Thermodynamic analysis of combined diesel engine and absorption unit—turbocharged engine

Applied Thermal Engineering. Vol. 16, No. 10, pp. 845--850, 1996

Pergamon

1359-4311(95)00065-8

Copyright © 1996 Elsevier Science Ltd Printed in Great Britain. All rights reserved 0890-4332/96 $15.00 + 0.00

TECHNICAL NOTE T H E R M O D Y N A M I C A N A L Y S I S O F C O M B I N E D DIESEL ENGINE AND ABSORPTION UNIT--TURBOCHARGED ENGINE M. Mostafavi*t and B. Agnew:~ i'Shiraz University, Shiraz, Iran; and ~Department of Mechanical, Materials and Manufacturing Engineering, University of Newcastle upon Tyne, Newcastle upon Tyn¢ NEI 7RU, UK (Received 9 October 1995)

Abstract--This paper describes the effects of pressure ratio and temperature ratio of the turbocharged diesel engine on the amount of net work and efficiency of the engine, and also the amount of energy available in the exhaust gases, which can be used to drive an absorption refrigeration machine. For an ideal cycle, the combined cycle is examined based on thermodynamic laws and the acceptance envelope identified. Then, based on the correlations derived, graphs have been plotted for the net work, efficiency and the amount of cooling load available in exhaust gases as a function of pressure ratio and temperature ratio of the cycle. Copyright © 1966 Elsevier Science Ltd. Keywords--Diesel absorption; CHP; energy recovery; thermodynamic analysis.

NOTATION

-

CpEX

c.o.p. cp k

mEX n~ nc P

Q~ Qcoollm

Qin.Nb

Qo Sol5

Sp35 Tin,ix T~t.Ex

T,

To T

nc-1 nc nt-I nt

--

1 tlc

X

=tlE×

k-i - k k-I ---k--

specific heat of exhaust gases coefficient of performance of the absorption compressor pressure ratio cp/cv = 1.4

mass flow rate of exhaust gases polytropic exponent for expansion polytropic exponent for compression pressure ratio of the cycle, P2/PI 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 temperature ratio of the cycle, T3/TO

Greek letters tl, compressor and compression efficiency = 0.89 expansion and turbine efficiency = 0.91 tiE ~lra~h

~HEI qH~

mechanical efficiency = 0.90 efficiency of the first heat exchanger efficiency of the second heat exchanger

*Author to whom correspondence should be addressed: Department of Mechanical, Materials and Manufacturing Engineering, University of Newcastle upon Tyne, Newcastle upon Tyne, NEI 7RU, UK

845

846

Technical Note effectiveness of the first heat exchanger effectiveness of the second heat exchanger

EHEI EHE2

Subscripts 1 2 3 4

4" 0

i.tur~ eAurbo

beginning of compression end of compression beginning of expansion end of expansion and beginning of the exhaust end of the exhaust ambient condition entrance of turbine of the turbocharger exit of turbine of the turbocharger INTRODUCTION

Supercharging can be defined as the introduction of air into an engine cylinder at a density greater than ambient. This allows a proportional increase in the fuel that can be burned and therefore raises the potential power output. The principal objective is to increase the power output, not to improve efficiency. Turbocharging is a specific method of supercharging in which no part of the engine power is absorbed by the compressor and specific fuel consumption is lower than it would be for a naturally aspirated engine. Turbocharging increases power by increasing the work done per engine cycle. Therefore the brake mean effective pressure and the mechanical efficiency increase [l I. Since the pressure of the exhaust gases leaving a turbine and, consequently, their temperatures are higher than the ambient pressure and temperature, there is still some waste energy that can be put to use. Exhaust gases, after expanding in the turbine, can be introduced to a generator of an absorption refrigeration machine to produce a chilling effect for external cooling purposes. The temperature of exhaust gases entering the absorption machine is dependent on the expansion ratio of the turbine. In this paper, based on thermodynamic laws, the amount of energy required to drive a compressor used for supercharging by an exhaust turbine has been calculated for compressor ratios of 1.5, 2, 2.5, 3 and 3.5. The net work and efficiency of a turbocharged engine and also the amount of cooling capacity available in the exhaust gases for cycle pressure ratios of 8-104 and cycle temperature ratios of 5-15 are calculated. THERMODYNAMIC

MODEL FOR COMBINED SUPERCHARGED AND ABSORPTION REFRIGERATION UNIT

DIESEL

ENGINE

The cycle examined in this study is a turbocharged diesel engine 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 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 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 air-conditioning purposes, is determined. The schematic diagram of the cycle is shown by Fig. 1 and the processes are given in and diagrams, Fig. 2. Based on the thermodynamic laws, the specific net work and the efficiency will be

P-V

T-S

W,~, 7~ mRTo -( ~-~ )~lE{T- [(cp)o],k ,,[~l¢p~+ (l _~lO]k}--(l~-a )(cp)"(P°--l)+ (l~--a )(rl--~h~h)[(cp)°-- l]

( k-l'~

-t,--~--)

(cp)"(P"-l) + f k -~-zi~7~

t,

Combustion ~ process

1~ l k

×

),~o~h

Expansion process

(,-o) ~-

[(cp) ~ - l]

T-(~pxe)o

W

42 Compression process

[

Exhaust process

I On I

l Absorption refrigeration

I Qcool Fig. 1

To

Technical Note HEAT

BALANCE

FOR

COMBINED

847 CYCLE

Referring to Fig. 1, a control volume can be established for each o f the sub-systems: Diesel engine

For the diesel engine a first-law analysis yields Q,. = Qcoolant+ W + QEx + QE QEX = mEXCpEXA T ~ = mEXCp~X(Ti..EX-- Tou,.Ex). H e a t exchanger number 1 ( / / E l )

Considering heat exchanger 1 the effectiveness and efficiency are defined as follows: C E x ( T I n , E X - To~,.Ex)

EHEI =

C~. (T,..EX -- T,) and -- Qin.ab

r/:-tE,- Q~x" So Qin.ab = ~/rtEI × QEX = r/HEI × mEXCpExATEx = r/.E* X mEX X CpEx(Ti..~x-- To~t.ex).

Absorption

For the absorber Q~..ab = Qou,.ab+ Q¢oo, and c.o.p.

=

Qcool Qin,ab

therefore Qcoo, = (c.o.p.)(Q~...b) = (C.O.p.)(rl.E,)(mEX)(C~Ex)(T,..~X -- To.,.Ex). S e c o n d heat e.xchanger (HE2)

In the case o f the second heat exchanger, an energy balance yields the following: ~/uE2Q~, + Qw.i. = Q ...... F r o m 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: C ~ ( T4 - 7"3) EHE2 =

Cmin(Tin.w- T3)

~cool

m~%~(Ti,,. -- Tout,,,.)= (rI.E,)(~I.E=)(C.O.p.)(mEX)(CpEx)(Ti..EX -- To~t.EX).

2

3

p--c

2'

4

4

p=¢

0 0

V

S Fig. 2

ATE 16/10----F

Technical Note

848

E f f i c i e n c y of t u r b o c h a r g e d , C.P.R. = 1.5, T5 to15 100 F'-

/•

g 86: ~

~

:

m/m•

•

8

16

24

• T=5 +T=7 OT=9 D T = 11

I

C

:

O T--13

~

A T =

32

40

48

56

64 72

80

88

15

96 104

P r e s s u r e ratio Fig. 3 N e t w o r k of t u r b o c h a r g e d , C.P.R. = 1.5, T5 to 0 15

20 15 ~

10 [--

I

. ~ ~

5 [-8

O T =5

.------n-'-u--n - n - n - n ~i''-T-~+ ~ + ~ + ~ + - + ~ + /~_..... x~ x^ _ _ x - - x^ ~ 4 - - ' ~ _ . . -+--- "x ~~ x ~ ^x - - x

! ~

16

24

32

'~ "0"~"~

'~ 0

,~ A 0 O"' O O

56

72

80

A

40

48

64

~

~

88

96

OT = 7 A T =9 x T=ll + T = 13

104

P r e s s u r e ratio Fig. 4

It is possible to express the exhaust gas temperature in terms of the cycle pressure ratio, cycle temperature ratio and the c o m p r e s s i o n pressure ratio, so the cooling capacity can be expressed as T4

=

To T(1 - r/E) + r/E x [(cp)~k_ "][1 + rk(P" -- 1)]k "

Let it be a s s u m e d that

U =

T ( I - n~) + r/~ x [ ( c p y " -

b-I

b-I

"1[1 + r/~(P" -

1

--

b-I

P [

To'U+To'U( T ) T ~ --

I

P

b , 7"4+ T,Ti.tu,bo = - 2

1)l*

b-i

]

To P'U+(U)(~-)'T ~

P

2

2P

then TimEX ~

Te,turbo

and We.turbo ~--- r/raech X Wt,turbo :=~" Wt,turbo We.turbo r/meeh

Also

mR( L ~ )(Tu.,bo-

T¢.,~,bo)=mRTo(l~a-a)(~-~h)[(cpy-1],

E f f i c i e n c y of t u r b o c h a r g e d , C.P.R. = 3.5, T5 to 15 100r-• I . , 5 "

.~

60b/•

~

20 ~ ~ . . ~ - ' r 8 " 16

__ - -

..~'=

=

=

i 24 32 40

O T=5 OT.7 AT=9 + T = 13

I)o L ). _L L oL J. 2

48 56 64 72 P r e s s u r e ratio Fig. 5

80 88

96 104

" T ° 15

Technical Note 20

849

Network o f turbocharged, C.P,R. = 3.5, T5 to 15

+T=7 OT=9 r'l T = 11 OT=13

5 0

~

8

16

24

32

40

48

56

64

72

~

80

88

96

104

P r e s s u r e ratio Fig. 6 which results in

b

I

I

+

l

l

Using the above equation and assuming 3% transmission losses [2] and following refs [3, 4, 5, 6], the coefficient o f the performance of the absorption unit has been set to 0.5, the cooling capacity is then is equal to Q,oo,

= (0.97)(c.o.p.)(tlH~O

rnExCp.ExTo b-I

I

DISCUSSION

OF

THE

RESULTS

The results o f the above analysis are shown in Figures 3--8. The efficiency and net specific work have been plotted for a wide range o f cycle pressure ratios from 8 to 104, whilst the compressor pressure ratio has been varied from i.5 to 3.5 in increments o f 0.5. In some cases the values o f the efficiency are below z e r o or greater than 100% and the net specific work is negative, which clearly is not acceptable. F o r brevity only the results relating to compressor ratios o f 1.5 and 3,5 are shown. Referring to the graphs o f cycle efficiency it can be seen that both compressor pressure ratios impose a m i n i m u m limit on the cycle temperature ratio for acceptable mode o f operation of approximately 7 in each case. The net specific work produced by the cycles is seen to reach a m a x i m u m value that is a function of the cycle pressure ratio and temperature ratio and then decrease with increased cycle pressure ratio as the temperature ratio remains constant; this is consistent with C o o l i n g capacity for air c o n d i t i o n i n g , . . " " = .

3.00 2.00

[] T--- 7 T=9 • T = 11

1.00

[] T = 15

0

o ~ 0

0

8

24

40

56

72

88

10

4

P r e s s u r e ratio Fig. 7 2.50 C,

2.00

g,

rO

C o o l i n g capacity for air c o n d i t i o n i n g , TC35, H.E. efficiency = 0.80

I

1.50

laT=7 raT=9 IT=ll rn T = 1 3 rn T - - 1 5

1.00 0.50

16

24

32

40

48

56

64

Pressure ratio

Fig. 8

72

80

88

96

104

850

Technical Note

well established results from previous work. The refrigeration effect that would be available to produce an external cooling effect for air-conditioning or other purposes is shown for the two cases considered in Figs 7 and 8. It can be seen that for all cycle configurations a net cooling effect can be produced but increasing the cycle pressure ratio produces diminishing returns with increasing compressor pressure ratio. The indications are that the choice of cycle parameters will be in the range of 24-40 for the cycle pressure ratio and 9 and above for the cycle temperature ratio.

CONCLUSION The following conclusions can be drawn from this work: For the range of cycle parameters examined it is possible to achieve an external cooling effect by interfacing an absorption refrigeration unit with a turbocharged diesel engine; The range of the diesel engine cycle operating parameters that would lead to satisfactory operation of the combined cycle examined would most probably be 24-40 for the cycle pressure ratio as this corresponds with the region of maximum work output and 9 and above for the cycle temperature ratio.

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. P. E. Hufford, Absorption chillers maximise cogeneration value. ASHRAE Trans. 10, 428-433 (1992). 3. N. B. Metha, Analysis of combined gas turbine and absorption-refrigeration cycles, M. S. Thesis, Illinois Institute of Technology (1970). 4. Xu Guany Qi, Cogeneration System of Utilising Residual Heatfrom IC Engine IEEE. IEEE Service Centre, pp. 135-138 (1985). 5. I. G. C. Dryden, The Efficient Use of Energy, 2nd Edition. Butterworth, London (1982). 6. American Society of Heating, Refrigeration and Air-conditioning Engineers, Fundamentals Handbook (SI). ASHRAE, Washington (1989).