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Thermodynamic analysis of combined diesel engine and absorption unit-turbocharged engine with intercooling
Applied Thermal Engineering Vol. 16, Nos g/9, pp. 733-740, 1996 Copyright 0 19% ElsevierScience Ltd
Pergamon
1359-4311(95)ooo64-x
Printed in Great Britain. All rights reserved 1359-431l/96 $15.00+ 0.00
THERMODYNAMIC ANALYSIS OF COMBINED DIESEL ENGINE AND ABSORPTION UNIT-TURBOCHARGED ENGINE WITH INTERCOOLING M. Mostafavi*t
and B. Agnew$
TShiraz University, Shiraz, Iran; and JDepartment of Mechanical, Materials and Manufacturing Engineering, University of Newcastle Upon Tyne, Newcastle Upon Tyne NE1 7RU, UK (Received 9 October 1995)
Abstract-Large diesel engines are well suited for power generation in remote areas, their high thermal efficiency and relative insensitivity to high ambient temperatures make them suitable for use in desert and tropical climates. The exhaust gases from such installations represent a significant amount of thermal energy that can be put to beneficial use. This paper explores the utilisation of the exhaust gases from a turbocharged and intercooled engine as an energy input to a vapour absorption refrigeration machine that is used primarily to cool the engine charge air in an intercooler. The influence of the engine configuration and performance parameters on the performance of an ideal system is investigated. Copyright 0 1996 Elsevier Science Ltd Keywords-Diesel
absorption; CHP; energy recovery; thermodynamic analysis.
NOTATION a
n,-I=lx_
b
n, - I -
n,
k-l k
% =t/fEx
k-l k
CIEX c.0.p. CP
n, specific heat of exhaust gases coefficient of performance of the absorption compressor pressure ratio
k
4 =1.4
rnEX
mass flow rate of exhaust gases
P
P2 pressure ratio of the cycle, f’~
C-8
SplS sp35 T,“.EX TO”,.EX T, TO
k?”
input heat for diesel engine heat of exhaust gases heat transfer to environment heat transfer to cooling system input heat for absorption cooling capacity of absorption cooling capacity of chilled water compressor pressure ratio of 1.5 compressor pressure ratio of 3.5 exhaust gases inlet temperature exhaust gases outlet temperature maximum temperature of the cycle ambient temperature
T
temperature
Ql. QEX QE ;“:Y
unit
T
ratio of the cycle, 2
polytropic exponent for expansion polytropic exponent for CompressionGreek letters compressor and compression efficiency = 0.89 expansion and turbine efficiency = 0.91 mechanical efficiency = 0.90 efficiency of the first heat exchanger efficiency of the second heat exchanger *Author to whom correspondence should be addressed. Present address: Department of Mechanical, Materials and Manufacturing Engineering, University of Newcastle Upon Tyne, Newcastle Upon Tyne NE1 7RU, UK. 733
M. Mostafavi
134 SHE!
&HE2
effectiveness effectiveness
and B. Agnew
of the first heat exchanger of the second heat exchanger
Subscripts
1 2 3 4 4” i,turbo e,turbo 0 0’
beginning of compression end of compression beginning of expansion end of expansion and beginning of the exhaust end of the exhaust entrance of turbine of the turbocharger exit of turbine of the turbocharger ambient condition exit of compressor
INTRODUCTION
The amount of charge in an engine cylinder and, hence, the power output are related to the charge density before the intake valve. Forced induction is therefore a recognised means of improving the performance of reciprocating engines but, in order to avoid excessive thermal loading of the engine, it is often necessary to precool or intercool the charge air. Compression of air in a compressor is accompanied by a temperature rise which depends on the pressure ratio, on the efficiency of the compressor, and on the heat losses in the compressor. If water at ambient temperature is available as a coolant, charge air cooling becomes feasible at boost ratios as low as 1.5. If the boost ratio is above 2, the charge air should always be cooled to reduce the thermal loading and thus to improve the operational reliability of the engine [l]. This also has the effect of reducing the emissions of CO, C,H! and NO,. Turbocharging reduces CO emissions by almost one-half, practically independent of output, NO, emissions increase slightly without charge air cooling, but fall by one-quarter with charge air cooling. C,H, emissions of turbocharged engines with charge air cooling are only half of those of naturally aspirated engines [l]. The beneficial aspects of charge air cooling can be considerable. In this paper, the concept of a diesel engine combined with an absorption refrigeration unit that can be used for cooling the charge air prior to ingestion to the engine cylinder or for other cooling purposes, such as air-conditioning, is examined. As in this proposed arrangement, the charge air is air cooled, effectively making the engine a self-contained unit by negating the necessity to provide external cooling water supplies. The work is based on the thermodynamic model of an ideal diesel cycle and an ideal absorption cycle that is used to estimate the net work and efficiency of the engine. The amount of cooling capacity required for intercooling and the amount of cooling capacity available in the exhaust gases have been calculated for a wide range of pressure ratios and temperature ratios of the diesel cycle.
THERMODYNAMIC
MODEL
FOR
COMBINED
CYCLE
The cycle examined in this study is turbocharged diesel engine with intercooling combined with an absorption refrigeration unit. The object of the study is the examination of the thermodynamic performance of each of the components of this cycle and to determine the most suitable configuration for combining it with an absorption unit. The specific work and efficiency of the cycle as a function of temperature ratio and pressure ratio of the cycle based on thermodynamic laws are calculated, then by writing a heat balance equation for the cycle the amount of heat available in the exhaust gases, which can be used for intercooling or air-conditioning purposes or both, is determined. The efficiency and net work have been plotted for a wide range of pressure ratios from 8 to 104. While the compressor pressure ratio has been varied from 1.5, 2, 2.5, 3 and 3.5. In some cases studied the values of efficiency are below zero or above 100, and the net work is negative, which, from the thermodynamic viewpoint, does not have any meaning. For brevity only the efficiency, net work and cooling capacity at compressor pressure ratios of 1.5 and 3.5 are shown, Figs 1-8, but the complete results are available in ref. [2]. A schematic diagram of the cycle is shown by Fig. 1 and the processes are given in P-V and T-S diagrams, Fig. 2. Based on the first law of thermodynamics, the specific net work and the efficiency of the cycle can be expressed as [2]
Turbocharged
engine with intercooling
3
Combustion process
735
W
Expansion process
2 Compression process
Exhaust
Qin
process
---
TO
Qin,ab
1
Absorption refrigeration
Qcool
Comp
To YPO Fig. 1. Schematic
diagram
of turbocharged
with intercooling
diesel engine.
(y--) l-a
I
-
k-l k H
1 qmech
T-
[(q)“-$$P)
I(v)” - 11
T-[(cpY’-%I(P) 3
2
3
-_--_--_--
p=c ,2’
z--K
45
1 1
0’
p=c
4’ ’
5
4
p=c -_--_--
--_-
0”
1
P
-------_--_--
T
0’
v=c
p=e
_&j$
0 0
V Fig. 2. P-V
and T-S diagrams
of turbocharged
s with intercooling
diesel engine.
M. Mostafavi and B. Agnew
736
Efficiency of turbocharged with intercooling, C.P.R. = 1.5, TO-T1 = 5, T5 to 15
.
T=5
Pressure ratio Fig. 3. Efficiency of turbocharged engine with intercooling-compression
pressure ratio of 1.5.
x(P.-l)T-[(cpYq(P)‘+ J.-(y) l-a a
x
[(cp)” - l]T-
(
[(W
- +-JY.
>
HEAT
BALANCE
FOR
COMBINED
CYCLE
Referring to Fig. 1, the diesel engine and absorption refrigeration unit are interfaced via heat exchangers that act as the charge air intercooler and deliver energy to the absorption unit. The following analysis is performed on the sub-systems that are shown in Fig. 1. The diesel engine
For the diesel engine a first law analysis yields Qin= Qcoo~ant + W+QEX+QE,
where QEX=
rn~cpEXAT~X
=
maxCpax(&x
-
TWX).
Heat exchanger number 1 (HEl)
Considering heat exchanger 1, the effectiveness and efficiency are defined as follows: CEX(T,,EX- Tout.~x) &HE1
=
Cmi,(Tin,~x- T,)
and ein.ab VHEI
2o
=
QEX ’
Network of turbocharged with intercooling, C.P.R. = 1.5, TO-T1 = 5, T5 to 15
r
0 T=5 OT=7 AT=9 x T=ll +T=13 . T=15
1s 10 5 0
8
16
24
32
40
48
56
64
72
80
88
96
104
Pressure ratio Fig. 4. Network of turbocharged engine with intercooling-compression
pressure ratio of 1.5.
Turbocharged engine with intercooling
737
Efficiency of turbocharged with intercooling. C.P.R. = 3.5, TO-T1 = 5, T5 to 15
.T=l3 8
16
24
32
40
48
56
64
72
80
88
96
104
Pressure ratio Fig. 5. Efficiency of turbocharged engine with intercooling-compression
pressure ratio of 3.5.
which leads to the following expression for the heat input to the absorption Qln.ab=
~~HEI X
QEX =
~HEI
X
mExCpExAT
EX =
~HEI
X
rnEx
X
C~EX(%,EX
unit: -
Tout.~~).
Absorption
The energy flows in the absorption
unit are linked as follows: (&nab
=
&t.ab
Qm,,
+
and using the definition of coefficient of performance, c.0.p. --- ;;I;
i.e. )
therefore the cooling capacity can then be expressed as Qcool =
Second heat exchanger
(C.O.P.)(Qin.ab)
=
(C.O.p.)(?HEl)(mEX)(c,EX)(T,".EX
TOu&.
-
(HE2)
In the case of the second heat exchanger an energy balance yields the following:
From the above and the definitions of heat exchanger effectiveness and efficiency the relationship between the mass and temperature difference of the chilled media and other system parameters can be established: CAT, - T,) EHR= C,,(Ti,,,, - T3) qHE2
=
mwc,4 Tn.w - LJ
*Qc =
&
=
qHE2
x
QCoo,= mwcpw(ATw)
(~HEl)(qHEZ)(C.O.P.)(mEX)(CpEX)(~n,EX
-
kEX).
It is possible to express the exhaust gas temperature in terms of the cycle pressure ratio, cycle temperature ratio and the compression pressure ratio, so, the cooling capacity can be expressed as
T4 = TO T(1 - VE) +
I]E X
[l + ?c(P -
l)lk
M. Mostafavi
738
and B. Agnew
Network of turbocharged with intercooling, C.P.R. = 3.5, TO-T1 = 5, T5 to 15 20 15 -
_.-.-m-m
T=5 0 T=l
??
x T=13 ?? T=15 8
16
24
32
40
48
64
56
Pressure Fig. 6. Network
of turbocharged
engine
72
80
88
96
ratio
with intercooling-compression
pressure
ratio
of 3.5.
Let it be assumed that
Q = To T(1 - $‘E)+
VE
x
[l + ?C(p”- 111”’ then T4..
The temperature
=
(TO)@ - l/b) .
Q(b- lib).T;‘b
Tllb . p
. Q’b
T,,
-
l/b)
. l-l/b
=
P
’
of the exhaust at the turbine inlet is To
. Q
+
To.Q’b
-
=To[P . Q + (Q)‘b - lib). T Ilb]
lib) . y
2P
2 and the temperature of the exhaust gases introduced temperature of the gases leaving the turbine, i.e. Tln.EX -
The work balance on the turbine-compressor W c,turbo
=
qmech x
to the first heat exchanger is equal to the
Tesurba.
unit produces the following:
W t,turbo=’
W c turbo Wwrbo
L
, &nech
which leads to the following expression relating the inlet temperature component efficiencies and the engine exhaust temperature: mR
2.5 r-
z
2.0
u”
1.5
E
d 2 i;
(
9
)
(xi.turbo
-
Te.turbo)
=
delivery in terms of the
mRTo
Cooling capacity for air conditioning, TC15P15, H.E. efficiency = 0.80
_
? ? T=l
111T=9 T=ll 0 T=13 0 T=15
??
1.0 0.5 n ”
8
16
24
32
40
48
56
Pressure Fig. 7. Cooling
capacity
available
in the exhaust
64
72
80
88
96
104
ratio
gases for air-conditioning of 1.5
for compression
pressure
ratio
Turbocharged
739
engine with intercooling
Cooling capacity for air conditioning, TC15P35, H.E. efficiency = 0.80
g2
2.0
? ? T=l ? ?T=9
1.5
u”
1.0
E 2 8
I
T=ll
0
T=13
I3 T=15
0.5 n 8
”
16
24
32
40
48
56
Pressure Fig. 8. Cooling
capacity
available
in the exhaust
64
72
80
88
96
104
ratio
gases for air-conditioning of 3.5.
for compression
pressure
ratio
Therefore
Temrbo
_ T -
Q.p
+
Q’b-
lib,.TW
0
Assuming a 3% transmission losses (duct work) [3] the cooling capacity available from the combined unit can be derived as follows:
Q cool
= (0.97)(C.O.P.)(~“E,)
m&.ExTo
1
(b WLTllb Q.p + p26
-(+$(+&)(&)KCPY-lly}. COOLING
CAPACITY
AVAILABL’E
FOR
AIR-CONDITIONING
For a turbocharged engine with intercooling, the amount of cooling capacity for air-conditioning is equal to the difference between the amount of cooling capacity available from the combined cycle, and the amount of cooling used for intercooling. Calculation
of the amount
of cooling capacity
necessary for intercooling
Calculations are based on the assumptions that the specific heat throughout the cycle is constant and the mass of air is equal to the mass of exhaust gases, so
~ Qc2 = E maCpaTO
Kcp)” -
11.
From the above calculations, the amount of cooling capacity necessary for intercooling can be calculated, so it is then straightforward to determine the amount of cooling available for air-conditioning, which is expressed below: Qd mEXCp,EX
= (~“E2)(o.97)(c*o.p.)(~“E,)
p . Q + $;-
lib) ’ T”b
TO
x [(cp>”-
-(G+%)($J
T
l] - 7
- z
[@PI - 11.
The depicted results have been based on a value of c.o.p. of 0.50 for the absorption suggested by [4, 5, 6, 71.
unit as
M. Mostafavi and B. Agnew
740
DISCUSSION
OF THE
RESULTS
The results of the above analysis are shown in the accompanying figures for different levels of charge air cooling and compression ratio. It can be seen that for the case of a turbocharger compression ratio of 1.5 or less it is possible to obtain both intercooling and external cooling for moderate degrees of intercooling but above this a choice must be made between intercooling and external cooling. The traditional means of intercooling are by air or liquid cooling using the engine coolant delivered from the engine radiator; methods that in regions of high ambient temperatures (45°C) are not ideal. The results indicate that intercooling will improve the efficiency and net work output of the diesel cycle and it is evident that the capacity exists in using an absorption cooler to cool the charge air to much lower temperatures than would be possible with the traditional methods. It is also possible to lower the point on the turbocharger pressure ratio at which intercooling becomes effective. For instance, with an ambient temperature of 45°C and a compression ratio of 1.5 and an efficiency of 70% the compressor delivery air temperature would be 374 K, which is below the normal engine coolant temperature and therefore could not be considered for intercooling. Using the engine coolant with the absorption system it is possible to cool this air by 60°C to 314 K, which is below the ambient temperature. The question then arises of the possibility of pre-cooling in the ambient conditions considered. The output of a turbocharged engine would be significantly larger than a similar naturally aspirated engine in the case described above. As the turbocharger compression ratio is increased the charge air temperature will also increase but the high degree of cooling available will reduce the overall engine thermal loading for a given work output, making the engine even less sensitive to the ambient conditions. CONCLUSION
The efficiency, net work and cooling load available for air-conditioning have been plotted for pressure ratios from 8 to 104, temperature ratios of 5-15 and compressor pressure ratios of 1.5-3.5, from which one can conclude that: (i) the feasibility of combining a turbocharged and intercooled diesel engine with an exhaust gas driven absorption refrigeration unit has been demonstrated; (ii) the influence of the cycle pressure ratio is such that the cooling capacity decreases with increased pressure ratio, irrespective of the cycle temperature ratio until the point is reached where there is insufficient cooling capacity to provide intercooling. This is worse for cycles with low temperature ratios and large degrees of intercooling. Acknowledgements-The first author would like to acknowledge, with thanks, financial support from the Ministry of Culture and Higher Education of the Islamic Republic of Iran, without which this study would not have been possible.
REFERENCES 1. K. Zinner, Supercharging of Internal Combustion Engines. Springer, Berlin (1978). 2. M. Mostafavi, Theoretical and experimental investigation of using power plant waste heat for absorption refrigeration applications, PhD Thesis (in preparation). 3. P. E. Hufford, Absorption chillers maximise cogeneration value. ASHRAE Trans. 10, 428433 (1992). 4. N. B. Metha, Analysis of a combined gas turbine and absorption-refrigeration cycles, M. S. Thesis, Illinois Institute
of Technology (1970). 5. Xu Guany Qi, Cogeneration
System of Utiiising Residual Hearfrom IC Engine IEEE, pp. 135-138. IEEE Service Centre (1985). 6. I. G. C. Dryden, The Eficient Use of Energy, 2nd Edition. Butterworths, London (1982). 7. American Society of Heating, Refrigeration and Air-conditioning Engineers, Fundamentals Handbook (SI). ASHRAE, Washington (1989).
Pergamon
1359-4311(95)ooo64-x
Printed in Great Britain. All rights reserved 1359-431l/96 $15.00+ 0.00
THERMODYNAMIC ANALYSIS OF COMBINED DIESEL ENGINE AND ABSORPTION UNIT-TURBOCHARGED ENGINE WITH INTERCOOLING M. Mostafavi*t
and B. Agnew$
TShiraz University, Shiraz, Iran; and JDepartment of Mechanical, Materials and Manufacturing Engineering, University of Newcastle Upon Tyne, Newcastle Upon Tyne NE1 7RU, UK (Received 9 October 1995)
Abstract-Large diesel engines are well suited for power generation in remote areas, their high thermal efficiency and relative insensitivity to high ambient temperatures make them suitable for use in desert and tropical climates. The exhaust gases from such installations represent a significant amount of thermal energy that can be put to beneficial use. This paper explores the utilisation of the exhaust gases from a turbocharged and intercooled engine as an energy input to a vapour absorption refrigeration machine that is used primarily to cool the engine charge air in an intercooler. The influence of the engine configuration and performance parameters on the performance of an ideal system is investigated. Copyright 0 1996 Elsevier Science Ltd Keywords-Diesel
absorption; CHP; energy recovery; thermodynamic analysis.
NOTATION a
n,-I=lx_
b
n, - I -
n,
k-l k
% =t/fEx
k-l k
CIEX c.0.p. CP
n, specific heat of exhaust gases coefficient of performance of the absorption compressor pressure ratio
k
4 =1.4
rnEX
mass flow rate of exhaust gases
P
P2 pressure ratio of the cycle, f’~
C-8
SplS sp35 T,“.EX TO”,.EX T, TO
k?”
input heat for diesel engine heat of exhaust gases heat transfer to environment heat transfer to cooling system input heat for absorption cooling capacity of absorption cooling capacity of chilled water compressor pressure ratio of 1.5 compressor pressure ratio of 3.5 exhaust gases inlet temperature exhaust gases outlet temperature maximum temperature of the cycle ambient temperature
T
temperature
Ql. QEX QE ;“:Y
unit
T
ratio of the cycle, 2
polytropic exponent for expansion polytropic exponent for CompressionGreek letters compressor and compression efficiency = 0.89 expansion and turbine efficiency = 0.91 mechanical efficiency = 0.90 efficiency of the first heat exchanger efficiency of the second heat exchanger *Author to whom correspondence should be addressed. Present address: Department of Mechanical, Materials and Manufacturing Engineering, University of Newcastle Upon Tyne, Newcastle Upon Tyne NE1 7RU, UK. 733
M. Mostafavi
134 SHE!
&HE2
effectiveness effectiveness
and B. Agnew
of the first heat exchanger of the second heat exchanger
Subscripts
1 2 3 4 4” i,turbo e,turbo 0 0’
beginning of compression end of compression beginning of expansion end of expansion and beginning of the exhaust end of the exhaust entrance of turbine of the turbocharger exit of turbine of the turbocharger ambient condition exit of compressor
INTRODUCTION
The amount of charge in an engine cylinder and, hence, the power output are related to the charge density before the intake valve. Forced induction is therefore a recognised means of improving the performance of reciprocating engines but, in order to avoid excessive thermal loading of the engine, it is often necessary to precool or intercool the charge air. Compression of air in a compressor is accompanied by a temperature rise which depends on the pressure ratio, on the efficiency of the compressor, and on the heat losses in the compressor. If water at ambient temperature is available as a coolant, charge air cooling becomes feasible at boost ratios as low as 1.5. If the boost ratio is above 2, the charge air should always be cooled to reduce the thermal loading and thus to improve the operational reliability of the engine [l]. This also has the effect of reducing the emissions of CO, C,H! and NO,. Turbocharging reduces CO emissions by almost one-half, practically independent of output, NO, emissions increase slightly without charge air cooling, but fall by one-quarter with charge air cooling. C,H, emissions of turbocharged engines with charge air cooling are only half of those of naturally aspirated engines [l]. The beneficial aspects of charge air cooling can be considerable. In this paper, the concept of a diesel engine combined with an absorption refrigeration unit that can be used for cooling the charge air prior to ingestion to the engine cylinder or for other cooling purposes, such as air-conditioning, is examined. As in this proposed arrangement, the charge air is air cooled, effectively making the engine a self-contained unit by negating the necessity to provide external cooling water supplies. The work is based on the thermodynamic model of an ideal diesel cycle and an ideal absorption cycle that is used to estimate the net work and efficiency of the engine. The amount of cooling capacity required for intercooling and the amount of cooling capacity available in the exhaust gases have been calculated for a wide range of pressure ratios and temperature ratios of the diesel cycle.
THERMODYNAMIC
MODEL
FOR
COMBINED
CYCLE
The cycle examined in this study is turbocharged diesel engine with intercooling combined with an absorption refrigeration unit. The object of the study is the examination of the thermodynamic performance of each of the components of this cycle and to determine the most suitable configuration for combining it with an absorption unit. The specific work and efficiency of the cycle as a function of temperature ratio and pressure ratio of the cycle based on thermodynamic laws are calculated, then by writing a heat balance equation for the cycle the amount of heat available in the exhaust gases, which can be used for intercooling or air-conditioning purposes or both, is determined. The efficiency and net work have been plotted for a wide range of pressure ratios from 8 to 104. While the compressor pressure ratio has been varied from 1.5, 2, 2.5, 3 and 3.5. In some cases studied the values of efficiency are below zero or above 100, and the net work is negative, which, from the thermodynamic viewpoint, does not have any meaning. For brevity only the efficiency, net work and cooling capacity at compressor pressure ratios of 1.5 and 3.5 are shown, Figs 1-8, but the complete results are available in ref. [2]. A schematic diagram of the cycle is shown by Fig. 1 and the processes are given in P-V and T-S diagrams, Fig. 2. Based on the first law of thermodynamics, the specific net work and the efficiency of the cycle can be expressed as [2]
Turbocharged
engine with intercooling
3
Combustion process
735
W
Expansion process
2 Compression process
Exhaust
Qin
process
---
TO
Qin,ab
1
Absorption refrigeration
Qcool
Comp
To YPO Fig. 1. Schematic
diagram
of turbocharged
with intercooling
diesel engine.
(y--) l-a
I
-
k-l k H
1 qmech
T-
[(q)“-$$P)
I(v)” - 11
T-[(cpY’-%I(P) 3
2
3
-_--_--_--
p=c ,2’
z--K
45
1 1
0’
p=c
4’ ’
5
4
p=c -_--_--
--_-
0”
1
P
-------_--_--
T
0’
v=c
p=e
_&j$
0 0
V Fig. 2. P-V
and T-S diagrams
of turbocharged
s with intercooling
diesel engine.
M. Mostafavi and B. Agnew
736
Efficiency of turbocharged with intercooling, C.P.R. = 1.5, TO-T1 = 5, T5 to 15
.
T=5
Pressure ratio Fig. 3. Efficiency of turbocharged engine with intercooling-compression
pressure ratio of 1.5.
x(P.-l)T-[(cpYq(P)‘+ J.-(y) l-a a
x
[(cp)” - l]T-
(
[(W
- +-JY.
>
HEAT
BALANCE
FOR
COMBINED
CYCLE
Referring to Fig. 1, the diesel engine and absorption refrigeration unit are interfaced via heat exchangers that act as the charge air intercooler and deliver energy to the absorption unit. The following analysis is performed on the sub-systems that are shown in Fig. 1. The diesel engine
For the diesel engine a first law analysis yields Qin= Qcoo~ant + W+QEX+QE,
where QEX=
rn~cpEXAT~X
=
maxCpax(&x
-
TWX).
Heat exchanger number 1 (HEl)
Considering heat exchanger 1, the effectiveness and efficiency are defined as follows: CEX(T,,EX- Tout.~x) &HE1
=
Cmi,(Tin,~x- T,)
and ein.ab VHEI
2o
=
QEX ’
Network of turbocharged with intercooling, C.P.R. = 1.5, TO-T1 = 5, T5 to 15
r
0 T=5 OT=7 AT=9 x T=ll +T=13 . T=15
1s 10 5 0
8
16
24
32
40
48
56
64
72
80
88
96
104
Pressure ratio Fig. 4. Network of turbocharged engine with intercooling-compression
pressure ratio of 1.5.
Turbocharged engine with intercooling
737
Efficiency of turbocharged with intercooling. C.P.R. = 3.5, TO-T1 = 5, T5 to 15
.T=l3 8
16
24
32
40
48
56
64
72
80
88
96
104
Pressure ratio Fig. 5. Efficiency of turbocharged engine with intercooling-compression
pressure ratio of 3.5.
which leads to the following expression for the heat input to the absorption Qln.ab=
~~HEI X
QEX =
~HEI
X
mExCpExAT
EX =
~HEI
X
rnEx
X
C~EX(%,EX
unit: -
Tout.~~).
Absorption
The energy flows in the absorption
unit are linked as follows: (&nab
=
&t.ab
Qm,,
+
and using the definition of coefficient of performance, c.0.p. --- ;;I;
i.e. )
therefore the cooling capacity can then be expressed as Qcool =
Second heat exchanger
(C.O.P.)(Qin.ab)
=
(C.O.p.)(?HEl)(mEX)(c,EX)(T,".EX
TOu&.
-
(HE2)
In the case of the second heat exchanger an energy balance yields the following:
From the above and the definitions of heat exchanger effectiveness and efficiency the relationship between the mass and temperature difference of the chilled media and other system parameters can be established: CAT, - T,) EHR= C,,(Ti,,,, - T3) qHE2
=
mwc,4 Tn.w - LJ
*Qc =
&
=
qHE2
x
QCoo,= mwcpw(ATw)
(~HEl)(qHEZ)(C.O.P.)(mEX)(CpEX)(~n,EX
-
kEX).
It is possible to express the exhaust gas temperature in terms of the cycle pressure ratio, cycle temperature ratio and the compression pressure ratio, so, the cooling capacity can be expressed as
T4 = TO T(1 - VE) +
I]E X
[l + ?c(P -
l)lk
M. Mostafavi
738
and B. Agnew
Network of turbocharged with intercooling, C.P.R. = 3.5, TO-T1 = 5, T5 to 15 20 15 -
_.-.-m-m
T=5 0 T=l
??
x T=13 ?? T=15 8
16
24
32
40
48
64
56
Pressure Fig. 6. Network
of turbocharged
engine
72
80
88
96
ratio
with intercooling-compression
pressure
ratio
of 3.5.
Let it be assumed that
Q = To T(1 - $‘E)+
VE
x
[l + ?C(p”- 111”’ then T4..
The temperature
=
(TO)@ - l/b) .
Q(b- lib).T;‘b
Tllb . p
. Q’b
T,,
-
l/b)
. l-l/b
=
P
’
of the exhaust at the turbine inlet is To
. Q
+
To.Q’b
-
=To[P . Q + (Q)‘b - lib). T Ilb]
lib) . y
2P
2 and the temperature of the exhaust gases introduced temperature of the gases leaving the turbine, i.e. Tln.EX -
The work balance on the turbine-compressor W c,turbo
=
qmech x
to the first heat exchanger is equal to the
Tesurba.
unit produces the following:
W t,turbo=’
W c turbo Wwrbo
L
, &nech
which leads to the following expression relating the inlet temperature component efficiencies and the engine exhaust temperature: mR
2.5 r-
z
2.0
u”
1.5
E
d 2 i;
(
9
)
(xi.turbo
-
Te.turbo)
=
delivery in terms of the
mRTo
Cooling capacity for air conditioning, TC15P15, H.E. efficiency = 0.80
_
? ? T=l
111T=9 T=ll 0 T=13 0 T=15
??
1.0 0.5 n ”
8
16
24
32
40
48
56
Pressure Fig. 7. Cooling
capacity
available
in the exhaust
64
72
80
88
96
104
ratio
gases for air-conditioning of 1.5
for compression
pressure
ratio
Turbocharged
739
engine with intercooling
Cooling capacity for air conditioning, TC15P35, H.E. efficiency = 0.80
g2
2.0
? ? T=l ? ?T=9
1.5
u”
1.0
E 2 8
I
T=ll
0
T=13
I3 T=15
0.5 n 8
”
16
24
32
40
48
56
Pressure Fig. 8. Cooling
capacity
available
in the exhaust
64
72
80
88
96
104
ratio
gases for air-conditioning of 3.5.
for compression
pressure
ratio
Therefore
Temrbo
_ T -
Q.p
+
Q’b-
lib,.TW
0
Assuming a 3% transmission losses (duct work) [3] the cooling capacity available from the combined unit can be derived as follows:
Q cool
= (0.97)(C.O.P.)(~“E,)
m&.ExTo
1
(b WLTllb Q.p + p26
-(+$(+&)(&)KCPY-lly}. COOLING
CAPACITY
AVAILABL’E
FOR
AIR-CONDITIONING
For a turbocharged engine with intercooling, the amount of cooling capacity for air-conditioning is equal to the difference between the amount of cooling capacity available from the combined cycle, and the amount of cooling used for intercooling. Calculation
of the amount
of cooling capacity
necessary for intercooling
Calculations are based on the assumptions that the specific heat throughout the cycle is constant and the mass of air is equal to the mass of exhaust gases, so
~ Qc2 = E maCpaTO
Kcp)” -
11.
From the above calculations, the amount of cooling capacity necessary for intercooling can be calculated, so it is then straightforward to determine the amount of cooling available for air-conditioning, which is expressed below: Qd mEXCp,EX
= (~“E2)(o.97)(c*o.p.)(~“E,)
p . Q + $;-
lib) ’ T”b
TO
x [(cp>”-
-(G+%)($J
T
l] - 7
- z
[@PI - 11.
The depicted results have been based on a value of c.o.p. of 0.50 for the absorption suggested by [4, 5, 6, 71.
unit as
M. Mostafavi and B. Agnew
740
DISCUSSION
OF THE
RESULTS
The results of the above analysis are shown in the accompanying figures for different levels of charge air cooling and compression ratio. It can be seen that for the case of a turbocharger compression ratio of 1.5 or less it is possible to obtain both intercooling and external cooling for moderate degrees of intercooling but above this a choice must be made between intercooling and external cooling. The traditional means of intercooling are by air or liquid cooling using the engine coolant delivered from the engine radiator; methods that in regions of high ambient temperatures (45°C) are not ideal. The results indicate that intercooling will improve the efficiency and net work output of the diesel cycle and it is evident that the capacity exists in using an absorption cooler to cool the charge air to much lower temperatures than would be possible with the traditional methods. It is also possible to lower the point on the turbocharger pressure ratio at which intercooling becomes effective. For instance, with an ambient temperature of 45°C and a compression ratio of 1.5 and an efficiency of 70% the compressor delivery air temperature would be 374 K, which is below the normal engine coolant temperature and therefore could not be considered for intercooling. Using the engine coolant with the absorption system it is possible to cool this air by 60°C to 314 K, which is below the ambient temperature. The question then arises of the possibility of pre-cooling in the ambient conditions considered. The output of a turbocharged engine would be significantly larger than a similar naturally aspirated engine in the case described above. As the turbocharger compression ratio is increased the charge air temperature will also increase but the high degree of cooling available will reduce the overall engine thermal loading for a given work output, making the engine even less sensitive to the ambient conditions. CONCLUSION
The efficiency, net work and cooling load available for air-conditioning have been plotted for pressure ratios from 8 to 104, temperature ratios of 5-15 and compressor pressure ratios of 1.5-3.5, from which one can conclude that: (i) the feasibility of combining a turbocharged and intercooled diesel engine with an exhaust gas driven absorption refrigeration unit has been demonstrated; (ii) the influence of the cycle pressure ratio is such that the cooling capacity decreases with increased pressure ratio, irrespective of the cycle temperature ratio until the point is reached where there is insufficient cooling capacity to provide intercooling. This is worse for cycles with low temperature ratios and large degrees of intercooling. Acknowledgements-The first author would like to acknowledge, with thanks, financial support from the Ministry of Culture and Higher Education of the Islamic Republic of Iran, without which this study would not have been possible.
REFERENCES 1. K. Zinner, Supercharging of Internal Combustion Engines. Springer, Berlin (1978). 2. M. Mostafavi, Theoretical and experimental investigation of using power plant waste heat for absorption refrigeration applications, PhD Thesis (in preparation). 3. P. E. Hufford, Absorption chillers maximise cogeneration value. ASHRAE Trans. 10, 428433 (1992). 4. N. B. Metha, Analysis of a combined gas turbine and absorption-refrigeration cycles, M. S. Thesis, Illinois Institute
of Technology (1970). 5. Xu Guany Qi, Cogeneration
System of Utiiising Residual Hearfrom IC Engine IEEE, pp. 135-138. IEEE Service Centre (1985). 6. I. G. C. Dryden, The Eficient Use of Energy, 2nd Edition. Butterworths, London (1982). 7. American Society of Heating, Refrigeration and Air-conditioning Engineers, Fundamentals Handbook (SI). ASHRAE, Washington (1989).