First industrial application of non-azeotropic mixture

First industrial application of non-azeotropic mixture* J. C. Blaise, T. D u t t o and J. L. A m b r o s i n o t § Electricit6 de France, Direction des Etudes et Recherches, B.P. N ° 1,77250 Moret S/Loing, France tInstitut Francais du P6trole, BP 311, 92506 Rueil-Malmaison Cedex, France Received 1 M a y 1989

Experiments were carried out on a heat pump which recovers energy from a refrigerating installation and warms water from 58 to 68°C. The aim of our experiment is to compare the COP and the thermal power of the heat pump using non-azeotropic mixture instead of R12. Therefore performances of the heat pump have been determined with R12 to obtain a reference, then R12 was replaced by a ternary mixture. The installation includes 31 sensors connected to a computer. We can evaluate the influence of non-azeotropic mixture on the volumetric and isentropic compression efficiencies and on the overall coefficient of heat transfer of the condenser. Tests with the mixture give the following results compared with R 12: + 20 % for the thermal power at the condenser; + 1.5 % for the COP; - 2% for the volumetric efficiency; - 10% for the overall coefficient of heat transfer at the same heat flux. The influence of refrigerant leakages on the mixture composition was studied. The leakages were simulated during running of the plant. It was noticed that the leakages at the input of the dry expansion evaporator and at the top of the receiver do not modify the concentration of the different components enough to change the working conditions of the heat pump. (Keywords: heat pump; non-azeotropiemixture; experimental results; leakage; R12 substitute)

La premi6re application industrielle d'un m61ange non az6otropique Des essais ont dtk mends sur une pompe it chaleur qui rdcupkre de l'dnergie it partir d'une installation frigorifique et qui chauffe de l'eau de 58 it 68°C. L'objectif de notre expdrience a dtd de comparer le COP et la puissance thermique de la pompe it chaleur en utilisant un mdlange non azdotropique plutdt que du R12. Aussi, on a ddtermind les performances de la pompe it chaleur avec du R12 afin d'obtenir une r@fdrence; puis, on a remplacd le R12 par un mdlange ternaire. L'installation comporte 31 capteurs connectds it un ordinateur. On a pu dvaluer l'infiuence d'un mdlange non azdotropique sur les rendements volumdtrique et isentropique et sur le coefficient de transfert de chaleur global du condenseur. Les essais effectuds avec le mdlange ont donnd les r@sultats suivants, comparks au R12: + 20% pour la puissance thermique au condenseur; + 1,5% pour le COP; - 2 % pour le rendement volumdtrique; - 1 0 % pour le coefficient de transfert de chaleur global it la m@me densitd de flux thermique. On a dtudid l'effet des fuites de frigorigdne sur la composition du mdlange. On a simuId les fuites au cours du fonctionnement de l'installation. On a constatO que les fuites, it l'entrde de l'dvaporateur it ddtente directe et it la partie supdrieure du rdservoir, ne modifiaient pas suffisamment la concentration des diffdrents composants pour pouvoir provoquer une modification des conditions de fonctionnement de la pompe it chaleur.

(Mots cl6s: pompe/t chaleur; m61angenon az6otropique; rbsultats exp6rimentaux; fuite; substitut du R12) Many controls, simulations and tests have been performed in laboratories or test rigs on non-azeotropic mixtures. Consequently, Electricit6 de France (EDF), the Institut Frangais du P6trole (IFP) and the company Q U I R I decided jointly to build a unit on an industrial site, using all their own expertise to optimize the operation and control. Such a test in an industrial environment is absolutely necessary to measure the reliability of this new technology.

2. a perfect counter-flow condenser; 3. electronic expansion valves to feed the dry-expansion evaporator. The heat pump includes 31 sensors on heat source (ammonia), on heat sink (water) and on refrigerant (R12 or non-azeotropic mixture). The flowmeters and the temperature and pressure gauges are connected to a computer. The installation has already been described in more detail 1.

Test installation Experiments were carried out on a heat pump which recovers energy from a refrigerating installation and warms water from 58°C to 68°C. As shown in Figure 1, the characteristics of the installation are: 1. two open type reciprocating compressors powered by electric motors of 7 5 k W ; the total swept volume is 496m 3 h -1 at 1450r.p.m.; * Revised version of a paper presented at the Purdue-IIR Conference 'Progress in the Design and Construction of Refrigeration Systems', Purdue University, West Lafayette, Indiana, USA, 18-21 July 1988 § To whom correspondenceshould be addressed 0140-7007/89/050255-O4503.00 © 1989 Butterworth & Co (Publishers) Ltd and IIR

Investigation method When we examined the measurements obtained, we were faced with all the problems of getting precise measurements in an industrial environment (meat salting factory) and particularly the problem of stability. The instability of the heat pump in operation is mostly due to variations of the ammonia condensing temperature in the refrigerant evaporator. In fact, the activity of salting varies during the day and this brings about important variation in the need of refrigeration and upsets the stability of the heat pump. So, in the measurements obtained, we looked for operating zones in which the variations of the ammonia pressure

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Industrial application of non-azeotropic mixture: J. C. Blaise et al. Wlte~

Figure 1 Schematic diagram of heat pump installation. (~), Temperature gauge; @ , pressure meter; ( ~ , flow meter; EGL, liquid-gas heat exchanger Figure 1 Schdma de l'installation de la pompe i~ chaleur.(~), Sonde de temperature; (~, Gauge de pression; ~ , d~bit m~tre; EGL ~changeur de chaleur liquide-gaz

simulation showed that the use of a ternary mixture would increase the heating capacity by 21% with a small increase of COP under an ammonia pressure equal to 12.5 bars. The C O P improvement is small because there is no temperature drop on heat source, and therefore no decrease of irreversibilities on the evaporator. So it is only due to the gliding temperature of i0 K in the counter-flow condenser. We can also observe the small influence of variations of the heat source on the advantages obtained with the non-azeotropic mixture. So a mixture is a fluid made to measure, but its advantages are maintained even when the working conditions differ from the initial ones. Volumetric efficiency Figure 4 shows the variations of volumetric efficiency versus compression ratio. First we notice that the values obtained are close to the values usually given for this type of compressor according to the pressure ratio between suction and discharge. For the same pressure ratio,

600

are small. We could then study the performance of the heat pump in relation to the temperature of the heat source. On a stable operating source, we determined an average operating point in order to calculate the performance. To evaluate the influence of a non-azeotropic mixture on the performance of a heat pump, compared with R12, we determined the following efficiencies for each fluid. 1. Volumetric efficiency: ratio of the flow-rate of refrigerant at the suction point to the swept volume of the compressor. 2. Isentropic efficiency: ratio of the power absorbed by an isentropic compression of the refrigerant to the power determined from the characteristics of the refrigerant at the suction and the discharge point, both for the same mass flow. 3. Coefficient of performance (COP): ratio of the heating capacity, Q, obtained at the heat sink to the electric power used by motors. 4. Overall coefficient of heat transfer of the condenser, h, (in kW °C -1 m -2) defined as:

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Q

DTL S ¢1.

where S=external surface of the condenser (in m2); Q = thermal power rejected by the condenser (in kW) and DTL=logarithmic mean temperature difference (in °C) between refrigerant and water. It is obtained from the logarithmic mean temperature differences of the desuperheating and condensing area z. Test results

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The curves for the average heating capacity and C O P of the heat pump versus ammonia pressure for the same input and output temperatures of water are shown in Figures 2 and 3. These results correspond to the expected values. In fact,

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Figure 3 COP versus ammonia pressure. • , R12; A, ternary mixture Figure 3 COP en fonction de la pression d'ammoniac. • , R12; A , m~lange ternaire

Industrial appfication of non-azeotropic mixture. J. C. Blaise et al. Table 1 Increase of thermal power and COP compared to R12 Tableau 1 Auomentation de la puissance thermique et du COP en comparaison avec le R12

PNH 3 (bar)

Heating capacity (%)

COP (%)

10.5 11.0 11.5 12.0

19.1 21.3 20.0 23.0

1.5 1.6 1.4 1.5

100

volumetric efficiency decreases about 2% when we use the non-azeotropic mixture instead of pure refrigerant R12. This small decrease of volumetric efficiency may be due to a variation of the ratio Cp/Cv (where Cp = heat capacity at constant pressure and Cv=heat capacity at constant volume), which is higher for the non-azeotropic mixture.

Isentropic efficiency The variation ofisentropic efficiency for different values of the pressure ratio is not observed from Figure 5. We can conclude that the mean isentropic efficiencies are 70% for pure refrigerant and 72% for ternary non-azeotropic mixture. So we can conclude that the use of this nonazeotropic mixture does not reduce isentropic and volumetric efficiencies of an open type reciprocating compressor.

Overall coefficient of heat transfer Experimental results (Figure 6) show a small decrease of e" lid

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Figure 4

Volumetric efficiency versus compression ratio, x, Nonazeo_tropic mixture; +, R12 Figure 4 Efficacitd volumdtrique en fonction du rapport de compression. x , Mdlange non-azdotropique;

+, R12

100

the overall coefficient of heat transfer of the condenser when we use the non-azeotropic mixture. On the heat pump equipped with a coiled condenser with bare pipes, the decrease can be equal to 10% for the same density of energy. But we can also compare the overall coefficient of heat transfer with the same temperature of the water at the input and the output of the condenser. To obtain the same temperature drop we make the water flow rate of the cold source vary in proportion to the thermal power rejected by refrigerant. In this case, the overall coefficient of heat transfer is about the same with R12 or non-azeotropic mixture. Figure 7 shows the variation of the temperature of R12 and non-azeotropic mixture with the same temperatures of water at the input and the output of the condenser (where Qtota~=thermal power ejected by the condenser, in kW). The same drop of temperature for water and non-azeotropic mixture is observed in the condenser. In this case, for the same input and output temperatures, the overall coefficient of heat transfer is 9 1 6 W m - 2 ° C for R12 and 9 3 3 W m - 2 °C for the nonazeotropic mixture. Influence of leakages After having evaluated the performance of a heat pump using a non-azeotropic mixture we studied the influence

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Figure 5 Isentropic efficiency versus compression ratio, x, Nonazeotropic mixture; +, R12 Figure 5 Efficacitd isentropique en fonction du rapport de compression.

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Overall coefficient of heat transfer of condenser• +, R12; x, ternary mixture Figure 6 Coefficient global de transfert de chaleur du condenseur. +,

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Industrial application of non-azeotropic mixture: J. C. Blaise et al.

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Table 2 Mass concentration after each leakage Tableau 2 Concentration massique apr~s chaquefuite.

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Mass concentration (%)

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At beginning

After leakage at input of

After leakage at top of

Component

of t e s t

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receiver

A B C

1.54 71.52 26.94

1.54 71.53 26.93

1.25 69.85 28.90

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Figure 7 Variation of temperature of R12 and non-azeotropic mixture with the same temperature of water at the input and output of condenser. Pure refrigerant R 12: Qtot~l= 413 kW; non-azeotropic mixture: Qtotal= 507 kW Figure 7 Variation de tempdrature du R12 et du mdlange non az~otropique aux rnOmes tempdratures d'eau ?t l'entr~e et d Ia sortie du condenseur. Frigorig~ne pur R12:Qtot.t=413 kW; mOlange non azkotropique: Qto~az=507 k W

of refrigerant leakage on the mixture composition. The leakages were simulated while running the plant. For this purpose, we created two leakages, one at the input of the dry-expansion evaporator (low pressure, liquid and gas phases) and the other at the safety valve of the receiver where gas is in equilibrium with liquid (high pressure). In order to determine the variations of concentration of all the components, we took a sample of 1 kg of liquid from the receiver before and after such leakage. A leakage of 30kg occurred at the input of the evaporator, the initial charge was 300 kg and the leakage lasted 4 h. The leakage at the top of the receiver lasted 4 h until gas crossed the expansion valve because refrigerant load of the heat pump became insufficient for good operation. The refrigerant loss was between 55 and 90 kg, i.e. 20 to 30% of the initial charge of 300 kg. Table 2 shows the mass concentration of each component of the ternary mixture at the end of each leakage. It is observed that the leakage at the input of the dry-expansion evaporator does not modify the concentration of the different components. These results obtained in an industrial installation agree with the results obtained in the laboratory 3. So we can conclude that the leakages created on a pipe, where a two phase mixture flows, do not modify the composition of a non-azeotropic mixture. The leakage at the top of the receiver causes a bigger decrease of the most volatile components A and B. We know the concentration of each component before and after the leakages, but we do not know the exact quantity

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Int. J. Refrig. 1989 Vol 12 September

of lost refrigerant. Although it is difficult to know the exact concentration of each component in the lost refrigerant, we can suppose that it is very close to the concentration of the mixture itself. Then, if we add the lost quantity of refrigerant, with initial concentration, to the remaining quantity, the concentration is about the same as the initial load. We can therefore conclude that leakages do not modify the composition of the mixture enough to change the working conditions of the heat pump. Conclusion

In this paper we have presented a series of measurements obtained on the first industrial heat pump using a non-azeotropic mixture of refrigerants. The tests carried out on the heat pump with a ternary non-azeotropic mixture and then with a pure fluid, R12, allowed us to verify the following: 1. the agreement of the increase of the heating capacity of COP with the expected values; 2. the absence of technical problems specific to the use of non-azeotropic mixtures (the same oil was used, no problem during charging refrigerant, starting or stopping compressors); 3. the very smal! influence of refrigerant leakage on the mixture composition. These tests have shown the reliability of a heat pump using a non-azeotropic mixture in an industrial environment in the case of refrigerant leakage. This type of mixture can be a substitute for R12, the use of which is controlled because of its adverse influence on the ozone layer. References 1 Blaise, J. C., Dutto, T., Ambrosino, J. L. Torreilles, M. An industrial application of non azeotropic mixture I l R Refrigeration Congress, Vienna, Austria (1987) 256--261

2 Kern,D. Process heat transfer McGraw-Hill Inc. New York, USA (1950) 3 Blaise, J. C., Dutto, T. Some practical results obtained with non-azeotropic mixture of refrigerants in high temperature heat pump, IIR CommissionE2, Purdue, Indiana, USA (1986)321-325