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The economy of nonazeotropic refrigerant mixtures in compression heat pumps
Heat Recovery Systems Vol. 6, No. 4, pp. 313-321, 1986 Printed in Great Britain.
0198-7593/86 $3.00+ .00 Pergamon Journals Ltd
THE ECONOMY OF NONAZEOTROPIC REFRIGERANT MIXTURES IN COMPRESSION HEAT PUMPS H. OFNER, H. SCHNITZERand F. MOSER Institute of Chemical Engineering, Technical University Graz, Inffeldgasse 25, A-8010 Graz, Austria
(Received 12 January 1986) Abstract--The relative economic advantages of heat pumps using nonazeotropic mixtures are examined.
1. I N T R O D U C T I O N In heat pumps with pure refrigerants as working fluids both the condensation and evaporation take place at constant temperature if the pressure drop is neglected. With a nonazeotropic mixture, however, the temperature will decrease during condensation and increase during evaporation. For that reason the performances of heat pumps can be improved, especially in applications with high temperature gradients of the heat sink and the heat source. This is one phenomenon that, in addition to other advantages of nonazeotropic mixtures, is described in this article. Especially in the process industries there are various possible applications of heat pumps for the recovery of waste heat [1]. Before reporting technical and economical evaluations on such machines reference is made to some basic work where at a high number of industrial plants the temperatures and the mass flowrates of waste heat streams were studied [2]. Here also the temperatures to which the heat has to be transformed to render it usable in the plant and save the maximum amount of primary energy had been worked out. From this study the most frequent boundary conditions for compression heat pumps in industrial plants are determined and machines are laid out with different pure working fluids and fluid mixtures. The economical effectiveness of the heat pumps is expressed in terms of the payback periods. In the article the main results of the author's doctoral thesis are described. For the details concerning the mathematical model, etc, see Ofner [3]. 2. T H E R M O D Y N A M I C A L B A C K G R O U N D - - P O S S I B L E IMPROVEMENTS WITH NONAZEOTROPIC MIXTURES One important thermodynamic consequence for a heat pump, when a binary nonazeotropic mixture is used instead of a pure working fluid, is shown schematically in Fig. 1. Here the
T L
4
Purefluid ~
3 Condensotion
T.
T.. T,
Evoporotion
Purefluid Heat fLux
Fig. 1. Temperature variations of working fluids during condensation and evaporation (pure fluids and non-azeotropic mixtures). 313
314
H.
O F N E R et aL
Dew-point curve BubbLe-point curve
Xl
X2
Concent rot ion
Fig. 2. The evaporation of a non-azeotroptc mtxture in the T,x-diagram.
temperature variations of the working fluid, the heat source and the heat sink are plotted in a temperature-heat diagram. In the case of a pure refrigerant the temperatures during the phase changes in the condenser and evaporator are constant if the pressure drop is neglected. For a nonazeotropic mixture the temperature wilt decrease along the condensation line and increase along the evaporation line. The reasons for these temperature variations are as follows. For a given pressure a nonazeotropic mixture will have bubble-point and dew-point curves as shown in principle in the T,x-diagram, Fig. 2. When a liquid is vapourized, the more volatile component will be enriched in the vapour phase, thus successively inere,asing the mole fraction for the less volatile component in the liquid phase. This means that the bubble-point of the liquid also is increased successively. If equilibrium between the liquid and the vapour phases is assumed along the whole heat exchanger, the temperature continues to increase until it reaches the dew-point of the vapour phase with the same mole fractions as the initial liquid phase. A corresponding discussion for the condensation process would explain the temperature decrease. The resulting smaller difference between the condensing and evaporating temperatures when a mixture is used (Fig. 1: To, and Te= are logarithmic mean temperatures of T3 and T~ or rather 7"1 and T2) indicates that the ratio between the compression work and the heating capacity can be decreased and hence that the coefficient of performance (COP) can be improved (COP = beat output in the condenser divided by the compression work). The ix~sibititi~sof increasing the COP are best for those heat pumps in which the temperature decrease of the heat source and the temperature increase of the heat sink are considerable. On the other hand this circumstance leads to smaller temperature differences in the evaporator and the condenser (between heat source/sink and refrigeran0 so that both o f these apparatus have to be designed to be larger. Furthermore they have to he arranged as countercurrent heat exchangers. In addition to the possibilities of improving the COP there are several other aspects of the heat pump cycle that may be improved. Here for instance the heating capacity can be increased by adding a fluid with a high density to a refrigerant. The pressure levels in the condenser and the evaporator can be adjusted to appropriate values and the compressor discharge temperature can be lowered for a given condensing temperature. 3. THE M A T H E M A T I C A L MODEL FOR THE D E T E R M I N A T I O N OF THE ECONOMY 3.1. Thermodynamic process simulation For the simulation of thermodynamic processes a computer program has ~ developed which solves the system of nonlinear equations of the steady state process flowsbeet by a sequential method [4]. The organisation of this program is such that a variety of thermodynamic cycleS, or rather the processes of different apparatus configurations, can be s i m u l ~ , Furthermore different data systems can be added to the program. Concerning the results publishedin thisartidethe state properties of the hydrocarbons and rcfrigerants (pure substances and mixtures) :were calculated by the equations of Lee and Kesler [5] and PlSeker [6].
Economy of nonazeotropic refrigerant mixtures
315
3.2. Apparatus design For the lay-out of the apparatus (evaporators. condensers, heat exchangers, valves and separators) another computer program is used. This program needs information about the thermodynamic process, which means that the process simulation program has to be executed in advance. More precisely speaking the latter program creates an input data file for the design program (Fig. 3). The geometry data of the apparatus and the external application conditions, such as temperatures of heat source and heat sink, have to be given separately. The program is organized in such a way that alternative models for different apparatus types can be linked to the system (e.g. pool boiling evaporator, boiling within horizontal tubes of falling film evaporator etc.) 3.3. Calculation of economy If the apparatus is designed the investment costs of the heat pump can be calculated. Here cost functions are used that have been developed at the Institute of Chemical Engineering, Technical University Graz [3]. The economical effectiveness of a heat pump is expressed by the pay back period. It is considered that the recovered heat alternatively is produced by a conventional heating device. Then the savings of the heat pump are mainly related to the costs of heating fuels conserved and result from the difference between the running costs of the conventional heater (oil fired) and those of the heat pump. The payback period again is calculated by an individual computer program where the required data are transferred automatically from the apparatus design program (Fig. 3). 3.4. General arrangement of the lay-out programs, economic optimization At the calculation of the economy of a heat pump the stated programs are executed in series (Fig. 3). They are organized in a way that they read all input data from only one data file. On this file the data are summarized into logical units that are identified by keywords. The execution of the programs happens from a command routine so that the whole system can be used if there was only one program. In addition a maximization of specific values can be done by a general maximization algorithm, which also is programed in separate programs. The maximization routine changes specific values in the input data file that represent the degrees of freedom of the optimization (Fig. 3). If for instance those are in the field of the process design and the payback period shall be minimized, both the process simulation program as well as the apparatus design program and the economy program have to be executed after each modification of one degree of freedom. 4. B O U N D A R Y CONDITIONS OF COMPRESSION HEAT PUMPS FOR I N D U S T R I A L A P P L I C A T I O N S - - G E N E R A L ASSUMPTIONS FOR TWO REPRESENTATIVE CASE STUDIES Some detailed investigations of heat pumps applied in industries and some studies of possible applications have been published in [2, 7]. Here 80 case studies have been carried out, covering applications in 16 branches of industry, such as the chemical industry, metals and wood processing,
siProcess mutotlont~\
f
1
',I
Apporotus izotl thm designl~I___ .=JaLgOri
Fig. 3. Organisation of computer programs for the economical optimization.
H. OENER et
316
W . o
Ap=0.2 bor " 90"C I
~
W
al.
O
~
.
Ap=0.2 Dor i?0° C
~
Ot'er
4~.c
41 *C 45"C 6o'c~) Ap'-0 2 bar Ap=0.2 bor Cose T Cose TT Fig. 4. Boundary conditions for the case studies.
food industry and others. Concerning the temperature of heat sources, it turned out that in 28% of the cases the heat source has to be cooled 0-5 ° (mainly sources of latent heat), in 14% 5-I0 °. in 48% 10-20 ° and in 10% more than 20 °. Starting out from these studies the boundary conditions of two cases were defined that seem to be representative of a high number of possible applications (Fig. 4). In case I the heat source is water that is cooled from 60 to 45°C. The heat is recovered to heat water from 70 to 90°C. In case II water is cooled from 45 to 41°C and the heat sink, also water, is heated from 60 to 70°C. In both cases the heat output in the condenser is 5000 kW. In order to keep the assumptions as general as possible the fluid flow of the heat source and the heat sink through the evaporator and the condenser is defined by the pressure loss. (With the pressure loss the velocities are defined which strongly influence the heat transfer coefficient and in this way the costs of the apparatus.) In addition the apparatus types have to be defined. Here three configurations of apparatus are studied (Figs 5-7). In configuration A a pool boiling evaporator and a horizontal tube bundle condenser are used. The subcooler and the intemaediate beat exchanger are designed as shell and tube heat exchangers. In the configuration B the evaporator is a horizontal tube bundle evaporator with the evaporating medium inside the tubes and in configuration C both the evaporator and the condenser are falling film apparatus. The compressors are screw compressors driven by electric motors. 5. E C O N O M I C A L COMPARISON OF COMPRESSION PUMPS OPERATING WITH PURE REFRIGERANTS NONAZEOTROPIC MIXTURES
HEAT AND
An important fact concerning the comparison of different machines are the criteria the comparison will be based on. It is considered here that only economically optimized machines shall
Fig. 5. Apparatus configuration A,
Economy of nonazeotropic refrigerant mixtures
317
-n~errnediote heat exchonoer
Fig. 6. Apparatus configuration B.
be compared among themselves. One criterion considered for a particular working fluid economically optimized heat pump, is the machine with the shortest payback period. That means that for each working fluid the thermodynamical process has to be found (process temperatures, pressure losses etc.) as well as the equipment (tube diameters, baffle spaces in the heat exchangers etc.) that result in the heat pump with the shortest payback period. Considering all process and design parameters variable for an optimization one obtains a huge multidimensional optimization problem. In order to reduce the high number of degrees of freedom the influences of the single parameters on the payback period were studied one by one and finally only those parameters which influence the payback period strongest were kept variable. It turned out that the choice of values concerning the equipment do not depend on the working fluid. They have to be chosen due to special guidelines but they are not parameters that have to
Condenser
5ubcooLer
~-JII I I
-
LZntermediote
heat. e x c l ~ l e r
Seperotor
Evoporotor
Fig. 7. Apparatus configuration C. H.R.$. 6/4--D
318
H
OFNER et al.
be optimized. For instance for convective boiling within horizontal tubes the heat transfer with all working fluids is best using the smallest tube diameters (assuming constant pressure losses). This means that the smallest tube the apparatus can be manufactured with has to be chosen. All the guidelines concerning the equipment are discussed in Ofner [3]. The parameters that define the thermodynamical process have to be optimized by all means. Here it is absolutely necessary that for instance the condensation and the evaporation temperatures (or in other words the temperature difference between heat sink source and the refrigerant in the condenser/evaporator) have to be chosen on the basis of economical aspects. Smaller temperature differences in the condenser and evaporator lead to higher COP but also higher investment costs. For that reason these temperatures are for all working fluids adjusted in such a way that one obtains the mimmum payback period of the whole machine. It is similar with the pressure losses of the refrigerant flowing through the apparatus. Higher pressure losses mean higher velociues and better heat transfer but inferior COP. In Table 1 the coefficients of performance (COP), the investment costs (IC) and the payback periods (t) are summarized for the apparatus configurations A, B and C (Figs 5-7), different working fluids (pure fluids) and the boundary conditions of case I (Fig. 4). For all detailed assumptions concerning the equipment and the economy the reader is referred to [3]. T h e media are chosen because the evaporation pressure lies above 1 bar and the condensation pressure does not exceed 30 bar. R114 is the substance that today is used without exception in this temperature range. It is not toxic, not inflammable or explosive and is stable at high temperatures. But, as one can see from Table 1, the physical and the thermodynamical properties are rather poor. Compared with other substances, expensive equipment is required and only low COPs can be achieved. Here with R21 and butane more economical machines can be laid out, but, as a matter of fact. these fluids can lead to other problems if used in heat pumps. Butane is highly inflammable and explosive and with R21 the compressor discharge temperture is higher than the proposed limit, for reasons of stability. Furthermore it is toxic and soluble in the compressor oil. Regarding all aspects it is questionable if the better economy data of R21 and butane really compensate for these disadvantages. In the following an attempt has been made to find better working fluids by using refrigerant mixtures. Here first of all the consequences of increasing the volumetric efficiency of R114 by adding R12 are studied (pure RI2 cannot be used in the present temperature range because of it's rather low critical temperature To,,,= 112°C). From Table 2 it can be seen that only with the equipment of the configuration B improvements are achieved. With a pool boiling evaporator (configuration A) no counter current heat exchange is possible. As at the nonazeotroplc mixture the temperature increases during the evaporation the evaporator inlet temperature has to be lowered so that the whole evaporation takes place below 45°C. This leads to worse COPs with
Table 1. Case 1 (Fig. 4) mixtures R114/R12 (investment costs (IC) in t01 Austrian Shillings, payback periods (t) in yr)
R114/RI2 (mol %) RII4 80/20 60/40 40/60 20/80
Configuration 8
A
. . . . . . . . . . . . . . COP I C . . . . t___~.f_op _ i c
5,01 4,90 4.78 4.73 4.94
10188 10719 10928 10898 11556
3.14 3.4I 3.60 3.65 3.63
5.08 5.35 5.60 5.39 5,27
11041 11589 11808 11205 11295
.
.
.
t
. . . . . .. c0£ if
3,34 3.28 3.18 3.14 3.26
5.05 5.15 4.86 4.97 4.88
C
. ,
10661 11094 10564 11323 11248
3.25 3.29 3.40 3.53 3~59
Table 2. Case I (Fig. 4) pure fluids (investment costs (IC) in 103 Austrian Shillings, Configuration B
A
Fluid
COP
lC '
RII R21 Rl14 pentane butane
4.99 5.18 5,01 5.16 5.12
11524 9543 10188 13442 8915
~
COP
ic-
3.57 2.81 3.14 3.98 2,67
5.27 5.20 5.08 5.40 5.26
11352 9562 11041 12171 9412
C
t -
?:oi"
)c
t
3.27 2.80 3.34 3.41 2.74
5.22 5.21 5.05 5.16 5.17
11235 9883 10661 12419 9821
3.28 2.89 3.25 3.68 2,90
Economy of nonazeotropic reffii~erant mixtures
319
Table 3. Case I (Fig. 4) mixtures Rl l/Rl2 (investment costs (IC) in l03 Austrian Shillings, payback periods (t) in yr) Configuration B RII/RI2 (mol%)
RI1 80/20 60/40 40/60 20/80
C
COP
IC
t
COP
IC
t
5.27 5.51 5.78 5.59 5.63
11352 10827 11424 11439 11463
3.27 2.96 2.97 3.09 3.07
5.22 5.25 4.97 5.04 4.70
11235 11630 11300 11683 10837
3.28 3.37 3.52 3.57 3.67
mixtures. By increasing the volumetric efficiency smaller (and cheaper) compressors can be used but due to the mixture larger heat exchangers are required (lower temperature differences, and lower heat transfer coefficient because of the combined heat and mass transfer, lower heat transfer coefficient because of inferior physical properties of the mixture). With falling film apparatus, counter current heat exchange is possible. In spite of that, no improvements can be observed with the mixtures by using the apparatus configuration C. This is the case as a total evaporation is impossible here. It is considered that only 70% evaporation takes place (mass vapour to mass vapour plus liquid is 0.7 at the evaporator outlet). This circumstance does not effect any disadvantages as far as fluids with isothermic evaporation are concerned. With nonazeotropic mixtures, however, the recycle stream from the separator (liquid stream from the separator to the evaporator inlet, Fig. 7) has a higher temperature than the stream at the outlet of the expansion valve. This leads to disadvantages for the whole process design that effect the low COPs. In the equimolar region of the mixture R114/R12 the temperature increase in the evaporator is up to 6 ° (taking into account the pressure loss and the partial evaporation in the expansion valve that lower the temperature increase in the evaporator). As in this case study the heat source is cooled 15° and an optimal adaptation of the process to the boundary condition is impossible with R114/R12. By mixing the components R11 and R12 one can get up to 15° temperature increase during evaporation. In Table 3 the results for different concentrations of R11/RI2 are summarized. Here only the configurations B and C are laid out. For reasons stated before no improvements can be achieved with falling film apparatus. Specifying a horizontal tube bundle evaporator the COP can he raised 10% compared with pure R11. It has to be noticed here that at the comparison oi theoretical (reversible) processes (Carnot process--Lorenz process) for the given boundary conditions due to the mixture a maximum of 45% improvement is possible. As the comparison conditions are totally different comparing optimized machines (minimum payback periods) the improvement is only 10% as now different real processes are compared among themselves. Using mixtures the heat transfer in all heat exchangers is worse. This means that more expensive equipment would be necessary, but for that reason the minima of the payback periods appear at higher values for the temperature differences in the condensers and evaporators and also at higher pressure losses for the refrigerant flow through the heat exchangers. Now the equipment is hardly more expensive for the use of the mixtures, but, on the other hand the increase of the COP is not as high as one may expect from theory. In Table 4 the results for mixtures of R21/R114 are shown. This mixture has an azeotrope at Table 4. Case ! (Fig. 4) mixtures R21/R114
(compressor discharge temperature (Ta.~_) in °C, investment costs (IC) in 10~ Austrian Shillings, payback periods (t) in yr) Configuration R21/R114 (tool %) T ~ . ¢ ~ COP IC R21 80/20 60/40 40/60 20/80 R114
132 120 Ill 104 100 96
5.20 5.18 5.12 5.18 5.1 ] 5.08
9562 9703 10002 10635 10834 11041
t
2.80 2.86 3.00 3.14 3.25 3.34
320
H.O~,~R
et al.
Table 5. Case II (Fig. 4) pure fluids (investment co6ts (IC) in l0 s Austrian Shillings. payback periods (t) in yr) Configuration B
A Fluid RII R21 R114 R12 pentane butane
COP 6.01 6.07 6.29 6.12 5.98 6.37
IC-13265 9729 10849 9445 15235 9268
t
COP
IC--
3.33 2.42 2.61 2.33 3.84 2.21
6.09 6.25 6.19 6.19 6.81 6.31
11360 9465 10524 9897 13317 9085
C
t
COP- IC-
2.82 2.29 2.57 2.42 3.01 2.18
6.04 6.53 6.M 6.02 6.48 6.42
11534 10205 10943 10470 14094 9983
2.88 2.38 2.62 2.62 3.31 2.36
Table 6. Case II (Fig. 4) mixtures R114/R12 (investment costs (IC) in 103 Austrian Shillings, payback periods (t) in yr) R114/RI2 (mol %) RII4 80/20 60/40 40[60 20180 RI 2
Configuration B
A
C
COP
IC
t
COP
IC
t
COP
IC
t
6.29 5.69 5.55 5.75 5.75 6.12
10849 10735 10953 t 1052 10189 9445
2.61 2.84 2.98 2.89 2.67 2.33
6.19 6.07 6.19 6.27 6.27 6.19
10524 10495 10672 10504 10186 9897
2.57 2.61 2.61 2.54 2.46 2.42
6.34 6.29 5.84 5.84 5.85 6.02
10943 11148 10552 10393 10357 10470
2.62 2.69 2.72 2.68 2.67 2.62
about 60% R114 (by mole), but the temperature increases during evaporation are very small at all concentrations. Beside the values already described here the compressor discharge tm~peratures (at diabatic compression) are also given in the table. The high temperature of R21 can be lowered drastically by adding more R114. This could be of practical importance as R21 should not be used above 121°C. Table 5 shows the results for pure fluids and the boundary conditions of ease II. In this temperature range the frequently proved refrigexant R12 can be used and it indeed seems to be the best refrigerant here too. The temperatures of the boundary conditions of case II are defined in such a way that with no refrigerant mixture improvements could be achieved. Table 6 shows the results for mixtures of Rl 14/R 12. 6. C O N C L U S I O N S The research's [3] purpose was to find the most economical compression heat pump for various industrial applications by using appropriate working fluids. Here first of all machines were laid out with different pure fluids, primarily considering the currently available refrigerants. Afterwards an attempt was made to find better working fluids by using refrigerant mixtures. It was questionable if with mixtures that lead to improved tbermodynamic heat pump cycles also more economical machines can be designed. During the design of the equipment special emphasis was laid on the economical optimization. Concerning heat pumps with different working fluids only economically optimized machines were compared among themselves. Furthermore the applicability of some condenser and evaporator types for different working fluids was studied. The investigations were made on the basis of two case studies that seem to be representative of a high number of possible industrial applications for compression heat pumps (Fig. 4). In one case the temperatures of the boundary conditions for the heat pump cycle were defined in a way that due to a nonazeotropic mixture a maximum o f 45% improvement for the coefficient of performance is possible (comparison of the reversible eyclcs: Carnot proces~Lorenz process). Here heat pumps with nonazcotropic mixtures were designed that clearly show better performance data compared to the machines with the conventional refrigerants. The co¢tfacient of performance could be improved by up to 10% with equal investment costs. In this way the payback period could be shortened by between 5 and 10%. Due to the boundary conditions no improvements could be observed with mixtures in the second ease study. Besides the attempt at improving the performance it was also tried to achieve more economical machines by raising the heating capacity. The idea was that at high volumetric eiticiences cheaper
Economy of nonazeotropic refrigerant mixtures
321
equipments are needed if one particular heat o u t p u t in the condenser is required. It turned out that cheaper compressors could indeed be used, but the heat exchanger costs increased due to the worse heat transfer with the use o f mixtures, so that the investment costs o f the machine could not be lowered at all. Concerning the a p p a r a t u s types the evaporators are o f main importance. With the use o f mixtures once-through evaporators are necessary (Fig. 6, convective boiling in tubes). With pool boiling evaporators no countercurrent heat transfer is possible and evaporators o f the falling film type are less a d v a n t a g e o u s as total evaporation is impossible. REFERENCES 1. W. Eder and F. Moser, Die Wdrmepumpe in der Verfahrenstechn&, Springer, Wien (1979). 2. F. Moser and H. Schnitzer, Energieeinsparung durch Wirmepumpen in Industrie und Gewerbe. Energiepolitische Schriftenreihe, herausgegeben yore Bundesministeriumf Handel, Gewerbeund lndustrie in Wien, Band 7. Springer, Wien (19Z3). 3. H. Ofner, Die Wirtschafllichkeit yon Kompressionswiirmepumpen mit nichtazeotropen Gemischen als Wfirmetrigermedium im industriellen Einsatz. Dissertation, Technische Universitfit Graz (1985). 4. H. Ofner, H. Schnitzer and F. Moser, Adaptation of general purpose flowsheet program for the design of energy recovering systems. Paper Presentation at 3rd Chemical Engineering Conf., 14-16 Sept., Graz (1982). 5. B. I. Lee and M. G. Kesler, A generalized thermodynamic correlation based on three-parameter corresponding states. AICHE J. 21, 510 and 1237 (1975). 6. U. Pl6cker, Berechnung von Hochdruckphasengleichgewichten mit einer Korrespondenzmethode unter besonderer Beriicksichtigung asymmetrischer Gemische. Dissertation, Technische Universit//t Berlin (1977). 7. H. Schnitzer and F. Moser, Experiences in the application of industrial heat pumps--conclusions from 80 case studies. Paper Presentation at International Symp. on the Industrial Application of Heat Pumps, Coventry, England (1982).
0198-7593/86 $3.00+ .00 Pergamon Journals Ltd
THE ECONOMY OF NONAZEOTROPIC REFRIGERANT MIXTURES IN COMPRESSION HEAT PUMPS H. OFNER, H. SCHNITZERand F. MOSER Institute of Chemical Engineering, Technical University Graz, Inffeldgasse 25, A-8010 Graz, Austria
(Received 12 January 1986) Abstract--The relative economic advantages of heat pumps using nonazeotropic mixtures are examined.
1. I N T R O D U C T I O N In heat pumps with pure refrigerants as working fluids both the condensation and evaporation take place at constant temperature if the pressure drop is neglected. With a nonazeotropic mixture, however, the temperature will decrease during condensation and increase during evaporation. For that reason the performances of heat pumps can be improved, especially in applications with high temperature gradients of the heat sink and the heat source. This is one phenomenon that, in addition to other advantages of nonazeotropic mixtures, is described in this article. Especially in the process industries there are various possible applications of heat pumps for the recovery of waste heat [1]. Before reporting technical and economical evaluations on such machines reference is made to some basic work where at a high number of industrial plants the temperatures and the mass flowrates of waste heat streams were studied [2]. Here also the temperatures to which the heat has to be transformed to render it usable in the plant and save the maximum amount of primary energy had been worked out. From this study the most frequent boundary conditions for compression heat pumps in industrial plants are determined and machines are laid out with different pure working fluids and fluid mixtures. The economical effectiveness of the heat pumps is expressed in terms of the payback periods. In the article the main results of the author's doctoral thesis are described. For the details concerning the mathematical model, etc, see Ofner [3]. 2. T H E R M O D Y N A M I C A L B A C K G R O U N D - - P O S S I B L E IMPROVEMENTS WITH NONAZEOTROPIC MIXTURES One important thermodynamic consequence for a heat pump, when a binary nonazeotropic mixture is used instead of a pure working fluid, is shown schematically in Fig. 1. Here the
T L
4
Purefluid ~
3 Condensotion
T.
T.. T,
Evoporotion
Purefluid Heat fLux
Fig. 1. Temperature variations of working fluids during condensation and evaporation (pure fluids and non-azeotropic mixtures). 313
314
H.
O F N E R et aL
Dew-point curve BubbLe-point curve
Xl
X2
Concent rot ion
Fig. 2. The evaporation of a non-azeotroptc mtxture in the T,x-diagram.
temperature variations of the working fluid, the heat source and the heat sink are plotted in a temperature-heat diagram. In the case of a pure refrigerant the temperatures during the phase changes in the condenser and evaporator are constant if the pressure drop is neglected. For a nonazeotropic mixture the temperature wilt decrease along the condensation line and increase along the evaporation line. The reasons for these temperature variations are as follows. For a given pressure a nonazeotropic mixture will have bubble-point and dew-point curves as shown in principle in the T,x-diagram, Fig. 2. When a liquid is vapourized, the more volatile component will be enriched in the vapour phase, thus successively inere,asing the mole fraction for the less volatile component in the liquid phase. This means that the bubble-point of the liquid also is increased successively. If equilibrium between the liquid and the vapour phases is assumed along the whole heat exchanger, the temperature continues to increase until it reaches the dew-point of the vapour phase with the same mole fractions as the initial liquid phase. A corresponding discussion for the condensation process would explain the temperature decrease. The resulting smaller difference between the condensing and evaporating temperatures when a mixture is used (Fig. 1: To, and Te= are logarithmic mean temperatures of T3 and T~ or rather 7"1 and T2) indicates that the ratio between the compression work and the heating capacity can be decreased and hence that the coefficient of performance (COP) can be improved (COP = beat output in the condenser divided by the compression work). The ix~sibititi~sof increasing the COP are best for those heat pumps in which the temperature decrease of the heat source and the temperature increase of the heat sink are considerable. On the other hand this circumstance leads to smaller temperature differences in the evaporator and the condenser (between heat source/sink and refrigeran0 so that both o f these apparatus have to be designed to be larger. Furthermore they have to he arranged as countercurrent heat exchangers. In addition to the possibilities of improving the COP there are several other aspects of the heat pump cycle that may be improved. Here for instance the heating capacity can be increased by adding a fluid with a high density to a refrigerant. The pressure levels in the condenser and the evaporator can be adjusted to appropriate values and the compressor discharge temperature can be lowered for a given condensing temperature. 3. THE M A T H E M A T I C A L MODEL FOR THE D E T E R M I N A T I O N OF THE ECONOMY 3.1. Thermodynamic process simulation For the simulation of thermodynamic processes a computer program has ~ developed which solves the system of nonlinear equations of the steady state process flowsbeet by a sequential method [4]. The organisation of this program is such that a variety of thermodynamic cycleS, or rather the processes of different apparatus configurations, can be s i m u l ~ , Furthermore different data systems can be added to the program. Concerning the results publishedin thisartidethe state properties of the hydrocarbons and rcfrigerants (pure substances and mixtures) :were calculated by the equations of Lee and Kesler [5] and PlSeker [6].
Economy of nonazeotropic refrigerant mixtures
315
3.2. Apparatus design For the lay-out of the apparatus (evaporators. condensers, heat exchangers, valves and separators) another computer program is used. This program needs information about the thermodynamic process, which means that the process simulation program has to be executed in advance. More precisely speaking the latter program creates an input data file for the design program (Fig. 3). The geometry data of the apparatus and the external application conditions, such as temperatures of heat source and heat sink, have to be given separately. The program is organized in such a way that alternative models for different apparatus types can be linked to the system (e.g. pool boiling evaporator, boiling within horizontal tubes of falling film evaporator etc.) 3.3. Calculation of economy If the apparatus is designed the investment costs of the heat pump can be calculated. Here cost functions are used that have been developed at the Institute of Chemical Engineering, Technical University Graz [3]. The economical effectiveness of a heat pump is expressed by the pay back period. It is considered that the recovered heat alternatively is produced by a conventional heating device. Then the savings of the heat pump are mainly related to the costs of heating fuels conserved and result from the difference between the running costs of the conventional heater (oil fired) and those of the heat pump. The payback period again is calculated by an individual computer program where the required data are transferred automatically from the apparatus design program (Fig. 3). 3.4. General arrangement of the lay-out programs, economic optimization At the calculation of the economy of a heat pump the stated programs are executed in series (Fig. 3). They are organized in a way that they read all input data from only one data file. On this file the data are summarized into logical units that are identified by keywords. The execution of the programs happens from a command routine so that the whole system can be used if there was only one program. In addition a maximization of specific values can be done by a general maximization algorithm, which also is programed in separate programs. The maximization routine changes specific values in the input data file that represent the degrees of freedom of the optimization (Fig. 3). If for instance those are in the field of the process design and the payback period shall be minimized, both the process simulation program as well as the apparatus design program and the economy program have to be executed after each modification of one degree of freedom. 4. B O U N D A R Y CONDITIONS OF COMPRESSION HEAT PUMPS FOR I N D U S T R I A L A P P L I C A T I O N S - - G E N E R A L ASSUMPTIONS FOR TWO REPRESENTATIVE CASE STUDIES Some detailed investigations of heat pumps applied in industries and some studies of possible applications have been published in [2, 7]. Here 80 case studies have been carried out, covering applications in 16 branches of industry, such as the chemical industry, metals and wood processing,
siProcess mutotlont~\
f
1
',I
Apporotus izotl thm designl~I___ .=JaLgOri
Fig. 3. Organisation of computer programs for the economical optimization.
H. OENER et
316
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Ap=0.2 bor " 90"C I
~
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al.
O
~
.
Ap=0.2 Dor i?0° C
~
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41 *C 45"C 6o'c~) Ap'-0 2 bar Ap=0.2 bor Cose T Cose TT Fig. 4. Boundary conditions for the case studies.
food industry and others. Concerning the temperature of heat sources, it turned out that in 28% of the cases the heat source has to be cooled 0-5 ° (mainly sources of latent heat), in 14% 5-I0 °. in 48% 10-20 ° and in 10% more than 20 °. Starting out from these studies the boundary conditions of two cases were defined that seem to be representative of a high number of possible applications (Fig. 4). In case I the heat source is water that is cooled from 60 to 45°C. The heat is recovered to heat water from 70 to 90°C. In case II water is cooled from 45 to 41°C and the heat sink, also water, is heated from 60 to 70°C. In both cases the heat output in the condenser is 5000 kW. In order to keep the assumptions as general as possible the fluid flow of the heat source and the heat sink through the evaporator and the condenser is defined by the pressure loss. (With the pressure loss the velocities are defined which strongly influence the heat transfer coefficient and in this way the costs of the apparatus.) In addition the apparatus types have to be defined. Here three configurations of apparatus are studied (Figs 5-7). In configuration A a pool boiling evaporator and a horizontal tube bundle condenser are used. The subcooler and the intemaediate beat exchanger are designed as shell and tube heat exchangers. In the configuration B the evaporator is a horizontal tube bundle evaporator with the evaporating medium inside the tubes and in configuration C both the evaporator and the condenser are falling film apparatus. The compressors are screw compressors driven by electric motors. 5. E C O N O M I C A L COMPARISON OF COMPRESSION PUMPS OPERATING WITH PURE REFRIGERANTS NONAZEOTROPIC MIXTURES
HEAT AND
An important fact concerning the comparison of different machines are the criteria the comparison will be based on. It is considered here that only economically optimized machines shall
Fig. 5. Apparatus configuration A,
Economy of nonazeotropic refrigerant mixtures
317
-n~errnediote heat exchonoer
Fig. 6. Apparatus configuration B.
be compared among themselves. One criterion considered for a particular working fluid economically optimized heat pump, is the machine with the shortest payback period. That means that for each working fluid the thermodynamical process has to be found (process temperatures, pressure losses etc.) as well as the equipment (tube diameters, baffle spaces in the heat exchangers etc.) that result in the heat pump with the shortest payback period. Considering all process and design parameters variable for an optimization one obtains a huge multidimensional optimization problem. In order to reduce the high number of degrees of freedom the influences of the single parameters on the payback period were studied one by one and finally only those parameters which influence the payback period strongest were kept variable. It turned out that the choice of values concerning the equipment do not depend on the working fluid. They have to be chosen due to special guidelines but they are not parameters that have to
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Fig. 7. Apparatus configuration C. H.R.$. 6/4--D
318
H
OFNER et al.
be optimized. For instance for convective boiling within horizontal tubes the heat transfer with all working fluids is best using the smallest tube diameters (assuming constant pressure losses). This means that the smallest tube the apparatus can be manufactured with has to be chosen. All the guidelines concerning the equipment are discussed in Ofner [3]. The parameters that define the thermodynamical process have to be optimized by all means. Here it is absolutely necessary that for instance the condensation and the evaporation temperatures (or in other words the temperature difference between heat sink source and the refrigerant in the condenser/evaporator) have to be chosen on the basis of economical aspects. Smaller temperature differences in the condenser and evaporator lead to higher COP but also higher investment costs. For that reason these temperatures are for all working fluids adjusted in such a way that one obtains the mimmum payback period of the whole machine. It is similar with the pressure losses of the refrigerant flowing through the apparatus. Higher pressure losses mean higher velociues and better heat transfer but inferior COP. In Table 1 the coefficients of performance (COP), the investment costs (IC) and the payback periods (t) are summarized for the apparatus configurations A, B and C (Figs 5-7), different working fluids (pure fluids) and the boundary conditions of case I (Fig. 4). For all detailed assumptions concerning the equipment and the economy the reader is referred to [3]. T h e media are chosen because the evaporation pressure lies above 1 bar and the condensation pressure does not exceed 30 bar. R114 is the substance that today is used without exception in this temperature range. It is not toxic, not inflammable or explosive and is stable at high temperatures. But, as one can see from Table 1, the physical and the thermodynamical properties are rather poor. Compared with other substances, expensive equipment is required and only low COPs can be achieved. Here with R21 and butane more economical machines can be laid out, but, as a matter of fact. these fluids can lead to other problems if used in heat pumps. Butane is highly inflammable and explosive and with R21 the compressor discharge temperture is higher than the proposed limit, for reasons of stability. Furthermore it is toxic and soluble in the compressor oil. Regarding all aspects it is questionable if the better economy data of R21 and butane really compensate for these disadvantages. In the following an attempt has been made to find better working fluids by using refrigerant mixtures. Here first of all the consequences of increasing the volumetric efficiency of R114 by adding R12 are studied (pure RI2 cannot be used in the present temperature range because of it's rather low critical temperature To,,,= 112°C). From Table 2 it can be seen that only with the equipment of the configuration B improvements are achieved. With a pool boiling evaporator (configuration A) no counter current heat exchange is possible. As at the nonazeotroplc mixture the temperature increases during the evaporation the evaporator inlet temperature has to be lowered so that the whole evaporation takes place below 45°C. This leads to worse COPs with
Table 1. Case 1 (Fig. 4) mixtures R114/R12 (investment costs (IC) in t01 Austrian Shillings, payback periods (t) in yr)
R114/RI2 (mol %) RII4 80/20 60/40 40/60 20/80
Configuration 8
A
. . . . . . . . . . . . . . COP I C . . . . t___~.f_op _ i c
5,01 4,90 4.78 4.73 4.94
10188 10719 10928 10898 11556
3.14 3.4I 3.60 3.65 3.63
5.08 5.35 5.60 5.39 5,27
11041 11589 11808 11205 11295
.
.
.
t
. . . . . .. c0£ if
3,34 3.28 3.18 3.14 3.26
5.05 5.15 4.86 4.97 4.88
C
. ,
10661 11094 10564 11323 11248
3.25 3.29 3.40 3.53 3~59
Table 2. Case I (Fig. 4) pure fluids (investment costs (IC) in 103 Austrian Shillings, Configuration B
A
Fluid
COP
lC '
RII R21 Rl14 pentane butane
4.99 5.18 5,01 5.16 5.12
11524 9543 10188 13442 8915
~
COP
ic-
3.57 2.81 3.14 3.98 2,67
5.27 5.20 5.08 5.40 5.26
11352 9562 11041 12171 9412
C
t -
?:oi"
)c
t
3.27 2.80 3.34 3.41 2.74
5.22 5.21 5.05 5.16 5.17
11235 9883 10661 12419 9821
3.28 2.89 3.25 3.68 2,90
Economy of nonazeotropic reffii~erant mixtures
319
Table 3. Case I (Fig. 4) mixtures Rl l/Rl2 (investment costs (IC) in l03 Austrian Shillings, payback periods (t) in yr) Configuration B RII/RI2 (mol%)
RI1 80/20 60/40 40/60 20/80
C
COP
IC
t
COP
IC
t
5.27 5.51 5.78 5.59 5.63
11352 10827 11424 11439 11463
3.27 2.96 2.97 3.09 3.07
5.22 5.25 4.97 5.04 4.70
11235 11630 11300 11683 10837
3.28 3.37 3.52 3.57 3.67
mixtures. By increasing the volumetric efficiency smaller (and cheaper) compressors can be used but due to the mixture larger heat exchangers are required (lower temperature differences, and lower heat transfer coefficient because of the combined heat and mass transfer, lower heat transfer coefficient because of inferior physical properties of the mixture). With falling film apparatus, counter current heat exchange is possible. In spite of that, no improvements can be observed with the mixtures by using the apparatus configuration C. This is the case as a total evaporation is impossible here. It is considered that only 70% evaporation takes place (mass vapour to mass vapour plus liquid is 0.7 at the evaporator outlet). This circumstance does not effect any disadvantages as far as fluids with isothermic evaporation are concerned. With nonazeotropic mixtures, however, the recycle stream from the separator (liquid stream from the separator to the evaporator inlet, Fig. 7) has a higher temperature than the stream at the outlet of the expansion valve. This leads to disadvantages for the whole process design that effect the low COPs. In the equimolar region of the mixture R114/R12 the temperature increase in the evaporator is up to 6 ° (taking into account the pressure loss and the partial evaporation in the expansion valve that lower the temperature increase in the evaporator). As in this case study the heat source is cooled 15° and an optimal adaptation of the process to the boundary condition is impossible with R114/R12. By mixing the components R11 and R12 one can get up to 15° temperature increase during evaporation. In Table 3 the results for different concentrations of R11/RI2 are summarized. Here only the configurations B and C are laid out. For reasons stated before no improvements can be achieved with falling film apparatus. Specifying a horizontal tube bundle evaporator the COP can he raised 10% compared with pure R11. It has to be noticed here that at the comparison oi theoretical (reversible) processes (Carnot process--Lorenz process) for the given boundary conditions due to the mixture a maximum of 45% improvement is possible. As the comparison conditions are totally different comparing optimized machines (minimum payback periods) the improvement is only 10% as now different real processes are compared among themselves. Using mixtures the heat transfer in all heat exchangers is worse. This means that more expensive equipment would be necessary, but for that reason the minima of the payback periods appear at higher values for the temperature differences in the condensers and evaporators and also at higher pressure losses for the refrigerant flow through the heat exchangers. Now the equipment is hardly more expensive for the use of the mixtures, but, on the other hand the increase of the COP is not as high as one may expect from theory. In Table 4 the results for mixtures of R21/R114 are shown. This mixture has an azeotrope at Table 4. Case ! (Fig. 4) mixtures R21/R114
(compressor discharge temperature (Ta.~_) in °C, investment costs (IC) in 10~ Austrian Shillings, payback periods (t) in yr) Configuration R21/R114 (tool %) T ~ . ¢ ~ COP IC R21 80/20 60/40 40/60 20/80 R114
132 120 Ill 104 100 96
5.20 5.18 5.12 5.18 5.1 ] 5.08
9562 9703 10002 10635 10834 11041
t
2.80 2.86 3.00 3.14 3.25 3.34
320
H.O~,~R
et al.
Table 5. Case II (Fig. 4) pure fluids (investment co6ts (IC) in l0 s Austrian Shillings. payback periods (t) in yr) Configuration B
A Fluid RII R21 R114 R12 pentane butane
COP 6.01 6.07 6.29 6.12 5.98 6.37
IC-13265 9729 10849 9445 15235 9268
t
COP
IC--
3.33 2.42 2.61 2.33 3.84 2.21
6.09 6.25 6.19 6.19 6.81 6.31
11360 9465 10524 9897 13317 9085
C
t
COP- IC-
2.82 2.29 2.57 2.42 3.01 2.18
6.04 6.53 6.M 6.02 6.48 6.42
11534 10205 10943 10470 14094 9983
2.88 2.38 2.62 2.62 3.31 2.36
Table 6. Case II (Fig. 4) mixtures R114/R12 (investment costs (IC) in 103 Austrian Shillings, payback periods (t) in yr) R114/RI2 (mol %) RII4 80/20 60/40 40[60 20180 RI 2
Configuration B
A
C
COP
IC
t
COP
IC
t
COP
IC
t
6.29 5.69 5.55 5.75 5.75 6.12
10849 10735 10953 t 1052 10189 9445
2.61 2.84 2.98 2.89 2.67 2.33
6.19 6.07 6.19 6.27 6.27 6.19
10524 10495 10672 10504 10186 9897
2.57 2.61 2.61 2.54 2.46 2.42
6.34 6.29 5.84 5.84 5.85 6.02
10943 11148 10552 10393 10357 10470
2.62 2.69 2.72 2.68 2.67 2.62
about 60% R114 (by mole), but the temperature increases during evaporation are very small at all concentrations. Beside the values already described here the compressor discharge tm~peratures (at diabatic compression) are also given in the table. The high temperature of R21 can be lowered drastically by adding more R114. This could be of practical importance as R21 should not be used above 121°C. Table 5 shows the results for pure fluids and the boundary conditions of ease II. In this temperature range the frequently proved refrigexant R12 can be used and it indeed seems to be the best refrigerant here too. The temperatures of the boundary conditions of case II are defined in such a way that with no refrigerant mixture improvements could be achieved. Table 6 shows the results for mixtures of Rl 14/R 12. 6. C O N C L U S I O N S The research's [3] purpose was to find the most economical compression heat pump for various industrial applications by using appropriate working fluids. Here first of all machines were laid out with different pure fluids, primarily considering the currently available refrigerants. Afterwards an attempt was made to find better working fluids by using refrigerant mixtures. It was questionable if with mixtures that lead to improved tbermodynamic heat pump cycles also more economical machines can be designed. During the design of the equipment special emphasis was laid on the economical optimization. Concerning heat pumps with different working fluids only economically optimized machines were compared among themselves. Furthermore the applicability of some condenser and evaporator types for different working fluids was studied. The investigations were made on the basis of two case studies that seem to be representative of a high number of possible industrial applications for compression heat pumps (Fig. 4). In one case the temperatures of the boundary conditions for the heat pump cycle were defined in a way that due to a nonazeotropic mixture a maximum o f 45% improvement for the coefficient of performance is possible (comparison of the reversible eyclcs: Carnot proces~Lorenz process). Here heat pumps with nonazcotropic mixtures were designed that clearly show better performance data compared to the machines with the conventional refrigerants. The co¢tfacient of performance could be improved by up to 10% with equal investment costs. In this way the payback period could be shortened by between 5 and 10%. Due to the boundary conditions no improvements could be observed with mixtures in the second ease study. Besides the attempt at improving the performance it was also tried to achieve more economical machines by raising the heating capacity. The idea was that at high volumetric eiticiences cheaper
Economy of nonazeotropic refrigerant mixtures
321
equipments are needed if one particular heat o u t p u t in the condenser is required. It turned out that cheaper compressors could indeed be used, but the heat exchanger costs increased due to the worse heat transfer with the use o f mixtures, so that the investment costs o f the machine could not be lowered at all. Concerning the a p p a r a t u s types the evaporators are o f main importance. With the use o f mixtures once-through evaporators are necessary (Fig. 6, convective boiling in tubes). With pool boiling evaporators no countercurrent heat transfer is possible and evaporators o f the falling film type are less a d v a n t a g e o u s as total evaporation is impossible. REFERENCES 1. W. Eder and F. Moser, Die Wdrmepumpe in der Verfahrenstechn&, Springer, Wien (1979). 2. F. Moser and H. Schnitzer, Energieeinsparung durch Wirmepumpen in Industrie und Gewerbe. Energiepolitische Schriftenreihe, herausgegeben yore Bundesministeriumf Handel, Gewerbeund lndustrie in Wien, Band 7. Springer, Wien (19Z3). 3. H. Ofner, Die Wirtschafllichkeit yon Kompressionswiirmepumpen mit nichtazeotropen Gemischen als Wfirmetrigermedium im industriellen Einsatz. Dissertation, Technische Universitfit Graz (1985). 4. H. Ofner, H. Schnitzer and F. Moser, Adaptation of general purpose flowsheet program for the design of energy recovering systems. Paper Presentation at 3rd Chemical Engineering Conf., 14-16 Sept., Graz (1982). 5. B. I. Lee and M. G. Kesler, A generalized thermodynamic correlation based on three-parameter corresponding states. AICHE J. 21, 510 and 1237 (1975). 6. U. Pl6cker, Berechnung von Hochdruckphasengleichgewichten mit einer Korrespondenzmethode unter besonderer Beriicksichtigung asymmetrischer Gemische. Dissertation, Technische Universit//t Berlin (1977). 7. H. Schnitzer and F. Moser, Experiences in the application of industrial heat pumps--conclusions from 80 case studies. Paper Presentation at International Symp. on the Industrial Application of Heat Pumps, Coventry, England (1982).