**AIR STANDARD CYCLES**

**Theoretical
Analysis**

** **

The accurate analysis of the various processes taking place in an internal combustion engine is a very complex problem. If these processes were to be analyzed experimentally, the analysis would be very realistic no doubt. It would also be quite accurate if the tests are carried out correctly and systematically, but it would be time consuming. If a detailed analysis has to be carried out involving changes in operating parameters, the cost of such an analysis would be quite high, even prohibitive. An obvious solution would be to look for a quicker and less expensive way of studying the engine performance characteristics. A theoretical analysis is the obvious answer.

A
theoretical analysis, as the name suggests, involves analyzing the engine
performance without actually building and physically testing an engine. It
involves *simulating* an engine operation with the help of thermodynamics
so as to formulate mathematical expressions, which can then be solved in order
to obtain the relevant information. The method of solution will depend upon the
complexity of the formulation of the mathematical expressions, which in turn
will depend upon the assumptions that have been introduced in order to analyze
the processes in the engine. The more the assumptions, the simpler will be the
mathematical expressions and the easier the calculations, but the lesser will
be the accuracy of the final results.

The simplest theoretical analysis involves the use of the air standard cycle, which has the largest number of simplifying assumptions.

**A
Thermodynamic Cycle**

In some practical applications, notably steam power and refrigeration, a thermodynamic cycle can be identified.

A thermodynamic cycle occurs when the working fluid of a system experiences a number of processes that eventually return the fluid to its initial state.

In
steam power plants, water is pumped (for which work W_{P} is required)
into a boiler and evaporated into steam while heat Q_{A} is supplied
at a high temperature. The steam flows through a turbine doing work W_{T}
and then passes into a condenser where it is condensed into water with
consequent rejection of heat Q_{R} to the atmosphere. Since the water
is returned to its initial state, the net change in energy is zero, assuming no
loss of water through leakage or evaporation.

An
energy equation pertaining only to the system can be derived. Considering a
system with one entering and one leaving flow stream for the time period t_{1}
to t_{2}

_{}

ΔQ is the
heat transfer across the boundary, +ve for heat *added to* the system and

–ve for heat *taken from* the system.

ΔW is the
work transfer across the boundary, +ve for work *done by* the system and

-ve
for work *added to* the system

_{} is the energy of all
forms *carried* by the fluid across the boundary *into* the system

_{} is the energy of all
forms *carried* by the fluid across the boundary *out* of system

ΔE_{system}
is the energy of all forms *stored* within the system, +ve for energy *increase*

-ve
for energy *decrease*

In the case of the steam power system described above

_{}

All
thermodynamic cycles have a heat rejection process as an invariable
characteristic and the net work done is always less than the heat supplied,
although, as shown in Eq. 2, it is equal to the sum of heat added and the heat
rejected (Q_{R} is a negative number).

The
thermal efficiency of a cycle, η_{th}, is defined as the fraction
of heat supplied to a thermodynamic cycle that is converted to work, that is

_{}

This
efficiency is sometimes confused with the enthalpy efficiency, η_{e},
or the fuel conversion efficiency, η_{f}

_{}

This definition applies to combustion engines which have as a source of energy the chemical energy residing in a fuel used in the engine.

Any
device that operated in a thermodynamic cycle, absorbs thermal energy from a
source, rejects a part of it to a sink and presents the difference between the
energy absorbed and energy rejected *as work to the surroundings* is
called a heat engine.

A heat engine is, thus, a device that produces work. In order to achieve this purpose, the heat engine uses a certain working medium which undergoes the following processes:

1. A compression process where the working medium absorbs energy as work.

2. A heat addition process where the working medium absorbs energy as heat from a

source.

3 An expansion process where the working medium transfers energy as work to the surroundings.

4. A heat rejection process where the working medium rejects energy as heat to a sink.

If the working medium does not undergo any change of phase during its passage through the cycle, the heat engine is said to operate in a non-phase change cycle. A phase change cycle is one in which the working medium undergoes changes of phase. The air standard cycles, using air as the working medium are examples of non-phase change cycles while the steam and vapor compression refrigeration cycles are examples of phase change cycles.

**Air Standard Cycles**

The air standard cycle is a cycle followed by a heat engine which
uses air as the working medium. Since the air standard analysis is the simplest
and most idealistic, such cycles are also called *ideal cycles* and the
engine running on such cycles are called *ideal engines.*

In order that the analysis is made as simple as possible, certain assumptions have to be made. These assumptions result in an analysis that is far from correct for most actual combustion engine processes, but the analysis is of considerable value for indicating the upper limit of performance. The analysis is also a simple means for indicating the relative effects of principal variables of the cycle and the relative size of the apparatus.

**Assumptions**

1. The working medium is a perfect gas with constant specific heats and molecular weight corresponding to values at room temperature.

2. No chemical reactions occur during the cycle. The heat addition and heat rejection processes are merely heat transfer processes.

3. The processes are reversible.

4. Losses by heat transfer from the apparatus to the atmosphere are assumed to be zero in this analysis.

5. The working medium at the end of the process (cycle) is unchanged and is at the same condition as at the beginning of the process (cycle).

In selecting an idealized process one is always faced with the fact that the simpler the assumptions, the easier the analysis, but the farther the result from reality. The air cycle has the advantage of being based on a few simple assumptions and of lending itself to rapid and easy mathematical handling without recourse to thermodynamic charts or tables or complicated calculations. On the other hand, there is always the danger of losing sight of its limitations and of trying to employ it beyond its real usefulness.

**Equivalent
Air Cycle**

A particular air
cycle is usually taken to represent an approximation of some real set of
processes which the user has in mind. Generally speaking, the air cycle
representing a given real cycle is called an *equivalent air cycle*. The
equivalent cycle has, in general, the following characteristics in common with
the real cycle which it approximates:

1. A similar sequence of processes.

2. Same ratio of maximum to minimum volume for reciprocating engines or maximum to minimum pressure for gas turbine engines.

3. The same pressure and temperature at a given reference point.

4. An appropriate value of heat addition per unit mass of air.

**The Carnot
Cycle**

This cycle was proposed by Sadi Carnot in 1824 and has the highest possible efficiency for any cycle. Figures 1 and 2 show the P-V and T-s diagrams of the cycle.

**Fig. 1 Fig.
2**

** **

Assuming
that the charge is introduced into the engine at point 1, it undergoes
isentropic compression from 1 to 2. The temperature of the charge rises from T_{min}
to T_{max}. At point 2, heat is added isothermally. This causes the air
to expand, forcing the piston forward, thus doing work on the piston. At point
3, the source of heat is removed and the air now expands isentropically to
point 4, reducing the temperature to T_{min} in the process. At point
4, a cold body is applied to the end of the cylinder and the piston reverses,
thus compressing the air isothermally; heat is rejected to the cold body. At
point 1, the cold body is removed and the charge is compressed isentropically
till it reaches a temperature T_{max} once again. Thus, the heat
addition and rejection processes are isothermal while the compression and
expansion processes are isentropic.

From thermodynamics, per unit mass of charge

Heat
supplied from point 2 to 3_{}

Heat
rejected from point 4 to 1_{}

Now
p_{2}v_{2} = RT_{max} (7)

And
p_{4}v_{4} = RT_{min} (8)

Since Work done, per unit mass of charge, W = heat supplied – heat rejected

_{}_{}

_{}

We have assumed that the compression and expansion ratios are equal, that is

_{}

Heat
supplied Q_{s} = R T_{max} ln (r) (10)

Hence, the thermal efficiency of the cycle is given by

_{}

From Eq. 11 it is seen that the thermal efficiency of the Carnot cycle is only a function of the maximum and minimum temperatures of the cycle. The efficiency will increase if the minimum temperature (or the temperature at which the heat is rejected) is as low as possible. According to this equation, the efficiency will be equal to 1 if the minimum temperature is zero, which happens to be the absolute zero temperature in the thermodynamic scale.

This equation also indicates that for optimum (Carnot) efficiency, the cycle (and hence the heat engine) must operate between the limits of the highest and lowest possible temperatures. In other words, the engine should take in all the heat at as high a temperature as possible and should reject the heat at as low a temperature as possible. For the first condition to be achieved, combustion (as applicable for a real engine using fuel to provide heat) should begin at the highest possible temperature, for then the irreversibility of the chemical reaction would be reduced. Moreover, in the cycle, the expansion should proceed to the lowest possible temperature in order to obtain the maximum amount of work. These conditions are the aims of all designers of modern heat engines. The conditions of heat rejection are governed, in practice, by the temperature of the atmosphere.

It is impossible to construct an engine which will work on the Carnot cycle. In such an engine, it would be necessary for the piston to move very slowly during the first part of the forward stroke so that it can follow an isothermal process. During the remainder of the forward stroke, the piston would need to move very quickly as it has to follow an isentropic process. This variation in the speed of the piston cannot be achieved in practice. Also, a very long piston stroke would produce only a small amount of work most of which would be absorbed by the friction of the moving parts of the engine.

Since the efficiency of the cycle, as given by Eq. 11, is dependent only on the maximum and minimum temperatures, it does not depend on the working medium. It is thus independent of the properties of the working medium.

**Piston Engine
Air Standard Cycles**

The cycles described here are air standard cycles applicable to piston engines. Engines bases on these cycles have been built and many of the engines are still in use.

** **

**The Lenoir
Cycle**

The Lenoir cycle
is of interest because combustion (or heat addition) occurs without compression
of the charge. Figures 3 and 4 show the P-V and T-s diagrams**.**

**
Fig. 3 Fig. 4**

** **

According to the cycle, the piston is at the top dead center, point 1, when the charge is ignited (or heat is added). The process is at constant volume so the pressure rises to point 2. From 2 to 3, expansion takes place and from 3 to 1 heat is rejected at constant pressure.

Heat
supplied, Q_{s} = c_{v}(T_{2} – T_{1})
(12)

Heat
rejected, Q_{r} = c_{p}(T_{3 }– T_{1})
(13)

Since
W = Q_{s} - Q_{r} (14)

W =
c_{v}(T_{2} – T_{1}) – c_{p}(T_{2} – T_{1})
(15)

Thus
_{} (16)

_{} (17)

Since _{} and _{}

_{}

Here,
r_{e} = V_{3}/V_{1}, the volumetric expansion ratio.
Equation 18 indicates that the thermal efficiency of the Lenoir cycle depends
primarily on the expansion ratio and the ratio of specific heats.

The
intermittent-flow engine which powered the German V-1 buzz-bomb in 1942 during
World War II operated on a modified Lenoir cycle. A few engines running on the
Lenoir cycle were built in the late 19^{th} century till the early 20^{th}
century.

**The Otto
Cycle**

The Otto cycle,
which was first proposed by a Frenchman, Beau de Rochas in 1862, was first used
on an engine built by a German, Nicholas A. Otto, in 1876. The cycle is also
called a *constant volume* or *explosion* cycle. This is the
equivalent air cycle for reciprocating piston engines using spark ignition.
Figures 5 and 6 show the P-V and T-s diagrams respectively.

**Fig. 5 Fig. 6**

At the start of the cycle, the cylinder contains a mass M of air at the pressure and volume indicated at point 1. The piston is at its lowest position. It moves upward and the gas is compressed isentropically to point 2. At this point, heat is added at constant volume which raises the pressure to point 3. The high pressure charge now expands isentropically, pushing the piston down on its expansion stroke to point 4 where the charge rejects heat at constant volume to the initial state, point 1.

The isothermal heat addition and rejection of the Carnot cycle are replaced by the constant volume processes which are, theoretically more plausible, although in practice, even these processes are not practicable.

The
heat supplied, Q_{s}, per unit mass of charge, is given by

c_{v(}T_{3}
– T_{2})

the heat
rejected, Q_{r} per unit mass of charge is given by

c_{v(}T_{4}
– T_{1})

and the thermal efficiency is given by

_{}_{}

Now _{}

And
since _{}

Hence,
substituting in Eq. 19, we get, assuming that r is the compression ratio V_{1}/V_{2}

_{}

In
a true thermodynamic cycle, the term *expansion ratio* and *compression
ratio* are synonymous. However, in a real engine, these two ratios need not
be equal because of the valve timing and therefore the term *expansion ratio*
is preferred sometimes.

Equation 20 shows that the thermal efficiency of the theoretical Otto cycle increases with increase in compression ratio and specific heat ratio but is independent of the heat added (independent of load) and initial conditions of pressure, volume and temperature.

Figure 7 shows a plot of thermal efficiency versus compression ratio for an Otto cycle for 3 different values of γ. It is seen that the increase in efficiency is significant at lower compression ratios.

r

**Figure
7**

** **

This is also seen in Table 1 given below, with γ = 1.4.

Table 1

r |
h |

1 |
0 |

2 |
0.242 |

3 |
0.356 |

4 |
0.426 |

5 |
0.475 |

6 |
0.512 |

7 |
0.541 |

8 |
0.565 |

9 |
0.585 |

10 |
0.602 |

16 |
0.67 |

20 |
0.698 |

50 |
0.791 |

From the table it is seen that if:

CR is increased from 2 to 4, efficiency increase is 76%

CR is increased from 4 to 8, efficiency increase is only 32.6%

CR is increased from 8 to 16, efficiency increase is only 18.6%

** **

**Mean effective pressure**:

It is seen that the air standard efficiency of the Otto cycle depends only on the compression ratio. However, the pressures and temperatures at the various points in the cycle and the net work done, all depend upon the initial pressure and temperature and the heat input from point 2 to point 3, besides the compression ratio.

A quantity of special interest in reciprocating engine analysis is
the mean effective pressure. Mathematically, it is the net work done on the
piston, W, divided by the piston *displacement* volume, V_{1} – V_{2}.
This quantity has the units of pressure. Physically, it is that constant
pressure which, if exerted on the piston for the whole outward stroke, would
yield work equal to the work of the cycle. It is given by

_{}

where Q_{2-3}
is the heat added from points 2 to 3.

Now

_{}_{}

Here r is the compression ratio, V_{1}/V_{2}

From the equation of state:

_{}

R_{0} is the universal gas constant

Substituting for
V_{1} from Eq. 3 in Eq. 2 and then substituting for V_{1} – V_{2}
in Eq. 1 we get

_{}

The
quantity Q_{2-3}/M is the heat added between points 2 and 3 *per unit
mass *of air* *(M is the mass of air and m is the molecular weight of
air); and is denoted by Q’, thus

_{}_{}

We
can non-dimensionalize the mep by dividing it by p_{1} so that we can
obtain the following equation

_{}

Since_{}, we can substitute it in
Eq. 25 to get

_{}

The
dimensionless quantity mep/p_{1} is a function of the heat added,
initial temperature, compression ratio and the properties of air, namely, c_{v}
and γ. We see that the mean effective pressure is directly proportional to
the heat added and inversely proportional to the initial (or ambient)
temperature.

We can
substitute the value of η from Eq. 20 in Eq. 26 and obtain the value of
mep/p_{1} for the Otto cycle in terms of the compression ratio and heat
added.

**Figure
8**

** **

** **Figure 8 shows plots of mep/p_{1} versus compression ratio
for different values of heat added function.

In
terms of the pressure ratio, p_{3}/p_{2 }denoted by r_{p}
we could obtain the value of mep/p_{1} as follows:

_{}

We can obtain a
value of r_{p} in terms of Q’ as follows:

_{}

Another
parameter, which is of importance, is the quantity mep/p_{3}. This can
be obtained from the following expression:

_{}

**Figure
9**

Figure
9 shows plots of the quantity mep/p_{3} versus r. This shows a decrease
in the value of mep/p_{3} when r increases.

**Choice of Q’**

** **

We have said that

_{}

M is the mass of charge (air) per cycle, kg.

Now, in an actual engine

_{}

M_{f} is
the mass of fuel supplied per cycle, kg

Q_{c} is
the heating value of the fuel, kJ/kg

M_{a} is
the mass of air taken in per cycle

F is the fuel
air ratio = M_{f}/M_{a}

Substituting for Eq. (B) in Eq. (A) we get

_{}

_{}

So, substituting
for M_{a}/M from Eq. (33) in Eq. (32) we get

_{}

For isooctane,
FQ_{c} at stoichiometric conditions is equal to 2975 kJ/kg, thus

Q’ = 2975(r – 1)/r (35)

At an ambient
temperature, T_{1} of 300K and c_{v} for air is assumed to be
0.718 kJ/kgK, we get a value of Q’/c_{v}T_{1} = 13.8(r – 1)/r.

Under fuel rich
conditions, φ = 1.2, Q’/ c_{v}T_{1} = 16.6(r – 1)/r.

Under fuel lean
conditions, φ = 0.8, Q’/ c_{v}T_{1} = 11.1(r – 1)/r

**The Diesel
Cycle**

This cycle, proposed by a German engineer, Dr. Rudolph Diesel to describe the processes of his engine, is also called the constant pressure cycle. This is believed to be the equivalent air cycle for the reciprocating slow speed compression ignition engine. The P-V and T-s diagrams are shown in Figs 10 and 11 respectively.

Figures 10 and 11

The cycle has processes which are the same as that of the Otto cycle except that the heat is added at constant pressure.

The
heat supplied, Q_{s} is given by

c_{p(}T_{3}
– T_{2})

whereas the heat
rejected, Q_{r} is given by

c_{v(}T_{4}
– T_{1})

and the thermal efficiency is given by

_{}_{}

From the T-s diagram, Fig. 11, the difference in enthalpy between points 2 and 3 is the same as that between 4 and 1, thus

_{}

_{}

_{}

_{}and _{}

Substituting in eq. 36, we get

_{}

Now _{}

_{}

When Eq. 38 is compared with Eq. 20, it is seen that the expressions are similar except for the term in the parentheses for the Diesel cycle. It can be shown that this term is always greater than unity.

Now _{}where r is the
compression ratio and r_{e} is the expansion ratio

Thus, the thermal efficiency of the Diesel cycle can be written as

_{}

Let
r_{e} = r – Δ since r is greater than r_{e}. Here, Δ
is a small quantity. We therefore have

_{}

We can expand the last term binomially so that

_{}

Also
_{}

We can expand the last term binomially so that

_{}

Substituting in Eq. 39, we get

_{}

Since
the coefficients of _{},
etc are greater than unity, the quantity in the brackets in Eq. 40 will be
greater than unity. Hence, for the Diesel cycle, we subtract _{} times a quantity greater
than unity from one, hence for the same r, the Otto cycle efficiency is greater
than that for a Diesel cycle.

If _{} is small, the square,
cube, etc of this quantity becomes progressively smaller, so the thermal
efficiency of the Diesel cycle will tend towards that of the Otto cycle.

From the foregoing we can see the importance of cutting off the fuel supply early in the forward stroke, a condition which, because of the short time available and the high pressures involved, introduces practical difficulties with high speed engines and necessitates very rigid fuel injection gear.

In practice, the diesel engine shows a better efficiency than the Otto cycle engine because the compression of air alone in the former allows a greater compression ratio to be employed. With a mixture of fuel and air, as in practical Otto cycle engines, the maximum temperature developed by compression must not exceed the self ignition temperature of the mixture; hence a definite limit is imposed on the maximum value of the compression ratio.

Thus Otto cycle engines have compression ratios in the range of 7 to 12 while diesel cycle engines have compression ratios in the range of 16 to 22.

We can obtain a value of r_{c} for a Diesel cycle in terms
of Q’ as follows:

_{}

We can
substitute the value of η from Eq. 38 in Eq. 26, reproduced below and
obtain the value of mep/p_{1} for the Diesel cycle.

_{}

In terms of the cut-off ratio, we can obtain another expression for
mep/p_{1} as follows:

_{}

For the Diesel cycle, the expression for mep/p_{3} is as
follows:

_{}

Modern high speed diesel engines do not follow the Diesel cycle. The process of heat addition is partly at constant volume and partly at constant pressure. This brings us to the dual cycle.

**The Dual
Cycle**

An important
characteristic of real cycles is the ratio of the mean effective pressure to
the maximum pressure, since the mean effective pressure represents the useful
(average) pressure acting on the piston while the maximum pressure represents
the pressure which chiefly affects the strength required of the engine
structure. In the constant-volume cycle, shown in Fig. 10, it is seen that the
quantity mep/p_{3} falls off rapidly as the compression ratio
increases, which means that for a given mean effective pressure the maximum
pressure rises rapidly as the compression ratio increases. For example, for a
mean effective pressure of 7 bar and Q’/c_{v}T_{1} of 12, the
maximum pressure at a compression ratio of 5 is 28 bar whereas at a compression
ratio of 10, it rises to about 52 bar. Real cycles follow the same trend and it
becomes a practical necessity to limit the maximum pressure when high
compression ratios are used, as in diesel engines. This also indicates that
diesel engines will have to be stronger (and hence heavier) because it has to
withstand higher peak pressures.

Constant
pressure heat addition achieves rather low peak pressures unless the
compression ratio is quite high. In a real diesel engine, in order that
combustion takes place at constant pressure, fuel has to be injected very late
in the compression stroke (practically at the top dead center). But in order to
increase the efficiency of the cycle, the fuel supply must be cut off early in
the expansion stroke, both to give sufficient time for the fuel to burn and
thereby increase combustion efficiency and reduce after burning but also reduce
emissions. Such situations can be achieved if the engine was a slow speed type
so that the piston would move sufficiently slowly for combustion to take place
despite the late injection of the fuel. For modern high speed compression
ignition engines it is not possible to achieve constant pressure combustion.
Fuel is injected somewhat earlier in the compression stroke and has to go
through the various stages of combustion. Thus it is seen that combustion is
nearly at constant volume (like in a spark ignition engine). But the peak
pressure is limited because of strength considerations so the rest of the heat
addition is believed to take place at constant pressure in a cycle. This has
led to the formulation of the dual combustion cycle. In this cycle, for high
compression ratios, the peak pressure is not allowed to increase beyond a
certain limit and to account for the total addition, the rest of the heat is
assumed to be added at constant pressure. Hence the name *limited pressure
cycle.*

The cycle is the equivalent air cycle for reciprocating high speed compression ignition engines. The P-V and T-s diagrams are shown in Figs.12 and 13. In the cycle, compression and expansion processes are isentropic; heat addition is partly at constant volume and partly at constant pressure while heat rejection is at constant volume as in the case of the Otto and Diesel cycles.

**Figure 12 Figure 13**

The
heat supplied, Q_{s} per unit mass of charge is given by

c_{v(}T_{3}
– T_{2}) + c_{p(}T_{3’} – T_{2})

whereas the heat
rejected, Q_{r} per unit mass of charge is given by

c_{v(}T_{4}
– T_{1})

and the thermal efficiency is given by

_{}

From thermodynamics

_{} the explosion or
pressure ratio and

_{} the cut-off ratio.

Now,
_{}

Also
_{}

And _{}

Thus
_{}

Also
_{}

Therefore, the thermal efficiency of the dual cycle is

_{}

We
can substitute the value of η from Eq. 46 in Eq. 26 and obtain the value
of mep/p_{1} for the dual cycle._{ }

In
terms of the cut-off ratio and pressure ratio, we can obtain another expression
for mep/p_{1} as follows:

_{}

For the dual cycle, the
expression for mep/p_{3} is as follows:

_{}

Since the dual cycle is also called the limited pressure cycle, the peak
pressure, p_{3}, is usually specified. Since the initial pressure, p_{1},
is known, the ratio p_{3}/p_{1} is known. We can correlate r_{p}
with this ratio as follows:

_{}

We
can obtain an expression for r_{c} in terms of Q’ and r_{p} and
other known quantities as follows:

_{}

We
can also obtain an expression for r_{p} in terms of Q’ and r_{c}
and other known quantities as follows:

_{}

Figure 14 shows a constant volume and a constant pressure cycle, compared with a limited pressure cycle. In a series of air cycles with varying pressure ratio at a given compression ratio and the same Q’, the constant volume cycle has the highest efficiency and the constant pressure cycle the lowest efficiency.

Figure
15 compares the efficiencies of the three cycles for the same value of _{} for the same initial
conditions and three values of p_{3}/p_{1} for the dual cycle.
It is interesting to note that the air standard efficiency is little affected
by compression ratio above a compression ratio of 8 for the limited pressure
cycle.

The
curves of mep/p_{3} versus compression ratio for the same three cycles
as above are given in Fig. 10. It is seen that a considerable increase in this
ratio is obtained for a limited pressure cycle as compared to the constant
volume or constant pressure cycles.

**Figure
14**

**Figure
15**

** **

**The Atkinson
Cycle**

This cycle is
also referred to as the *complete expansion cycle.* Inspection of the P-V
diagrams of the Otto, Diesel and Dual cycles shows that the expansion process
to point 4 does not reach the lowest possible pressure, namely, atmospheric
pressure. This is true of all real engines; when the exhaust valve opens, the
high pressure gases undergo a violent blow down process with consequent
dissipation of available energy. This is necessary so as to allow the gases to
flow out due to pressure difference and hence reduce the piston work in driving
out the gases. The air standard cycle shows a loss of net work because of the
reduction in area of the P-V diagram.

In the Otto cycle, if the expansion is allowed to completion to point 4’ (Fig. 16) and heat rejection occurs at constant pressure, the cycle is called the Atkinson cycle.

The heat
supplied, Q_{s} per unit mass of charge is given by

c_{v(}T_{3}
– T_{2})

**Figure
16**

whereas the heat
rejected, Q_{r} per unit mass of charge is given by

c_{p(}T_{4}
– T_{1})

and the thermal efficiency is given by

_{}

_{}_{}

Now _{}

As
before, _{}

And _{}

The efficiency is therefore given by

_{}

If
we denote the expansion ratio as V_{4’}/V_{3}, we can rewrite
the thermal efficiency as

_{}

Since the Atkinson cycle area under the P-V diagram is larger than the corresponding Otto cycle, the efficiency, for the same compression ratio and heat input, will be higher.

An engine can be built to make use of complete expansion, but the stroke length of such an engine will be extremely long and will not be economically feasible to offset the improvement in power and efficiency. Also, there are some operational problems with such a cycle.

**The Miller
Cycle**

This cycle, proposed by Ralph Miller, (Fig. 17), is applicable for engines with very early or late closing of the inlet valve. If the valve closes before the piston reaches bottom center, at point 1, the charge inside will first expand to point 7. Compression will be from point 7 through 1 to point 2. Work done in expansion from 1 to 7 is the same as the compression work from point 7 to 1. The actual compression work will be from 1 to 2.

If the valve closes after the piston crosses the bottom center, it will do so again at point 1. Compression will begin after the valve closes. For this case, process 1 to 7 and 7 to 1 will not exist.

The
parameter, λ, is defined as the ratio of the expansion ratio r_{e}
to the compression ratio, r_{c}, thus:

**Fig. 17****Fig.
17**

(Equations 58-61)

The thermal efficiency of the Miller cycle is a function of the compression ratio, the specific heat ratio, the expansion ratio and the heat added. The ratio of the Miller cycle thermal efficiency and the equivalent Otto cycle thermal efficiency is plotted against λ in Fig. 18 (Taken from Ferguson and Kirkpatrick[1]). For high values of λ, the Miller cycle is more efficient. A plot of the ratio of indicated mean effective pressures of the two cycles against λ (Fig.19, also taken from the same reference) shows that the Miller cycle is at a significant disadvantage. This is because, as λ increases, the fraction of the displacement volume that is filled with the inlet fuel-air mixture decreases, thereby decreasing the IMEP. The decrease in the IMEP for the Miller cycle can be compensated by supercharging the inlet mixture.

**Figure
18**

**Figure
19**

** **

**The Brayton
Cycle**

The Brayton cycle is also referred to as the Joule cycle or the gas turbine air cycle because all modern gas turbines work on this cycle. However, if the Brayton cycle is to be used for reciprocating piston engines, it requires two cylinders, one for compression and the other for expansion. Heat addition may be carried out separately in a heat exchanger or within the expander itself.

The pressure-volume and the corresponding temperature-entropy diagrams are shown in Figs 20 and 21 respectively.

**Fig.
20**

**Fig.
21**

** **

The cycle consists of an isentropic compression process, a constant pressure heat addition process, an isentropic expansion process and a constant pressure heat rejection process. Expansion is carried out till the pressure drops to the initial (atmospheric) value.

Heat
supplied in the cycle, Q_{s}, is given by

C_{p}(T_{3}
– T_{2})

Heat
rejected in the cycle, Q_{s}, is given by

C_{p}(T_{4}
– T_{1})

Hence the thermal efficiency of the cycle is given by

_{}

Now _{}

And
since _{}

Hence,
substituting in Eq. 62, we get, assuming that r_{p} is the pressure
ratio p_{2}/p_{1}

_{}

This is numerically equal to the efficiency of the Otto cycle if we put

_{}

so
that _{}

where r is the volumetric compression ratio.

For
gas turbines it is convenient to speak of pressure ratio p_{2}/p_{1}
rather than the compression ratio V_{2}/V_{1} unless we are
talking of a reciprocating type of Brayton cycle engine. Reciprocating engines
that operate on this cycle would require a very long stroke so that the working
medium can expand to atmospheric pressure. This will increase the friction
power and hence reduce the brake power.

The heat addition at constant pressure of the Brayton cycle makes it more efficient than the diesel cycle although the latter also has a constant pressure heat addition. This is because expansion in the former cycle proceeds to atmospheric pressure rather than to a higher pressure in the former cycle.

The spark ignition and compression ignition engines are more efficient than the gas turbine. This is because the SI and the CI engines operate at higher peak cycle temperatures. Moreover, the compression and expansion processes are more efficient in the piston-cylinder system due to lower fluid friction and turbulence. On the other hand, the mass flow rate through a gas turbine is much greater than that through a CI or SI engine; hence the gas turbine is ideally suited for higher power than the CI engine. The gas turbine may be provided with intercooling during compression, reheating during expansion, and regeneration prior to heat addition. These are techniques used to increase the power and efficiency of a simple gas turbine.

Gas
turbines generally run at maximum fuel-air ratios that are about a quarter of
the chemically correct ratio. Hence, such cycles analysis may be carried out
with Q’ = 2980/4 = 745 kJ/kg air. There is no concept of a clearance volume in
a gas turbine so the value of M_{a}/M in eq. 32 is taken as unity.

For
a gas turbine, the ratio of work per unit time (or power) to the volume of air
at inlet conditions (per unit time) or W/V_{1} has units of pressure.
Its significance is similar to that of mean effective pressure in reciprocating
engines.

A gas turbine cycle of the type described above, at the most, gives an idea of the upper limit of possible cycle efficiency. It does not, however, predict the trends of real gas turbine performance very well, even when the compressor, combustion chamber and turbine efficiencies are assumed to be constant.

**Comparison of
Air Cycles**

** **

** **

**Fig 22**

The Lenoir, Otto, Diesel, dual, Atkinson and Brayton cycles may be compared for similar parametric conditions. In all these comparisons it is assumed that initial conditions of pressure and temperature are identical. One of the parameters that may be kept the same for all the cycles is the heat input. Another factor that may be kept the same is the compression ratio, maximum pressure or maximum temperature. Another set of factors that can be kept the same are the temperature and pressure. In the case of the dual cycle, another comparison is on the basis of different proportions of heat added at constant pressure relative to heat added at constant volume.

**Same
Compression Ratio and Heat Input**

It is already seen that adding heat at constant volume results in the highest maximum pressure and temperature for the Otto and Atkinson cycles. Adding heat at constant pressure results in the lowest maximum temperature and pressure for the constant pressure Diesel cycle. The corresponding values for the dual cycle lie in between those for the Otto and Diesel cycles.

This is seen Fig. 22, which gives the pressure-volume diagrams for all cycles except the Lenoir cycle which has no compression therefore no compression ratio. The Lenoir cycle (with the same heat input) has a peak temperature between the Otto and Diesel cycles but a displacement volume of about 6.6 times that of the other cycles.

**Fig. 23**

In the temperature-entropy diagram shown in Fig. 23, the area under a curve represents the heat (added or rejected as the case may be). For the Otto, dual and Diesel cycles, the area under the lower constant volume line represents the heat rejected, between the entropy limits of any given cycle. From Fig. 23, it is seen that the area under the heat rejected curve is the least for the Otto cycle and the highest for the Diesel cycle, while for the limited pressure (dual) cycle, it lies in between. Since the heat rejected by the Otto cycle is the lowest, and

_{},

it the most efficient while the Diesel is the least efficient because in this cycle, the heat rejected is the highest. The efficiency of the dual cycle lies in between. This explains why a petrol engine will be more efficient than a diesel engine if the compression ratio is the same. Unfortunately, a diesel engine cannot have the same compression ratio as that of a petrol engine because the diesel fuel would not be able to auto-ignite. However, it is clear from the foregoing in any engine, the addition of heat should be such that maximum possible expansion of the working fluid should occur in order to obtain the maximum thermal efficiency.

** **

** **When comparing the constant-pressure heat rejection curves of the
Brayton, Atkinson and Lenoir cycles, the heat rejection is the highest for the
Lenoir and hence its efficiency is the lowest. Thus, the relative values of the
heat rejection (in ascending order) and the corresponding thermal efficiency (in
descending order) are as follows: Atkinson, Otto, Brayton (numerically equal to
that of the Otto), dual, Diesel, and Lenoir cycles. The main reason why the
diesel cycle is at a disadvantage is its lower isentropic expansion ratio.

**Fig. 24**

**Same Maximum
Pressure and Heat Input**

A comparison of all the cycles except the Lenoir air cycle with the same maximum pressure and heat input, as seen in Fig. 24, indicates that the heat rejected is the lowest for the Brayton cycle and the highest for the Otto cycle. The relative order of efficiencies (in ascending order is as follows: Otto, Atkinson, dual, diesel, and Brayton cycles. The Lenoir cycle would not be feasible because the temperatures reached at the end of combustion would be extremely high. A much lower compression ratio must be used with the Otto and Atkinson cycles than with the Brayton and diesel cycles in order to attain the same maximum pressure. The compression ratio for the dual cycle will lie in between.

This explains why the diesel engine (which follows the dual cycle more closely) is more efficient than a petrol engine. In a real engine, the maximum pressures would be comparable and so also the heat inputs because the heating values of the two fuels are more or less similar.

**Same Maximum Temperature and Heat Input**

The same conclusion as in the previous case can be obtained in this case, that is, the Otto cycle is the least efficient and the Brayton is the most efficient. Here also the compression ratio of the Otto and Atkinson cycles will have to be kept much lower than that of the Brayton and Diesel cycles. The dual cycle case falls in between.

**Same Maximum
Pressure and Maximum Temperature**

For this case,
the heat rejected for those cycles where the heat is rejected at constant
volume is the same. ** **However, the heat added in the diesel cycle is the
highest, making it the most efficient cycle, followed by the dual and Otto
cycles. The heat rejected for those cycles where the heat is rejected at
constant pressure will be lower, and since the heat added in the Brayton cycle
is higher than that for the Atkinson cycle, the Brayton cycle is the most
efficient of them all, followed by the Atkinson, Diesel, dual and Otto cycles.
The compression ratio in the Diesel cycle will be higher than that of the Otto.

**Same Maximum
Pressure and Output**

While the temperature-entropy plots are best suited for comparing cycles on the basis of heat input and temperatures, the pressure-volume diagram is best suited for comparing cycles on the basis of pressure and work output. The temperature-entropy diagram would nevertheless be still required in order to determine the efficiency. The temperature-entropy curves will be similar to the case of same maximum pressure and same heat input. Hence the order of efficiency will be the same, that is Otto, Atkinson, dual, Diesel, and Brayton cycles.

**Additional
Information on the Miller Cycle**

Taken from Everything2.com:-

The Miller Cycle, developed by American engineer Ralph Miller in the 1940's, is a modified Otto Cycle that improves fuel efficiency by 10%-20%. It relies on a supercharger/turbocharger, and takes advantage of the superchargers greater efficiency at low compression levels. As with other forced induction engines more power can be had from a smaller engine, but without the efficiency penalties usually associated with forced induction (e.g. a Miller Cycle v6 can get the power of a v8 yet still retain the fuel efficiency of a v6).

During the intake stroke the supercharger overcharges the cylinder with fuel and air, and during the first bit of the compression stroke the intake valves are left open and some of the overcharge is pushed out. While the overcharge is being forced out and until the intake valves close the piston isn't pushing against anything and in effect the compression stroke is shortened compared to the 'normal' power stroke. While the supercharger normally employed does use some of the engines power, it's much less than the power saved from the shortened compression stroke. The lower friction associated with the smaller engine also improves efficiency

Taken from Wikipedia:-

A traditional Otto cycle engine uses four "strokes", of which two can be considered "high power" – the compression stroke (high power consumption) and power stroke (high power production). Much of the internal power loss of an engine is due to the energy needed to compress the charge during the compression stroke, so systems that reduce this power consumption can lead to greater efficiency.

In the Miller cycle, the intake valve is left open longer than it would be in an Otto cycle engine. In effect, the compression stroke is two discrete cycles: the initial portion when the intake valve is open and final portion when the intake valve is closed. This two-stage intake stroke creates the so called "fifth" cycle that the Miller cycle introduces. As the piston initially moves upwards in what is traditionally the compression stroke, the charge is partially expelled back out the still-open intake valve. Typically this loss of charge air would result in a loss of power. However, in the Miller cycle, this is compensated for by the use of a supercharger. The supercharger typically will need to be of the positive displacement type due its ability to produce boost at relatively low engine speeds. Otherwise, low-rpm torque will suffer.

A key aspect of the Miller cycle is that the compression stroke actually starts only after the piston has pushed out this "extra" charge and the intake valve closes. This happens at around 20% to 30% into the compression stroke. In other words, the actual compression occurs in the latter 70% to 80% of the compression stroke. The piston gets the same resulting compression as it would in a standard Otto cycle engine for less work.

To understand the reason for the delay in closing the intake valve, consider the action of the crankshaft, piston and connecting rod in creating a mechanical advantage. At bottom dead center ("BDC") or top dead center ("TDC"), the rotational axis of the crank comes into alignment with the wrist pin, and the big end of the crank. When these three points (rotational axis of the crank, wrist pin center, and big end center) are in alignment, there is no lever arm to create or use rotational energy. But as the crank rotates a bit, the big end of the crank moves away from alignment with the other two points, creating the mechanical leverage needed to do the work of compression. By delaying the closing of the inlet port, compression of the air in the cylinder is delayed to a point where the crankshaft is once again very effective. In the meantime, the air charge has been easily pushed out of the cylinder and back upstream in the inlet tract where it meets the pressurized charge from the supercharger head-on, causing the inlet pressure to increase just as the inlet port closes. In the inlet tract, the supercharger continues to add pressure until the inlet valve opens again. The net gain comes from moving the work of compression away from the most inefficient region of the crank rotation, namely the rotation near BDC, and letting the work of compression be done during the near-BDC period by the more efficient Supercharger. This trick of inlet timing and compression allows the crank to turn freely around BDC and makes Miller Cycle engines free revving and fuel efficient.

The Miller cycle results in an advantage as long as the supercharger can compress the charge using less energy than the piston would use to do the same work. Over the entire compression range required by an engine, the supercharger is used to generate low levels of compression, where it is most efficient. Then, the piston is used to generate the remaining higher levels of compression, operating in the range where it is more efficient than a supercharger. Thus the Miller cycle uses the supercharger for the portion of the compression where it is best, and the piston for the portion where it is best. In total, this reduces the power needed to run the engine by 10% to 15%. To this end, successful production engines using this cycle have typically used variable valve timing to effectively switch off the Miller cycle in regions of operation where it does not offer an advantage.

In a typical spark ignition engine, the Miller cycle yields an additional benefit. The intake air is first compressed by the supercharger and then cooled by an intercooler. This lower intake charge temperature, combined with the lower compression of the intake stroke, yields a lower final charge temperature than would be obtained by simply increasing the compression of the piston. This allows ignition timing to be advanced beyond what is normally allowed before the onset of detonation, thus increasing the overall efficiency still further.

An additonal advantage of the lower final charge temperature is that the emission of NOx in diesel engines is decreased, which is an important design parameter in large diesel engines on board ships and power plants.

Efficiency is increased by raising the compression ratio. In a typical gasoline engine, the compression ratio is limited due to self-ignition (detonation) of the compressed, and therefore hot, air/fuel mixture. Due to the reduced compression stroke of a Miller cycle engine, a higher overall cylinder pressure (supercharger pressure plus mechanical compression) is possible, and therefore a Miller cycle engine has better efficiency.

The benefits of utilizing positive displacement superchargers come with a cost. 15% to 20% of the power generated by a supercharged engine is usually required to do the work of driving the supercharger, which compresses the intake charge (also known as boost).

[1]Ferguson and Kirkpatrick, “Internal Combustion
Engines”, 2^{nd} Ed., John Wiley & Sons New York, 2001