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EFFICIENCY OF COMBUSTION
Nico Woudstra
Delft University of Technology
Thermal Power Engineering
Mekelweg 2, 2628 CD Delft
The Netherlands
Phone: +31 15 278 2178 Fax: +31 15 278 2460
E-mail:
n.woudstra@wbmt.tudelft.nl
ABSTRACT
From second law evaluations (entropy or exergy evaluations) it appears that thermodynamic
losses of boilers and furnaces are much higher than the thermal efficiencies do suggest. With
thermal losses of around 5 % the thermodynamic losses (exergy losses) of a boiler can be 50 % or
more. The combustion process is responsible for a significant part of these losses. Various
atmospheric combustion processes are investigated in order to obtain a better insight in the effect
of the main design parameters like the applied fuel, its composition and moisture content, the air
factor and the air preheat temperature. Therefore a series of system calculations with natural gas,
hard coal and wood as fuel have been made. Value diagrams have been derived to visualise the
exergy loss of combustion and to explain their cause. For the investigations the computer
program Cycle-Tempo has been used to perform the necessary system calculations, to draw value
diagrams and to analyse the results. It appears that the type of fuel and the extent of air preheat
mainly determine the exergy loss of combustion. Preheating of combustion air can reduce the
exergy losses of combustion considerably. But even with air preheat temperatures up to 1000 ˚C
exergy losses of combustion are still high.
Keywords:
atmospheric combustion, biomass, coal, natural gas, Cycle Tempo, exergy
analysis, exergy efficiencies, value diagrams.
NOMENCLATURE
100 %, can be achieved depending on the applied
fuel and boiler type. These high thermal efficiencies
do suggest that combustion processes are highly
optimised and do not need further improvement with
regard to their thermodynamic performance. Second
law (entropy or exergy) evaluations however show
that thermodynamic losses of boilers and furnaces
are much lager than the thermal efficiencies do
suggest. With thermal losses of around 5 % the
thermodynamic losses still can be in the order of 30
%. These high thermodynamic losses (exergy losses)
are mainly caused by the combustion process as can
be demonstrated by the value diagram in Figure 3.1;
the shaded areas represent the thermodynamic
losses. The backgrounds of the value diagram are
explained in [1].
ex
F
specific exergy fuel [kJ/kg]
Ex
exergy [kJ]
LHV
F
lower heating value fuel [kJ/kg]
Q
heat [kJ]
T
temperature [K]
T
0
environmental temperature [K]
W
power [kJ]
INTRODUCTION
The production of power and heat in
industrialised countries is almost entirely based
on the combustion of fuels. Usually combustion
takes place in boilers or furnaces; well designed
boilers have high thermal efficiencies of more
than 90 %. Even very high efficiencies, close to
Because of availability and costs air is usually
applied for the combustion of fuels. In case of
fossil fuels adiabatic combustion temperatures are
around 2000 ˚C. Heat can be made available by
cooling down the flue gas from this temperature
will not automatically result in higher overall plant
efficiencies. In case of power production, heat must
be transferred to a power cycle. When the
thermodynamic average temperature, at which heat
transfer to the power cycle takes place, remains
unchanged, reducing the thermodynamic losses of
combustion will primarily increase the
thermodynamic loss of heat transfer from flue gas to
the power cycle. Therefore measures to decrease the
thermodynamic losses of combustion are useful only
if they are accompanied with improved conditions of
the heat absorbing processes.
T
0
1 -
T
T
stack
PRELIMINARY CONSIDERATIONS
In actual combustion plants it must be assured that
100 % fuel conversion will take place. To enable
complete conversion within a limited time the
amount of air should be higher than the
stoichiometric quantity. However the air factor, the
ratio between the actual air quantity divided by the
stoichiometric quantity, should be as low as possible
since the excess air reduces the adiabatic
combustion temperature and consequently the
thermodynamic average temperature of heat
transfer. The actual air factor depends on the type of
fuel as well as the design of the combustion plant.
Solid fuels will require higher air factors than
gaseous fuels. The required air factor is further
depending on combustion temperature, residence
Q
ex
F
flue gas
Q
Figure 3.1 Value diagram of adiabatic
combustion of a fuel
to ambient temperature. The resulting
temperature curve is shown in Figure 3.1; the
area below this curve equals the exergy of the
heat transferred from the flue gas. As the total
area of this diagram equals the exergy of the fuel,
the slantly shaded area represents the exergy loss
of combustion. In a boiler however flue gas is not
cooled down to ambient temperature, but leaves
the boiler at the “stack temperature” (see Figure
3.1). The residual heat from the flue gas is also
lost; in Figure 3.1 the horizontal shaded area
equals the corresponding exergy loss. From the
value diagram it will be clear that the rather low
combustion temperature and the need for cooling
down the flue gasses in order to extract the
required heat, are the main causes of the large
exergy losses. The thermodynamic average
temperature of heat transfer from the flue gasses
is around 1000 ˚C. The adiabatic combustion
temperature and thus also the thermodynamic
efficiency can be raised seriously by using
oxygen instead of air to oxidise the fuel. However
in general this option is not feasible because of
thermodynamic losses and high costs of oxygen
separation plants. When using air, the adiabatic
combustion temperature depends only on the
properties of fuel and air. The determining
parameters are the fuel type, their composition
and moisture content, the air temperature and air
factor at combustor inlet.
Reducing the thermodynamic losses of
combustion in atmospheric combustion systems
Figure 3.2 Combustion system with air preheat
time of the reactants in the combustion chamber or
furnace and the way of mixing, the turbulence, of
fuel and air flows. An air factor of 1.05 and 1.10 is
sufficient for gaseous fuels, while solid fuels will
require air factors of 1.10 to 1.20.
Air preheating makes it possible to increase the
temperature of combustion and consequently the
thermodynamic average temperature of heat
transfer will be increased. Assuming a
combustion system as shown in Figure 3.2, air is
preheated before it is supplied to the combustion
chamber. Hot flue gasses are passed to the heat
transfer system to supply heat to a power cycle or
other processes after adiabatic combustion. The
flue gas is further cooled down in the air
preheater and finally discharged to the stack.
1
T
0
1 -
T
0
1
Q
af
ex
F
Figure 3.4 Value diagram of a combustion system
with air preheat
(after removal of the air preheater)
T
0
1 -
T
In actual systems heat transfer in the preheater will
require a substantial temperature difference between
air and flue gas. And thus the exergy loss of the
preheater will not be negligible. In Figure 3.5 the
temperature curve of the heated air and thus also the
additional exergy loss of the air preheater is
included. The balance of the exergy loss reduction
of combustion and the additional exergy loss of the
air preheater determines the overall effects of air
preheat.
0
Q
air preheat
Q
air preheat
Q
from system
exergy loss of combustion
Figure 3.3 Value diagram of a combustion
system with air preheat
The resulting cooling curve of the flue gas is
shown in the value diagram of Figure 3.3. In this
diagram it is supposed that heat transfer in the
preheater can occur without temperature
difference between air and flue gas; this is of
course unrealistic but simplifies the explanation
of the effect. The available heat from the flue gas
equals the sum of the heating value of the fuel
and the heat of the preheated air. Therefore the
value diagram of Figure 3.3 can be obtained by
extending the horizontal axis of the value
diagram of combustion without reheat (see Figure
3.1) with the heat from the preheated air and
extrapolating the flue gas temperature curve to
the point of elevated combustion temperature.
The exergy loss of combustion with air preheat is
represented by the slantly shaded area in Figure
3.3. Heat transfer in the air preheater occurs
between two flows within the system. By
removing the air preheater from the diagram of
Figure 3.3 the value diagram as shown in Figure
3.4 is obtained. From this diagram it will be clear
that air preheat reduces the exergy loss of
combustion. The difference in exergy loss
between the system with reheat and the system
without reheat is also shown in Figure 3.5.
1
T
0
1 -
T
0
Q
from system
ex
F
Figure 3.5 Value diagram with exergy loss
reduction due to air preheat and
additional exergy loss of the preheater
SYSTEM EVALUATION
For evaluating various combustion options three
system models are specified, that are applied to
calculate system performance with Cycle-Tempo, a
computer program for the evaluation and
optimisation of energy systems. Since understanding
of the effect of design parameters on the combustion
process is the main objective of this evaluation,
technical limitations are sometimes ignored.
(nr.2). Heat is transferred in the air preheater from
the flue gas flow to the air flow. As the flue gas flow
is stronger than the air flow, the temperature
difference between flue gas and air will become
smaller when increasing the air temperature. The
minimum allowable temperature difference is set at
40 K. With this temperature difference at the hot
side of the preheater the maximum air preheat
temperature is applied. The highest air temperature
that can be obtained depends on the type of fuel as
well as the air factor. To allow for higher air
preheating temperatures the flue gas flow must be
splitted so that the flow that passes the air preheater
is just sufficient to deliver the necessary heat for
preheating the combustion air. The remainder of the
4
3
3
1
1
2
2
4
7
5
6
5
6
H
8
7
8
Figure 3.6 Combustion system without air
preheat
The first system model is a simple combustion
system without air preheat as shown in Figure
3.6. An air blower (nr.2) passes the air to the
combustion chamber (nr.3). The combustion
chamber is assumed to be adiabatic; then the flue
gas leaves the combustion chamber at high
temperature. Hot flue gas is cooled from
adiabatic combustion temperature to 100 ˚C in
heat exchanger (nr.5) and discharged to the
atmosphere via the stack (nr.7) by the forced draft
fan (nr.6). Thus the flue gas temperature at the
inlet of the forced draft fan is 100 ˚C; this
temperature is applied for all system alternatives
disregarding the type of fuel. This enables a
useful comparison of the performance of the
4
4
3
3
5
114
7
2
5
112
8
111
12
8
7
10
6
H
H
110
11
9
H
2
115
9
1
1
6
113
Figure 3.8 Combustion system with splitted
flue gas flow for air preheat
flue gas flow can be cooled in another heat
exchanger (nr. 11 in Figure 3.8), where heat is
transferred to a power cycle or other heat absorbing
processes. In theory any desired air temperature can
be obtained by this system. Without considering
limitations with respect to the technical feasibility of
these systems it is assumed that air temperatures up
to 1000 ˚C can be achieved. For all air temperatures
the temperature difference between the flows at the
hot side of the air preheater is set at 40 K. To
evaluate the effect of the type of fuel three fuels are
considered: natural gas, hard coal and biomass
(wood). The composition of Slochteren gas has been
chosen for the natural gas. The composition of the
fuels (ultimate analysis for the solid fuels) as fed to
the combustor is shown in Table 3.1. Also the lower
heating value (
LHV
F
) and the exergy (
ex
F
) of the fuel
are given. Estimated exergy values of the solid fuels
are based on [2]. Also the composition of the
environment chosen for determining exergy values
is based on this reference. For the evaluation
complete combustion is assumed for all considered
air factors; also the stack temperature is supposed to
be the same for all alternatives.
4
4
3
3
5
7
2
5
8
8
7
6
H
H
110
2
9
9
6
1
1
Figure 3.7 Combustion system with air preheat
system alternatives. As the utilisation of the heat
transferred in heat exchanger (nr.5) is beyond the
scope of this evaluation, this part of the system is
only specified to remove the transferred heat.
The second system model represents a
combustion system with air preheat and is shown
in Figure 3.7. The only difference with the first
model (without air preheat) is the air preheater
stack. In addition to the exergy of the fuel, the
electric motor driven air blower and forced draft fan
also supply exergy to the system. For the reference
case the electrical input of the motors is 0.35 % of
the fuel exergy. Therefore the total exergy loss of
combustor, air preheater and stack is a little bit
higher than the apparent loss resulting from the
system exergy efficiency.
The value diagram of the reference case is shown in
Figure 3.9. The diagram differs slightly from the
diagrams discussed previously as the length of the
horizontal axis does not equal the exergy of the fuel
but the quantity of heat transferred from the flue gas
by cooling the flue gas from adiabatic combustion
temperature till ambient temperature. As not all
natural gas solid fuels
comp. mol.fract. elem.+ mass fractions
comp. coal wood
CH
4
0.8129 C 0.5990 0.4496
C
2
H
6
0.0287 H
2
0.0534 0.0475
C
3
H
8
0.0038 N
2
0.0115 0.0148
C
4
H
10
0.0015 O
2
0.1694 0.2361
C
5
H
12
0.0004 S 0.0135 0.0011
C
6
H
14
0.0005 Cl
2
0.0030 -
N
2
0.1432 F
2
0.0002 -
O
2
0.0001 H
2
O - 0.2000
CO
2
0.0089 ash 0.1500 0.0509
total 1.0000 1.0000 1.0000
LHV
F
38.00 24.61 16.01
(MJ/kg)
ex
F
39.39 26.47 18.06
(MJ/kg)
Table 3.1 Composition of applied fuels
Value diagram
Heat Exchgr. 5
600
600
500
500
400
400
DESCRIPTION REFERENCE CASE
In order to explain the quantities that will be used
to describe the effect of the combustion
parameters a description of a reference system,
with air preheat, is presented first. The results for
the reference case are presented in table 3.2.
Natural gas is chosen as the fuel for the reference
case. With an air factor of 1.05 and a temperature
difference of 40 K between flue gas and air at the
high temperature side of the air preheater an air
300
300
200
200
100
100
Stack
15
0
15.1
25.3
125
Transmitted heat [MW]
Figure 3.9 Value diagram of the reference case
water vapour is condensed at ambient temperature,
this quantity is somewhat less than the higher
heating value of the fuel. The shaded area below the
right part of the temperature curve, indicated as Heat
Exchgr. 5, represents the exergy of the heat
transferred from the flue gas in heat exchanger nr.5.
The remaining shaded areas represent the exergy
losses in preheater and stack. The sudden change in
slope of the temperature curve in the "Stack” is
caused by condensing water vapour. Cooling the
flue gas from stack temperature down to ambient
temperature occurs after the gases have left the stack
of course, but the losses are generally indicated as
“stack loss”. The figure illustrates that the water
content of the flue gas seriously affects these stack
losses.
units natural gas
air factor - 1.05
temp. combustion air ˚C 279
adiabatic comb. temp. ˚C 2106
system exergy efficiency % 72.80
exergy losses
combustor % 23.14
air preheater % 1.01
stack % 3.34
Table 3.2 Combustion of natural gas;
reference case
temperature of 279 ˚C is obtained, resulting in an
adiabatic combustion temperature of 2106 ˚C.
The system performance, indicated with the
system exergy efficiency, is defined as the exergy
of the heat transferred from the flue gas in heat
exchanger 5 (see Figure 3.7) divided by the
exergy of the supplied fuel. Exergy losses are
presented for the combustor, air preheater and
EFFECT OF THE TYPE OF FUEL
Table 3.3 shows the results for the combustion of
different fuels assuming an air factor of 1.10 and
combustion without air preheat. Thus air is supplied
to the combustor at ambient temperature (=15 ˚C).
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