HCFC22-based absorption cooling systems, part III: Effects of different absorbers and condenser temperatures

INTERNATIONAL JOURNAL OF ENERGY RESEARCH, VOL. 20,483-494 (1996)

HCFC22-BASED ABSORPTION COOLING SYSTEMS, PART 111: EFFECTS OF DIFFERENT ABSORBER AND

CONDENSER TEMPERATURES M. FATOUH* AND S. SRINIVASA MURTHY

Rejkgemtion and Airronditioninghbomtoty, Depamant of Mechanical Engineehg Indian Institute of Techlogv, Madms 036, India

SUMMARY

Generally in a vapour absorption refrigeration system (VARS) heat rejection temperatures at absorber (T,) and condenser (T,) are taken to be equal. However, different temperatures can exist when the cooling water flows in series through the two components. Under such situations, it is essential to know which of Ta and T, has greater influence on the performance of the VARS. Here the influence of different T, and T, on the performance of a single-stage VARS working with HCFC22 as a refrigerant and three organic solvents, namely DMA, DMF and DEMTEG, as absorbents is studied. Results are obtained over a wide range of operating temperatures. To improve the performance of HCFC22-based VARS, results reveal that (i) the cooling water in parallel pipe connections should be used at low values of temperatures at evaporator, cooling water and heat source, and (ii) cooling water should first flow through condenser and then through the absorber when evaporator and heat source temperatures are high over the complete range of cooling water temperatures. COP,,, is more sensitive to T, than to T,.

KEY WORDS: HCFC22-absorbent pair; heat rejection temperatures; cooling water pipe connection; performance improvement

INTRODUCTION

Since in a vapour absorption refrigeration system (VARS) the condenser and absorber are cooled by the same medium, the temperatures in both absorber and condenser are usually taken to be equal. In practice cooling water may flow in parallel or in series through the condenser and the absorber of an absorption machine. If it is passed through them in parallel, their temperatures may be very nearly the same; but if it passed through them in series, one will operate at a higher temperature than the other.

Richter and Schumacher (1974) have reported that usually a parallel layout of cooling water supply through condenser and absorber is adopted in industrial NH,-H,O absorption plants. However, Kouremenos et al. (1989) have presented a compound NH,/H,O-H,O/LiBr absorption refrigeration system in which cooling water flows in series from absorber to condenser. Herold and Radermacher (1991) have mentioned that cooling water first goes to condenser then goes to the absorber in a series piping arrangement for aqueous ternary hydroxide working fluid in an absorption heat pump. From these observations it may be said that the direction of the cooling water path depends on the type of the working fluid.

Effects of operating temperatures on the performance of HCFC22-based VARS have been reported in the literature (Jelinek et al., 1978; Borde et al., 1979; Borde and Jelinek, 1987; Dan and Srinivasa Murthy, 1989; Fatouh and Srinivasa Murthy, 1995). Generally, both absorber and condenser temperatures are

*Permanent address: Mechanical Power Engineering Department, Faculty of Engineering and Technology, El-Mattaria, Cairo, Egypt.

CCC 0363-907X/96/060483-12 0 1996 by John Wiley & Sons, Ltd.

Received 9 Muy 1994 Revived 25 Muy 1594

484 M. FATOUH AND S. SRINIVASA MURTHY

assumed to be equal, i.e. cooling water flows in parallel through the absorber and condenser. However, the coefficient of performance and circulation ratio are presented for HCFC21-DMF by Kumar et al. (1991) for two cases: first one in which both absorber and condenser temperatures are equal; and second one in which the absorber temperature is higher than the condenser temperature.

Hence, the analysis presented here gives special attention to the possibility that temperatures of the absorber and the condenser can be euqal or different in HCFC22-based VARS. In practice, this can be achieved by connecting the cooling water pipe line to the absorber and condenser in parallel or in series. HCFC22-DMA, HCFC22-DMF and HCFC22-DMETEG as working fluids are used.

THEORETICAL CONSIDERATIONS

A schematic diagram of a single stage VARS is shown in Figure 1. The assumptions and analysis are presented by the authors in their earlier paper (Fatouh and Srinisava Murthy, 1995).

Since heat quantities and performance characteristics such as circulation ratio (CR), thermodynamic coefficient of performance (COP,,,) and second law efficiency ( E ) are referred to while discussing the results, relevant equations are tabulated here in Table I for ready reference. The thermodynamic property data for HCFC22, as well as HCFC22-DMA, HCFC22-DMF and HCFC22-DMETEG solutions, are taken from Fatouh and Srinisava Murthy (1993a, 1993b).

RESULTS AND DISCUSSIONS

Simulation studies are performed in the following ranges of operating temperatures:

- evaporator temperatures, T, = -10 to 10 "C - condensor temperature, T, = 20 to 50 "C

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Figure 1. Schematic diagram of continuous VARS

HCFC22-ABSORPTION SYSTEM 111 485

Table I. Equations of heat quantities and performance characteristics from (Fatouh and Srinisava Murthy, 1995)

Equation Equation no.

Note: Ta = T, , T, = T,, T, = TI, and Tg = T4 = T7, X,, = X , and Xws = X4; M, = M,, = M8 = M,, M,, = M4 and Mss = Ml

All subscripts refer to state points in Figure 1.

- absorber temperature, Ta = 20 to 50 "C - generator temperature, Tg = 60 to 50 "C - solution heat exchanger effectiveness, E,, = 0.85 - absorber mass transfer effectiveness, E, = 1.0 - generator mass transfer effectiveness Eg = 1-0 - cooling capacity, Q, = 1.0 kW

Here, the results for each HCFC22-absorbent pair are presented for three cases with respect to absorber and condenser temperatures as follows:

Case 1: in which both absorber and condenser temperatures are equal and vaxy in the range 20 to 50 "C. Case 2 in which condenser temperature is varied in the range 20 to 50 "C for an absorber temperature

Case 3: in which absorber temperature is varied in the range 20 to 50 "C for a condenser temperature of of 30 "C.

30 "C.

Influence of operating conditions is discussed with reference to the results for HCFC22-DMA system. The other two systems, HCFC22-DMF and HCFC22-DMETEG, follow the same trend.

Comparison of the three cases for HCFC22-DMA VARS at Tg of 60 and 90 "C, and T, of 0 and 10 "C are made in Figure 2. This is done to investigate which of the two heat rejection temperatures, the absorber temperature or condenser temperature, exerts greater influence on the performance. It can be mentioned here that temperature on the horizontal coordinate represents equal absorber and condenser temperatures for Case 1, condenser temperature for Case 2 and absorber temperature for case 3.

Before embarking on the comparative studies, the effects of the operating temperatures in the above cases on the heat quantities at solution heat exchanger, generator, absorber and condenser, and performance characteristics, namely circulation ratio, Carnot and thermodynamic coefficients of performance and second law efficiency, are discussed hereunder.

Pe$ormance of VXRS for equal Ta and T, (Case 1)

In practice, equal absorber and condenser temperatures can be achieved by connecting cooling water pipe lines in parallel through the absorber and condenser. The influences of equal absorber and


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condenser temperatures on performance are reported by the authors in their earlier paper (Fatouh and Srinavasa Murthy 1995). Hence, the results are used for comparison with the other two cases.

3.2 Effect of condenser temperature, T, (Case 2) The effects of condenser temperature when other operating parameters are kept constant can be studied from Figure 2.

Circulation ratio increases with condenser temperature as illustrated in Figure 2(A) and 2(B). When condenser temperature increases at a particular generator temperature, the weak solution concentration at the generator exit increases. Therefore. both numerator and denominator of equation (2) decrease. Since the rate of decrease of the numerator (1 - x w s ) is less than that of the denominator ( x s s -xws) , the circulation ratio increases with condenser temperature when other operating temperatures are kept constant. As a result both strong and weak solution mass flow rates also increase.

For a set of temperature levels at absorber, evaporator and generator, it is observed that both h, and h,, are not changed; refrigerant mass flowrate and h, increase; and h, and h6 decrease when the condenser temperature is increased.

As a result of the above, the value of (h , - h6) increases with condenser temperature. For the above reason and also because of the increase in weak solution mass flowrate with condenser temperature, from equation (3) one can predict that the solution heat exchanger duty increases. This trend is evident from Figures 2(C) and 2(D).

As h, increases and h6 decreases, the quantity ( h , - h6) increases. For this reason and also because of the increase in refrigerant mass flowrate, the first term in the right-hand side of equation (4) increases. When h6 decreases and h, is unchanged (because T, and P, are constant), the quantity (h6 - h , ) decreases. However, the rate of decrease in (h6 - h, ) is lower compared to the rate of increase in the strong solution mass flowrate. Therefore, the second term in the right-hand side of equation (4) also increases. Hence, Q, increases because of the increase in both terms in the right-hand side of equation (4) when T, increases. This is shown in Figures 2(E) and 2(F).

When T, increases, the quantity (hlo - h6) increases because h,, remains constant while h6 decreases. Also, the refrigerant mass flowrate increases because of the decrease in cooling effect. Thus, the first term in the right-hand side of equation ( 5 ) increases. It has already been mentioned, that the product of the strong solution mass flowrate and (h6 - h, ) increases. Thus, Q, increases with T, for given operating parameters, as shown in Figures 2(G) and 2(H).

Variation of the refrigerant mass flowrate with T, is similar to that of a vapour compression refrigeration system, i.e., as T, increases at constant T,, the refrigerant mass flowrate increases. However, the enthalpy difference (h , - h,) decreases slightly as the condenser temperature increases because h8 increases at a faster rate than h,. Therefore, Q,, the product of the mass refrigerant flowrate and enthalpy change, is slightly changed, as shown in equation (6).

From equation (71, when T, increases, both the numerator and the denominator increase. But the latter increases at a faster rate than the former, thus COP, decreases at a higher T,.

Variation of COP,,, with condenser temperature is given in Figures 2(I) and 2(J). It is clear that as T, increases, COP,, decreases. This is because of the increase in Q,. Since the constant cooling load is constant at 1 kW, COP decreases. This in fact is a mirror reflection of Figures 2(E) and 2(F).

Figures 2(K) and 2(L) give the variation of second law efficiency with T,. It is found that the rate of increase in the numerator due to increase in the term (T8/TlO) - 1) is higher than that in Q, in equation (9). As a result, second law efficiency increases initially. It is seen that 4 reaches a peak value and starts to fall at higher condenser temperature. This trend is due to the sharp increases in Q,, due to the sudden increase in the circulation ratio already seen in Figures 2(K) and 2(L).

Effect of absorber temperature, T, (Case 3)

Figures 2(A) and 2(B) indicate that the circulation ratio increases with absorber temperature. When the absorber temperature increases at constant evaporator temperature, the strong solution concentration

HCFCZZABSORPTION SYSTEM 111 489

decreases; thereby concentration spread decreases, which contributes to an increased circulation ratio. Hence, both strong and weak solution mass flowrates increase.

As the absorber temperature alone is changed, the refrigerant mass flowrate, which depends only on the condenser and evaporator temperatures, is unchanged. Also, enthalpies h,, h,, h8 and h,, are not significantly affected. But h6 and h, increase with absorber temperature. It can be noted that the rate of increase in h6 (absorber inlet) is lower than that in h, (absorber outlet). Based on the above changes in specific enthalpies, the values of enthalpy difference of (h6 - h,), ( h , - h6), (h,,, - h6) and (h , - h6) are found to decrease.

Because of rate of increase in weak solution mass flowrate is higher than that of decrease in ( h , - h6),

the solution heat exchanger duty increases with the absorber temperature when other parameters are held constant, as given by equation (3) and also shown in Figure 2(C) and 2(D).

From the abovementioned enthalpy changes, one can deduce that the first term in the right-hand side of equation (4) decreases because of the decrease in the value of ( h , - h,) and in the constant refrigerant mass flowrate. Also, the second term in the same equation increases because the rate of decrease in enthalpy difference ( h , - h , ) is lower than that of the increase in the strong solution mass flowrate. Thus the total effect of both terms of equation (4) is to increase heat input at the generator when the absorber temperature increases. This is illustrated in Figures 2(E) and 2(F).

As the absorber temperature increases, the heat rejected at the absorber increases, as given in Figures 2(G) and 2(H). The reason for this is that the rate of increase in the second term is higher than that of the decrease in first term in the right-hand side of equation (5). The first term decreases owing to the decrease in (h, , - h6) for constant refrigerant mass flowrate while the reasons for the increase in the second term are given in the previous paragraph.

Because the enthalpy difference ( h , - h,) and refrigerant mass flowrate remain constant when T, is varied, T, does not have any influence on Q,, as given by equation (6).

Based on equation (7), when To increases, the denominator increases while the numerator decreases owing to the reduction in the term (T, - TI) , both of which contribute to reducing COP,.

The thermodynamic coefficient of performance (COP,,) as a function of absorber temperature is plotted in Figures 20) and 2(J). It is clear that COP,,, decreases with To when other operating parameters are held constant. This is because of the increase in Q, with T,, as discussed with reference to Figures 2(E) and 2(F).

Figures 2(K) and 2(L) show the variation of second law efficiency ( E ) with T,. It is seen that second law efficiency increases to a maximum value then decreases. When the influence of the term (1 - (T,/T4)) is more pronounced than that of Q,, E increases, and vice versa.

Comparison among Cases 1, 2 and 3

It is important to note here that the curves intersect at the heat rejection temperature of 303.15 K (30 "C) where T, = T,, which corresponds to Case 1. Beyond this intersection point, T, > T, for Case 2 and T, > T, for Case 3. Before the intersection point, it will be vice versa, i.e. T, < T, for Case 2 and T, < T, for Case 3.

Comparison of the circulation ratio is made in Figures 2(A) and 2(B). At low generator temperature, it is observed that the lowest CR value can be obtained by Case 1 while both Cases 2 and 3 exhibit nearly the same values of CR over the complete range of operating conditions. At a higher generator temperature, as seen in Figure 2(B), Case 3 offers slightly lower CR, particularly when the condenser temperatures are higher than T,. This is due to the fact that the rate of decrease in concentration spread for Case 3 is slightly higher than that in Case 2.

Figures 2(C) and 2(D) show a comparison of solution heat exchanger duty for Cases 1, 2 and 3. It is clear that low Q,, can be achieved by Cases 1, 2 and 3 in that order when evaporator and generator temperatures are low, as presented in Figure 2(C). For high evaporator temperature at low generator temperature and low heat rejection temperature, there is no essential change in Q,, for the cases considered. At high generator temperature, it is clear from Figure 2(D) that change in solution heat


exchanger duty due to Cases 2 and 3 is negligible at a low heat rejection temperature. However, at high heat rejection temperatures Case 3 reveals lower Qh values while Case 1 yields the lowest Qh at low heat rejection temperatures. Therefore, to obtain low Qh the working conditions at which T, > T, are preferable at high generator temperatures. This suggests that for the HCF22-DMA refrigeration system a series cooling pipe line arrangement, in which cooling water flows first through the condenser and then through the absorber, is beneficial towards achieving low Qh when the heat source temperatures are high.

Figures 2(E) and 2(F) show a comparison of Q,. It is seen from Figure 2(E) that Case 2 shows lower Q, than Case 3 at high T, in the region where T, > T,, while the smallest value of Q, can be achieved by Case 1 at low evaporator temperature. Figure 2(F) for high generator temperature reveals that low Q, can be obtained by Case 2 before the intersection point and by Case 3 after the intersection point. As mentioned earlier, T, < T, for Caser 2 before the intersection point at low heat rejection temperature and also for Case 3 after the intersection point. As mentioned earlier, T, < T, for Case 2 before the intersection point. As mentioned earlier, T, < T, for Case 2 before the intersection point at low heat rejection temperature and also for Case 3 after the intersection point at high heat rejection temperature. Thus, for the considered range of evaporator temperatures, from the viewpoint of low Q,, series pipe line connections in which cooling water flows first through the condenser and then through the absorber may be recommended, particularly when the heat source temperature is high.

A cornparison of absorber heat loads is made in Figures 2(G) and 2(H). It is seen that a lower absorber heat load can be obtained by Case 2 before the intersection point, i.e., at low heat rejection temperatures. At high heat rejection temperatures, i.e., after the intersection point when T, > T,, Case 3 rejects low absorber heat. Thus to achieve low Q,, a series piping arrangement for cooling water in which water enters first the condenser then goes to the absorber can be beneficial, because this yields a higher absorber than condenser temperature. Also, it is clear that Case 1, in which the absorber and condenser are equal, rejects the lowest Q, at low T, when the heat source and the cooling water temperature are low.

Figures 2(I) and 2(J) give a comparison of Copt,, for the cases considered. This variation is the mirror reflection of Q, variation discussed earlier with reference to Figures 2(E) and 2(F). Case 2 yields higher COP values when T, < T,. Also when T, > T,, i.e., beyond the intersection point, Case 3 gives higher COP values. It is clear that cop,,, for Case 2 is higher than that for Case 3 at low cooling water temperatures when the heat source temperature is low. However, the highest Copt,, can be obtained if the absorber and the condenser temperatures are equal when evaporator, cooling water and heat source temperatures are low. This reveals that at low heat source and cooling water temperatures for high T,, one can expect that, if the absorber temperature is higher than the condenser temperature, high cop,,, can be achieved. Also when T, equals T,, COP,,, can be obtained at low evaporator, cooling water and heat source temperatures.

From these figures, it is seen that the change in cop,,, due to a given change of condenser temperature at constant absorber temperature is higher than that due to the same change in absorber temperature at constant condenser temperature. This confirms that Copt,, is more sensitive to variation in condenser temperature in HCF22-based VARS than to variation in absorber temperature.

Comparisons of second law efficiency ( E ) are given in Figures 2(K) and 2(L). Since E depends both on Q, and temperature levels, the variation of second law efficiency is different from that of cop,,,. At lower heat rejection temperatures, higher E is observed by Case 3 to the left of the intersection point, i.e., when T, < T,. However, at higher heat rejection temperature, i.e., beyond the intersection point, Case 3 for which T, > T, yields higher E . This means that cooling water direction may depend on whether the heat rejection temperature is low or high.

Similar trends for heat quantities, as well as performance characteristics, for HCFC22-DMF and HCFC22-DMETEG systems are observed. For instance, comparison of circulation ratio, coefficient of performance and second law efficiency are illustrated in Figure 3 for HCFC22-DMF and in Figure 4 for HCFC22-DMETEG.

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HCFC22-ABSORPTION SYSTEM I11 493

CONCLUSIONS

From the above study, the following conclusions are made:

(1) In HCFC22-absorbent VARS, COP is more sensitive to condenser than absorber temperature. (2) To achieve low heat quantities and high COP and 4, a series piping arrangement for cooling water in

which water first enters the condenser then goes to the absorber can be beneficial when the evaporator, heat source and cooling water temperatures are high, because this results in a higher absorber than condenser temperature.

(3) Case 1 in which T, and T, are equal, yields low CR, Qh, Qg and Q, as well as high COP,, and E for each working fluid for low evaporator temperature at low heat source and cooling water temperatures. This reveals that, at low evaporator, cooling water and heat source generator temperatures, a parallel pipe connection for cooling water which gives equal temperature at the absorber and the condenser may be recommended to improve the system performance.

NOMENCLATURE

COP = coefficient of performance CR = circulation ratio E = effectiveness h = specific enthalpy (kJ/kg) M = mass flowrate (kg/s) P = pressure (bar) Q T = temperature (K) X E = second law efficiency

= heat transfer rate (kW)

= weight fraction of R22 in solution (kg,/kg,)

Subscripts

1-10 a

cr e e9 g h r

th

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ss

ws

= state points in Figure 1 = absorber = condenser = Carnot = evaporator = equilibrium = generator = solution heat exchanger = refrigerant = strong solution = thermodynamic = weak solution

REFERENCES

Borde, I. and Jelinek, M. (1987). ‘Development of absorption refrigeration units for cold storage of agricultural products’, Int. I .

Borde, I., Jelinek, M. and Yaron, I. (1979). ‘Continuous solar heated absorption cooling unit for industrial applications’, Int. Solar

Dan, P. D. and Srinivasa Murthy, S. (1989). ‘Comparative thermodynamic study of fluorocarbon refrigerant based vapour absorption

Fatouh, M. and Srinivasa Murthy, S. (1993a). ‘Comparison of R22-absorbent pairs for absorption cooling based on P-T-X data’, 1.

Refn’g., 10, 53-56.

Fonun, Hambug, Vol. 11, pp. 511-521.

heat pumps’, Int. J. Energy Resemh, 13, 1-22.

Renewable Energy, 3,31-37.


Fatouh, M. and SMivasa Murthy, S. (1993b). ‘Comparison of R22-absorbent pairs for absorption cooling based on H-T-X data, J.

Fatouh, M. and SMivasa Murthy, S. (1995). HCFC22 based vapour absorption refrigeration systems, Part I: parametric studies, Int.

Herold, K. E. and Radermacher, R. (1991). ‘Development of an absorption heat pump water heater using aqueous ternary

Jelinek, M., Borde, I. and Yaron, I. (1978). ‘Enthalpy-concentration diagram of the system R22-DMF and performance

Kouremenos, D. A., Rogdakk, E. and Antonopoulos, K. A. (1989). A highefficiency, compound NH,/H,O-H,O/LiBr absorption

Kumar, S., Prevost, M. and Bugarel, R. (1991). ‘Comparison of various working pairs for absorption refrigeration systems.

Richter, K. H. and Schumacher, G. (1974). ‘Design operation and comparative evaluation of industrial ammonia-water absorption

Renewable E-,, 3,262-271.

J. Ene~gy Research (submitted).

hydroxide working fluid‘, Int, J. Ref~ig., 14,156-167.

characteristics of refrigeration cycle operating with this system’, ASHRAE Tmnr., 84, Pt. 11, 60-67.

refrigeration system’, E-, 14,893-905.

Application of R21 and R22 as refrigerants’, Int. J. Refrig., 14, 304-310.

plant’, Linde Reports on Science and Technology, 20, 26-55.

Documents

HCFC22-based absorption cooling systems, part III: Effects of different absorbers and condenser temperatures