharger, engine dynamics, governor, fuel
pump operation, and exhaust emissions during transients. The various limitations of the models are discussed
together with the main aspects of transient operation (e.g. turbocharger lag, combustion and friction deterioration),
which diversify it from the steady-state.
The stringent regulations concerning engine exhaust emissions dominate the automotive industry, forcing
manufacturers to new developments. Sophisticated, high pressure common rail injection systems, exhaust gas
recirculation and variable geometry turbochargers are applied for reduction of fuel consumption, pollutant emissions
and noise. In this work the publications concerning transient diesel engine simulation will be surveyed, provided
they include both engine thermodynamic and powertrain dynamic submodels. To maintain the paper size at an
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acceptable level, it was decided that works which put emphasis on controllers design rather than thermodynamics,
i.e. applying optimization algorithms, or Bode diagrams, will not be covered. Most of the attention will be paid to
the simulations that follow the filling and emptying modeling technique. This, to the opinion of the authors, had
proven so far to be the best approach to adequately shed light into the engine processes during transients. Moreover,
taking into account its computational requirements, it was the most promising one for successful transient exhaust
emissions predictions.
The transient response of compression ignition engines forms a significant part of their operation and it was of
critical importance, owing to the often non-optimum performance involved as regards turbocharged engines.
Moreover, the launch of emission directives, in the form of Transient Cycles, have directed engine manufacturers to
deal with overall (vehicles’) transient performance, since it was well established that the transient operation
contributes much more to the total amount of emissions than the corresponding steady-state one. The majority of
transient diesel engine simulations are based on the filling and emptying approach, which combines adequate insight
into the relevant phenomena and requires limited computational time. Quasi-linear models are met often, owing to
the low computational time required that makes them ideal for real time simulations.
Reliable study of pollutants emissions during transient operation via the use of suitable models is the most
important objective for the future. Its accomplishment is, for the moment, limited, due to the high computational time
required for the analysis of hundreds of cycles. Fortunately, experimental investigations of transient operation and
computational power of personal computers both flourish during the last years, forming a sound basis for an even
more successful investigation of dynamic engine operation in the near future.
Donepudi Jagadish, Ravi Kumar Puli and K. Madhu Murthy (2011): A zero dimensional model had been used
to investigate the combustion performance of a single cylinder direct injection diesel engine fuelled by biofuels with
options like supercharging and exhaust gas recirculation. The numerical simulation was performed at constant
speed. The indicated pressure, temperature diagrams are plotted and compared for different fuels. The emissions of
soot and nitrous oxide are computed with phenomenological models. Content was prediction of heat release and gas
properties, Calculation of ignition delay, Nitric Oxide Formation, Frictional Power Calculations, Indicated Power
and Brake Power calculations, Combustion duration, Prediction of soot formation, Frictional Power Calculations.
The peak pressures was lowered with ethanol blending with diesel. And supercharging operation resulted in
little lowering the maximum combustion pressure, temperatures in comparison to no supercharging case. The peak
cylinder pressures are low with biodiesel blends in comparison to diesel since the heating value of biodiesel is lower
than that of diesel resulting in lower heat release. Poor atomization and slow heat release rate also the reason for
lower peak pressure for pure biodiesel.
In future the experimental work is also carried out with biodiesel (palm stearin methyl ester) diesel blends,
ethanol diesel blends
P.A. Lakshminarayanan and Y.V. Aghav. The phenomenological combustion models are practical to describe
diesel engine combustion and to carry out parametric studies. This is because the injection process, which can be
relatively well predicted with the phenomenological approach, has the dominant effect on mixture formation and
subsequent course of combustion. The need for improving these models was also established by incorporating
developments happening in engine designs. A phenomenological model consisting of sub-models for combustion
and emissions are proposed in detail in this chapter. With more and more “model based control programs” used in
the ECU controlling the engines, phenomenological models are assuming importance now. The full CFD based
models though give detailed insight into the combustion phenomena and guide the design engineer, they are too
slow to be handled by the ECU’s or for laying out the engine design.
Md. Moinul Islam, Mohammad Anisur Rahman, Mohammad Zoynal Abedin.. An experimental investigation
had carried out for the first law analysis of a DI (direct-ignition) diesel engine running on straight soybean oil (SVO)
preheated at 50, 75, 100oC with different loads at varying speeds of 1750, 2000, 2250 rpm. The results show that
preheated straight soybean oil may be a practical replacement of the conventional diesel fuel with a small power and
efficiency drop. The brake thermal efficiency of the engine apparently increases with increased preheats temperature
of the soybean oil fuel and at 100oC it becomes very much comparable with the performance trends obtained using
diesel fuel.
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3. CONCLUSION
A comprehensive survey on First Law of Thermodynamics using diesel and biodiesel fuels is presented in this
literature. Extensive theory on the energy balance is given including the application of the first law of
thermodynamics, variation in heat transfer correlation for wall heat loss evaluation, thermodynamic models etc
Modelling and energy analysis, zero dimensional single zone combustion model simulation has been carried out to
predict the single cylinder constant speed diesel engine performance.




The single-zone thermodynamic models are comparatively easier to apply and exhibit reasonable accuracy
but for better accuracy multi-zone models are preferable.
The engine performance improved with low quantity blends of biodiesel to diesel, this indicated by the
higher maximum combustion temperature and pressure.
Biodiesels have lower thermal efficiency than diesel fuel .For same amount of output power, the BSFC of
biodiesels is much higher than diesel fuel. The heat losses except the exhaust loss are higher while using
biodiesels due to the presence of excessive oxygen molecules. Thermal insulation can be a good solution in
this regard; mean while it will increase the exhaust heat loss. The brake power decreases with the increase
of cetane number. The cooling water loss is comparatively higher than the exhaust heat loss for biodiesels.
Modifying the equations if necessary so that it could be applied over a much wider range of speed and load.
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p-ISSN: 2394-8280
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