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Performance of an advanced absorption cycle with R125 and different absorbents
Energy 29 (2004) 2501–2515 www.elsevier.com/locate/energy
Performance of an advanced absorption cycle with R125 and different absorbents A. Levy a,, M. Jelinek a, I. Borde a, F. Ziegler b a
b
Department of Mechanical Engineering, Ben-Gurion University of the Negev, P.O. Box 653, 84105 Beer Sheva, Israel Technische Universita¨t Berlin, Institut fu¨r Energietechnik, Ernst-Reuter-Platz 1, D-10587 Berlin, Germany
Abstract The performance of an advanced triple-pressure level (TPL) single-stage absorption cycle with refrigerant R125 and various organic absorbents were studied. In the developed TPL cycle, a jet ejector of a special design is added at the absorber inlet. The device serves two major functions: it facilitates pressure recovery and improves the mixing process between the weak solution and the refrigerant vapor coming from the evaporator. These effects enhance the absorption process of the refrigerant vapor into the solution drops. To facilitate the design of a jet ejector for absorption machines, a numerical model of simultaneous heat and mass transfer between the liquid and the gas phases in the ejector was developed. Based on the computerized simulation program, a parametric study of a TPL cycle was carried out. Comparison was made between the performances of the TPL and the common double pressure level (DPL) absorption cycle with refrigerant R125 and various organic absorbents. In addition, the influence of the jet ejector on the performance of the absorption cycle and the size of the unit was studied. # 2004 Published by Elsevier Ltd.
1. Introduction Various types of absorption heat pumps, both single-stage and multistage, can exploit a varv iety of heat sources for cooling and refrigeration (<0 C). The type of absorption heat pump and the nature of the working fluids are usually determined by the temperature of the heat source and the cooling or refrigeration demands. Utilization of low-potential heat sources (70– v 120 C) for cooling and refrigeration is limited by the properties of the working fluids and the cycle configuration of the heat pump. For utilization of low-potential heat sources for cooling
Corresponding author. Tel.: +972-7-6477092; fax: +972-7-6472813. E-mail address: [email protected] (A. Levy).
0360-5442/$ - see front matter # 2004 Published by Elsevier Ltd. doi:10.1016/j.energy.2004.03.045
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Nomenclature coefficient of performance ( Qe =ðQg þ Wp Þ) the pressure recovered by the ejector divided by the pressure difference between the nozzle inlet and outlet, in percentage f circulation ratio (kg strong solution/kg refrigerant) P pressure (Pa) heat rejected by the evaporator (J) Qe heat transferred to the generator (J) Qg heat transferred at the solution heat exchanger (J) Qhs Qhs/Qe the heat transferred at the solution heat exchanger divided by the heat rejected by the evaporator (J/J) v evaporator temperature ( C) Te v generator temperature ( C) Tg v cooling water temperature ( C) Tw
COP dPrec
v
and refrigeration to <0 C, a single-stage absorption heat pump based on organic working fluids is preferable due to the imposed limitations of the conventional working fluids such as ammonia–water or water–lithium bromide. The ammonia–water combination requires a heat v v source temperature above 120 C for cooling and refrigeration to <0 C. Such a system is a high-pressure system that requires a rectification column [1]. Ammonia has acceptable thermophysical properties, but it is a flammable, toxic, strongly irritant fluid, and is corrosive to copper. The water–lithium bromide solution can be used with a heat source temperature above v 70 C for air-conditioning but not for cooling and refrigeration because of the limitation for the v evaporator temperature (>0 C). This system operates under vacuum and does not require a rectification column. The water–lithium bromide solution is highly corrosive and more viscous than water. Thus corrosion inhibitors are required. The limitations of using common working v fluids [2] for utilizing low-potential heat sources (80–120 C) for cooling and refrigeration v (<0 C) are thus self-evident. With a view to overcoming these limitations, the working fluids used in the studies of our group are based on environmentally acceptable fluorocarbon (HFC) refrigerants and organic absorbents [3–12]. The refrigerants are not toxic or corrosive, and the organic working fluids are environmentally acceptable. In our high-pressure system, the condenser and the absorber are water cooled, and a rectification column is not needed, since the difference between the normal v boiling points of the absorbent and the refrigerant is about 200 C. The performance of absorption systems is expressed in terms of the coefficient of performance (COP) and the circulation ratio (f). The COP is defined as the heat supplied from the evaporator divided by the heat supplied to the generator and the energy supplied to the pump. Sometimes, the energy supplied to the pump is neglected in the denominator. The circulation ratio (f) is defined as the ratio between the mass flow rate of the strong solution in the pump to the mass flow rate of the refrigerant. In this paper in addition to COP and f a pump work ratio (PWR), which is the ratio of pump work to generator heat is given. In a conventional double-pressure
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level (DPL) single-stage absorption cycle with organic pairs the COP is about 0.5, f is in the range of 3–7 and the PWR is in the range of 1–5%. In the present study, an advanced single-stage absorption heat pump, i.e., a triple-pressurev level (TPL) cycle was used to utilize a low-potential heat source for cooling to 5 to 15 C. Similar absorption cycles have been suggested by Borde and Jelinek [9,10] and Chen [13], in which a jet ejector was used as a device for mixing and pressure recovery. Borde and Jelinek [9] conducted an experimental study on a heat pump. They showed the suitability to integrate a self-made jet ejector as a mixing unit. Borde and Jelinek [10] conducted a theoretical study on a heat pump with organic working fluids. In their mathematical model, the pressure recovery was not taken into account. Thus the jet ejector was considered only as a mixing device. Chen [13] used a Bernoulli type equation for analysis of the two-phase mixture flow through the ejector. This analysis was based on a single-phase model of pseudo-fluid without taking heat transfer, mass transfer and momentum balance into account. To overcome the above mentioned simplifications, theoretical studies on the performance of the jet ejector parameters were conducted [14–16]. To facilitate the design of the jet ejector for absorption machines, a two-phase numerical model of simultaneous mass, momentum and heat transfer in the jet ejector was developed [11,12]. In these studies, a combination of computerized simulation programs, one for the conventional single-stage absorption cycle (DPL) and the other for the jet ejector, was developed to examine the influence of the jet ejector on the performance of the TPL single-stage absorption cycle. 2. TPL single-stage absorption cycle v
To utilize a low-potential heat source for cooling to 5 to 15 C, an advanced single-stage absorption heat pump can be used, i.e., a TPL cycle. This cycle may be implemented in a number of different ways, for example, by a design known as the compression/absorption cycle [17– 20], in which a compressor is inserted between the evaporator and the absorber. The main disadvantage of the latter device is the extra electrical energy required for the compressor. In our TPL single-stage heat pump, a specially designed jet ejector is introduced at the absorber inlet (Fig. 1), with the purpose of increasing the absorber pressure relative to the evaporator pressure (by partially recovering the high-pressure of the generator), improving the mixing process and the pre-absorption by the weak solution of the refrigerant coming from the evaporator [14–16]. The mixing process in the jet ejector is very intense as a result of spray generation of the liquid phase and of extensive subcooling of the weak solution in the solution heat exchanger. A schematic representation of jet ejector is given in Fig. 2 and the processes of the solution sub-cycle between the generator and the absorber are given in Fig. 3 for both the DPL and the TPL cycles. In the last figure, solid and dashed lines are representing equilibrium and non-equilibrium processes, respectively. The line N–M represents the pressure drop process due to expansion process at the DPL cycle. In the TPL cycle, the line N–M represents energy conversion (potential to kinetic energy) at the nozzle. The point M represents the start of the mixing process between the poor solution and the refrigerants coming from the evaporator. The line M–D represents the pressure recovery process at the diffuser. As can be seen at the pressure–concentration diagram (Fig. 3a), the main difference between the solution sub-cycles is illustrated by the larger concentration differences between the absorber and the generator outlets in the TPL
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Fig. 1. Schematic representation of a TPL single-stage advanced absorption cycle (Sol. H.E., solution heat exchanger; Ref. H.E., refrigerant heat exchanger; P, solution pump).
cycle in comparison with the DPL cycle. Due to the concentration difference between the absorbers’ outlets (nhTPL nhDPL ), vapor generation at the TPL cycle will start at lower temperature at high-pressure (points STPL and SDPL at Fig. 3b). 2.1. Jet ejector The jet ejector consists basically of a nozzle and a diffuser. The device is characterized by the facts that there are no moving parts and there is no requirement for an additional energy source. At the nozzle outlet, production of a full cone spray of small high-velocity drops of a
Fig. 2. Schematic representation of the jet ejector.
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Fig. 3. Process diagram of the solution sub-cycles for both the DPL and the TPL (A, absorber outlet; D, diffuser outlet; G, generator outlet; P, pump outlet; S, first equilibrium point of the strong solution at high-pressure; N, nozzle inlet; M, nozzle outlet/diffuser inlet).
weak solution serves to transform potential energy (pressure difference between the generator and the evaporator) into kinetic energy. At the diffuser inlet (low pressure of the evaporator), the stream of drops mixes with the refrigerant vapor coming from the evaporator. In the diffuser, pressure recovery (defined as the difference of pressure between the outlet and the inlet of the diffuser) is thus obtained by transforming the droplet kinetic energy to potential energy. As a result, the velocities of both the solution drops and the refrigerant gas are reduced. To facilitate the design of the jet ejector for absorption machines, a numerical model of simultaneous mass, momentum and heat transfer between the liquid and gas phases in the jet ejector was developed. The model was based on the following assumptions: steady state; adiabatic flow; dilute phase flow; mass, momentum and heat transfer only between the two phases; the effect of vapor transport on drag and droplet external diffusion resistance were neglected; friction forces with the diffuser wall negligible; droplets comprise a binary mixture of absorbent and refrigerant; constant droplet diameter; electrical forces and surface tension neglected, temperature and composition gradients in the droplets were taken into account; the dispersed phase is a binary mixture, and hence all its thermodynamic properties were calculated upon the basis of weight fraction of the refrigerant in the solution, n, and its other thermophysical properties were obtained by applying the mixture theory. The numerical solution was obtained by means of Gear’s fifth-order BDF method (often called Gear’s stiff method.), which is available in the IMSL library [21]. The influence of the jet ejector design parameters on the pressure recovery, temperature and concentration of the refrigerant in the solution, and velocities of the gas and liquid drops, has previously been described by Levy et al. [22]. The parametric study involved examination and ways of augmentation of the mass transfer process in the diffuser with the ultimate aim of designing a compact and efficient unit. The numerical simulation demonstrated that pressure recovery and pre-absorption in the jet ejector would improve the efficiency of the absorption process and hence the overall efficiency of the refrigeration cycle.
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2.2. Application of the TPL absorption cycle In commonly used absorption heat pumps (when the pressure drop is negligible), the absorber pressure is equal to the evaporator pressure. However, in the TPL cycle, the absorber pressure is higher than evaporator pressure, which leads to improvement of the cycle performance. This improvement can be obtained by decreasing the circulation ratio f (Case 1) or by changing the governing cycle temperatures as follows: lowering the refrigeration temperature, i.e., the evaporator temperature Te (Case 2); lowering the heat source temperature, i.e., the generator temperature Tg (Case 3); or raising the cooling water temperature, i.e., the condenser and absorber temperature Tw (Case 4), as summarized in Table 1. Implementation of the four cases was explained and analyzed by Jelinek et al. [12].
3. System analysis and discussion A combination of the computerized simulation programs, one for the conventional singlestage double-pressure level (DPL) absorption cycle and the other for the jet ejector, facilitated evaluation of the influence of the jet ejector on the performance of the TPL single-stage absorption cycle. In all cases, calculation of the performances (COP, f, Qhs/Qe, which is defined as heat transferred by the solution heat exchanger in relation to the cooling capacity, and dPrec, which is defined as the pressure recovered by the ejector related to the pressure difference between nozzle inlet and outlet) of the TPL and DPL cycles was carried out for the refrigerant R125 with the following absorbents: Dimethylether of tetraethyleneglycol (DMETEG), Nmethyl-2-pyrrolidone (NMP), N-methyl-e-caprolactam (MCL), N-N0 dimethylacetamide (DMAC), Dimethylethylenurea (DMEU) and Dimethylpropylenurea (DMPU). 3.1. Comparison between the TPL absorption cycle and the DPL absorption cycle The calculations for the DPL absorption cycle (absorber pressure ¼ evaporator pressure) and the TPL absorption cycle (absorber pressure ¼ evaporator pressure plus the recovered pressure) were carried out under the following operating conditions: cooling water temperature v v v 25 C (condenser temperature 32 C and absorber temperature 28 C), generator temperature in v v the range of 70–140 C, and evaporator temperatures of 5 C. For both cycles, the pressure Table 1 Comparison between the possibilities provided by the TPL vs. the common single-stage (DPL) absorption cycle in terms of the outside temperatures and the circulation ratio v
Case 1 Case 2 Case 3 Case 4 a
v
v
Tg ( C)
Te ( C)
Tw ( C)
a
a
a
Lower
a
Lower
a
a
Lower
a
a
a
a
a
Higher
a
Equal to the value of the common single-stage absorption cycle.
f
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drop along the cycle was taken as zero and the outlet temperature of the poor solution from the v solution heat exchanger is 5 C above the absorber outlet temperature. For the same governing temperatures of both the DPL and TPL cycles, the higher absorber pressure (by the amount of the pressure recovery) for the TPL cycle and the same absorber outlet temperature (line D–ATPL) results in a higher weight fraction of the refrigerant in the solution (point ATPL) as can be seen in Fig. 3. This increase in the weight fraction at the absorber outlet leads to a reduction in f, i.e., a reduction in the mass flow rate of the strong solution in the pump, and consequently to a reduction in the amount of heat to be transferred at the solution heat exchanger. The calculated COP, f and Qhs/Qe for the same evaporator temperature in TPL and DPL absorption cycles with the following working fluids R125-DMETEG, R125NMP, R125-MCL, R125-DMAC, R125-DMEU and R125-DMPU are shown in Figs. 4–8. The calculated pressure recovery in the ejector, as expressed by the term dPrec (in %), increases as the generator temperature decreases. Increases of dPrec up to 8% were obtained, as v can be seen in Fig. 4. For example, at generator temperature of 90 C, the values of dPrec for the investigated working fluids are in the range of 5–6.5%. The values of f in the TPL cycle are lower than those of the DPL cycle (Fig. 5). At the same operating conditions, the reduction in the values of f can reach 40%. For example, at generator v temperature of 90 C, the values of f in the DPL cycle for the investigated working fluids are in the range of 4–5.8 while in the TPL cycle the values are in the range of 3.2–4.2. The reduction in f is in the range of 20–28%. The values of PWR in the TPL cycle are lower than those in the DPL cycle as shown in Fig. 6. At high generator temperature the values are similar, but as the generator temperature v decreases the differences are increased. For example, at generator temperature of 90 C, the
Fig. 4. Variation of the calculated pressure recovery (%) in the jet ejector with generator temperature for evaporator v temperature of 5 C as calculated for various absorbents.
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Fig. 5. Comparison between the calculated f in TPL absorption cycle (a) and in DPL absorption cycle (b) as a funcv tion of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents.
values of PWR in the DPL cycle for the investigated working fluids are in the range of 1.7–2.6% while in the TPL cycle the values are in the range of 1.3–1.8%. The reduction in PWR is in the range of 24–31%. The lower values of PWR in the TPL cycle are due to the lower f at the TPL cycle.
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Fig. 6. Comparison between the calculated PWR in TPL absorption cycle (a) and in DPL absorption cycle (b) as a v function of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents.
The relative size of the solution heat exchanger can be expressed in terms of Qhs/Qe. The values of Qhs/Qe in the TPL cycle are lower than those in the DPL cycle (Fig. 7). At the same operating conditions, the reduction in Qhs/Qe can reach 40% of the Qhs/Qe in the DPL cycle.
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Fig. 7. Comparison between the calculated Qhs/Qe in TPL absorption cycle (a) and in DPL absorption cycle (b) as a v function of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents. v
For example, at generator temperature of 90 C, the values of Qhs/Qe in the DPL cycle for the investigated working fluids are in the range of 2.8–5.2 while in the TPL cycle the values are in the range of 2–3.5. The reduction in Qhs/Qe is in the range of 29–33%. Fig. 8 present a comparison between the calculated COP in TPL absorption cycle (Fig. 8a) and in DPL absorption cycle (Fig. 8b) as a function of the generator temperature for evaporv ator temperature of 5 C as calculated for various absorbents. As can be seen, higher values
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Fig. 8. Comparison between the calculated COP in TPL absorption cycle (a) in DPL absorption cycle (b) and the difv ference in COP in percent (c) as a function of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents.
of the COP were obtained by the TPL cycle in comparison with those obtained by the DPL cycle. It can also be seen that these higher COP were achieved at lower generator temperatures. The optimal operating conditions, for a constant evaporator temperature, are those that achieve the highest COP. The values of the generator temperature and f for these maximum values of the COP are given in Table 2 for the mentioned working fluids. Table 2 illustrates the influence of pressure recovery on the performance of the TPL cycle in comparison with the DPL cycle. At the same evaporator temperatures, same cooling water temperature and the highest COP, the main benefit of the TPL cycle is evident in the reduction of the generator temperav ture by 9–12 C, the increasing of the maximum COP’s by 5–6% while the circulation ratios remained almost the same ( 3%). Since the LMTD of the solution heat exchanger becomes lower and the circulation ratios remained practically constant, the size of the solution heat exchanger decreases up to 19%.
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Table 2 Tg, f and Qhs/Qe at maximum calculated COP for various absorbents with R125 in both cycles and the differences between them DPL
DMETEG NMP MCL DMAC DMEU DMPU
TPL
(TPL–DPL)
COP
Tg v ( C)
f
Qhs/Qe
COP
Tg
f
Qhs/Qe d(COP) %
dTg v ( C)
d(f) %
d(Qhs/Qe) %
0.46 0.54 0.53 0.54 0.56 0.52
102 96 102 102 100 101
4.0 3.3 3.9 3.3 4.2 3.8
4.3 2.4 3.6 2.9 3.0 2.9
0.49 0.57 0.56 0.57 0.59 0.54
90 87 92 93 88 92
3.9 3.4 3.9 3.4 4.2 3.8
3.5 2.2 3.1 2.6 2.5 2.5
12 9 10 9 12 9
2.5 3.0 0 3.0 0 0
18.6 8.3 13.9 10.3 16.7 13.8
6.1 5.0 5.9 5.0 6.1 5.4
The results of this analysis for the various working fluids in both cycles while the generator v temperature is 100 C are shown in Table 3. It can be seen from Table 3 that the COP’s were increased in the range of 3–5%, the circulation ratios decreases in the range of 13–21% and the Qhs/Qe’s decreased in the range of 21–35%. It should be pointed out that since the circulation ratio decreases while the cycle temperatures remained constant, a reduction in the size of the solution heat exchanger was obtained. 3.2. Comparison between the working fluids performances The best working fluid can be considered as the one with the highest COP while the circulation ratio f, the generator temperature and the value of the heat transferred by the solution heat exchanger divided by the heat rejected by the evaporator, Qhs/Qe, are as low as possible. Lower circulation ratio f and lower value of Qhs/Qe leads to a smaller size of the absorption machine. As can be seen in Table 2, the maximum COP in the TPL cycle was obtained by the solution R125-DMEU (0.59) followed by R125-NMP (0.57), R125-DMAC (0.57), R125-MCL (0.56), R125-DMPU (0.54) and R125-DMETEG (0.49). It should be pointed out that although R125v DMEU has the highest COP at low generator temperature (88 C), its circulation ratio is the Table 3 v COP, f and Qhs/Qe for various absorbents with R125 in both cycles at generator temperature of 100 C and evaporv ator temperature of 5 C DPL
DMETEG NMP MCL DMAC DMEU DMPU
TPL
(TPL–DPL)
COP
f
Qhs/Qe
COP
f
Qhs/Qe
d(COP) %
d(f) %
d(Qhs/Qe) %
0.46 0.54 0.53 0.54 0.56 0.52
4.2 3.0 4.1 3.4 4.2 3.9
4.4 2.6 3.7 3.0 3.0 2.9
0.49 0.56 0.55 0.56 0.59 0.54
3.35 2.6 3.35 2.9 3.3 3.25
3.3 1.7 2.8 2.35 2.2 2.3
2.9 3.3 4.8 4.1 4.8 4.9
20.2 13.3 18.3 14.7 21.4 16.7
25.0 34.6 24.3 21.7 26.7 20.7
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highest (4.2) and the Qhs/Qe value is moderate in comparison with the other solutions. Therefore, the second and third solutions have to be taken into consideration. Since the COP, the circulation ratio and the value of Qhs/Qe of R125-NMP and R125-DMAC are almost the same, R125-NMP is preferable among them because it can be operated under lower generator temv v perature, 87 C, compared to 93 C with R125-DMAC. This order was obtained also with the DPL cycle. When operating conditions of the cycle were known and the generator temperature was taken v as 100 C (see Table 3), the highest COP was achieved in the following order: R125-DMEU (0.59) followed by R125-DMAC (0.56), R125-NMP (0.56), R125-MCL (0.55), R125-DMPU (0.54) and R125-DMETEG (0.49). It should be pointed out that although R125-DMEU has the highest COP, its circulation ratio and the value of Qhs/Qe are moderate in comparison with the other solutions. Therefore, the second and third solutions have to be taken into consideration. Although, the COP of R125-DMAC is a little bit higher than the COP of R125-NMP, the later is preferable because its lower f (2.6 compared to 2.9 with R125-DMAC) and the value of Qhs/ Qe is lower (1.7 compared to 2.35 with R125-DMAC).
4. Conclusions Integrating a specially designed jet ejector at the absorber inlet resulted in an advanced TPL single-stage absorption cycle. The jet ejector facilitated pressure recovery and improved the mixing between the weak solution and the refrigerant vapor coming from the evaporator. The influence of the jet ejector on the performance of the TPL absorption cycle was evaluated by comparing the performance of the TPL absorption cycle with that of the DPL cycle under the same operating conditions. The analysis was made with six environmentally acceptable working fluids; R125-DMETEG, R125-NMP, R125-MCL, R125-DMAC, R125-DMEU and R125DMPU. Among the four options of improvements in the TPL absorption cycle performances [12], the ability to reduce the circulation ratio f was investigated. From this study, the improvement in the performance of the TPL cycle in comparison to the DPL cycle with the mentioned working fluids is evident. The TPL cycle can be operated by v lower heat source (9–12 C) and yet with higher COP’s of 5–6% and required smaller solution heat exchanger up to 19% (see Table 2). However, for the same operation condition, the TPL cycle provided higher COP’s of 3–5%, lower circulation ratios of 13–21% and required smaller solution heat exchanger of 22–35% (see Table 3). It should be noted that these improvements are functions of the operating conditions and the thermophysical properties of the working fluids. The interdependencies between the different parameters are quite complex. It is obvious, that at the end of the day cost functions for the different apparatuses have to be incorporated into the optimisation, so that an economic optimum can be found. Concerning the working fluids based on the refrigerant R125 and the six absorbents, it was found that the solution R125-DMEU showed the best performances followed by R125-NMP, R125-DMAC, R125-MCL, R125-DMPU and R125-DMETEG. The calculated COP’s in the TPL cycle were in the range of 0.45–0.59 and the circulation ratios were in the range of 2–5.
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Acknowledgements This work was partially supported by the German–Israeli Scientific Cooperation Unit, Ministry of Science, under the program BMBF-MOS, Grant-GR-1718.
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Performance of an advanced absorption cycle with R125 and different absorbents A. Levy a,, M. Jelinek a, I. Borde a, F. Ziegler b a
b
Department of Mechanical Engineering, Ben-Gurion University of the Negev, P.O. Box 653, 84105 Beer Sheva, Israel Technische Universita¨t Berlin, Institut fu¨r Energietechnik, Ernst-Reuter-Platz 1, D-10587 Berlin, Germany
Abstract The performance of an advanced triple-pressure level (TPL) single-stage absorption cycle with refrigerant R125 and various organic absorbents were studied. In the developed TPL cycle, a jet ejector of a special design is added at the absorber inlet. The device serves two major functions: it facilitates pressure recovery and improves the mixing process between the weak solution and the refrigerant vapor coming from the evaporator. These effects enhance the absorption process of the refrigerant vapor into the solution drops. To facilitate the design of a jet ejector for absorption machines, a numerical model of simultaneous heat and mass transfer between the liquid and the gas phases in the ejector was developed. Based on the computerized simulation program, a parametric study of a TPL cycle was carried out. Comparison was made between the performances of the TPL and the common double pressure level (DPL) absorption cycle with refrigerant R125 and various organic absorbents. In addition, the influence of the jet ejector on the performance of the absorption cycle and the size of the unit was studied. # 2004 Published by Elsevier Ltd.
1. Introduction Various types of absorption heat pumps, both single-stage and multistage, can exploit a varv iety of heat sources for cooling and refrigeration (<0 C). The type of absorption heat pump and the nature of the working fluids are usually determined by the temperature of the heat source and the cooling or refrigeration demands. Utilization of low-potential heat sources (70– v 120 C) for cooling and refrigeration is limited by the properties of the working fluids and the cycle configuration of the heat pump. For utilization of low-potential heat sources for cooling
Corresponding author. Tel.: +972-7-6477092; fax: +972-7-6472813. E-mail address: [email protected] (A. Levy).
0360-5442/$ - see front matter # 2004 Published by Elsevier Ltd. doi:10.1016/j.energy.2004.03.045
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Nomenclature coefficient of performance ( Qe =ðQg þ Wp Þ) the pressure recovered by the ejector divided by the pressure difference between the nozzle inlet and outlet, in percentage f circulation ratio (kg strong solution/kg refrigerant) P pressure (Pa) heat rejected by the evaporator (J) Qe heat transferred to the generator (J) Qg heat transferred at the solution heat exchanger (J) Qhs Qhs/Qe the heat transferred at the solution heat exchanger divided by the heat rejected by the evaporator (J/J) v evaporator temperature ( C) Te v generator temperature ( C) Tg v cooling water temperature ( C) Tw
COP dPrec
v
and refrigeration to <0 C, a single-stage absorption heat pump based on organic working fluids is preferable due to the imposed limitations of the conventional working fluids such as ammonia–water or water–lithium bromide. The ammonia–water combination requires a heat v v source temperature above 120 C for cooling and refrigeration to <0 C. Such a system is a high-pressure system that requires a rectification column [1]. Ammonia has acceptable thermophysical properties, but it is a flammable, toxic, strongly irritant fluid, and is corrosive to copper. The water–lithium bromide solution can be used with a heat source temperature above v 70 C for air-conditioning but not for cooling and refrigeration because of the limitation for the v evaporator temperature (>0 C). This system operates under vacuum and does not require a rectification column. The water–lithium bromide solution is highly corrosive and more viscous than water. Thus corrosion inhibitors are required. The limitations of using common working v fluids [2] for utilizing low-potential heat sources (80–120 C) for cooling and refrigeration v (<0 C) are thus self-evident. With a view to overcoming these limitations, the working fluids used in the studies of our group are based on environmentally acceptable fluorocarbon (HFC) refrigerants and organic absorbents [3–12]. The refrigerants are not toxic or corrosive, and the organic working fluids are environmentally acceptable. In our high-pressure system, the condenser and the absorber are water cooled, and a rectification column is not needed, since the difference between the normal v boiling points of the absorbent and the refrigerant is about 200 C. The performance of absorption systems is expressed in terms of the coefficient of performance (COP) and the circulation ratio (f). The COP is defined as the heat supplied from the evaporator divided by the heat supplied to the generator and the energy supplied to the pump. Sometimes, the energy supplied to the pump is neglected in the denominator. The circulation ratio (f) is defined as the ratio between the mass flow rate of the strong solution in the pump to the mass flow rate of the refrigerant. In this paper in addition to COP and f a pump work ratio (PWR), which is the ratio of pump work to generator heat is given. In a conventional double-pressure
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level (DPL) single-stage absorption cycle with organic pairs the COP is about 0.5, f is in the range of 3–7 and the PWR is in the range of 1–5%. In the present study, an advanced single-stage absorption heat pump, i.e., a triple-pressurev level (TPL) cycle was used to utilize a low-potential heat source for cooling to 5 to 15 C. Similar absorption cycles have been suggested by Borde and Jelinek [9,10] and Chen [13], in which a jet ejector was used as a device for mixing and pressure recovery. Borde and Jelinek [9] conducted an experimental study on a heat pump. They showed the suitability to integrate a self-made jet ejector as a mixing unit. Borde and Jelinek [10] conducted a theoretical study on a heat pump with organic working fluids. In their mathematical model, the pressure recovery was not taken into account. Thus the jet ejector was considered only as a mixing device. Chen [13] used a Bernoulli type equation for analysis of the two-phase mixture flow through the ejector. This analysis was based on a single-phase model of pseudo-fluid without taking heat transfer, mass transfer and momentum balance into account. To overcome the above mentioned simplifications, theoretical studies on the performance of the jet ejector parameters were conducted [14–16]. To facilitate the design of the jet ejector for absorption machines, a two-phase numerical model of simultaneous mass, momentum and heat transfer in the jet ejector was developed [11,12]. In these studies, a combination of computerized simulation programs, one for the conventional single-stage absorption cycle (DPL) and the other for the jet ejector, was developed to examine the influence of the jet ejector on the performance of the TPL single-stage absorption cycle. 2. TPL single-stage absorption cycle v
To utilize a low-potential heat source for cooling to 5 to 15 C, an advanced single-stage absorption heat pump can be used, i.e., a TPL cycle. This cycle may be implemented in a number of different ways, for example, by a design known as the compression/absorption cycle [17– 20], in which a compressor is inserted between the evaporator and the absorber. The main disadvantage of the latter device is the extra electrical energy required for the compressor. In our TPL single-stage heat pump, a specially designed jet ejector is introduced at the absorber inlet (Fig. 1), with the purpose of increasing the absorber pressure relative to the evaporator pressure (by partially recovering the high-pressure of the generator), improving the mixing process and the pre-absorption by the weak solution of the refrigerant coming from the evaporator [14–16]. The mixing process in the jet ejector is very intense as a result of spray generation of the liquid phase and of extensive subcooling of the weak solution in the solution heat exchanger. A schematic representation of jet ejector is given in Fig. 2 and the processes of the solution sub-cycle between the generator and the absorber are given in Fig. 3 for both the DPL and the TPL cycles. In the last figure, solid and dashed lines are representing equilibrium and non-equilibrium processes, respectively. The line N–M represents the pressure drop process due to expansion process at the DPL cycle. In the TPL cycle, the line N–M represents energy conversion (potential to kinetic energy) at the nozzle. The point M represents the start of the mixing process between the poor solution and the refrigerants coming from the evaporator. The line M–D represents the pressure recovery process at the diffuser. As can be seen at the pressure–concentration diagram (Fig. 3a), the main difference between the solution sub-cycles is illustrated by the larger concentration differences between the absorber and the generator outlets in the TPL
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Fig. 1. Schematic representation of a TPL single-stage advanced absorption cycle (Sol. H.E., solution heat exchanger; Ref. H.E., refrigerant heat exchanger; P, solution pump).
cycle in comparison with the DPL cycle. Due to the concentration difference between the absorbers’ outlets (nhTPL nhDPL ), vapor generation at the TPL cycle will start at lower temperature at high-pressure (points STPL and SDPL at Fig. 3b). 2.1. Jet ejector The jet ejector consists basically of a nozzle and a diffuser. The device is characterized by the facts that there are no moving parts and there is no requirement for an additional energy source. At the nozzle outlet, production of a full cone spray of small high-velocity drops of a
Fig. 2. Schematic representation of the jet ejector.
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Fig. 3. Process diagram of the solution sub-cycles for both the DPL and the TPL (A, absorber outlet; D, diffuser outlet; G, generator outlet; P, pump outlet; S, first equilibrium point of the strong solution at high-pressure; N, nozzle inlet; M, nozzle outlet/diffuser inlet).
weak solution serves to transform potential energy (pressure difference between the generator and the evaporator) into kinetic energy. At the diffuser inlet (low pressure of the evaporator), the stream of drops mixes with the refrigerant vapor coming from the evaporator. In the diffuser, pressure recovery (defined as the difference of pressure between the outlet and the inlet of the diffuser) is thus obtained by transforming the droplet kinetic energy to potential energy. As a result, the velocities of both the solution drops and the refrigerant gas are reduced. To facilitate the design of the jet ejector for absorption machines, a numerical model of simultaneous mass, momentum and heat transfer between the liquid and gas phases in the jet ejector was developed. The model was based on the following assumptions: steady state; adiabatic flow; dilute phase flow; mass, momentum and heat transfer only between the two phases; the effect of vapor transport on drag and droplet external diffusion resistance were neglected; friction forces with the diffuser wall negligible; droplets comprise a binary mixture of absorbent and refrigerant; constant droplet diameter; electrical forces and surface tension neglected, temperature and composition gradients in the droplets were taken into account; the dispersed phase is a binary mixture, and hence all its thermodynamic properties were calculated upon the basis of weight fraction of the refrigerant in the solution, n, and its other thermophysical properties were obtained by applying the mixture theory. The numerical solution was obtained by means of Gear’s fifth-order BDF method (often called Gear’s stiff method.), which is available in the IMSL library [21]. The influence of the jet ejector design parameters on the pressure recovery, temperature and concentration of the refrigerant in the solution, and velocities of the gas and liquid drops, has previously been described by Levy et al. [22]. The parametric study involved examination and ways of augmentation of the mass transfer process in the diffuser with the ultimate aim of designing a compact and efficient unit. The numerical simulation demonstrated that pressure recovery and pre-absorption in the jet ejector would improve the efficiency of the absorption process and hence the overall efficiency of the refrigeration cycle.
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2.2. Application of the TPL absorption cycle In commonly used absorption heat pumps (when the pressure drop is negligible), the absorber pressure is equal to the evaporator pressure. However, in the TPL cycle, the absorber pressure is higher than evaporator pressure, which leads to improvement of the cycle performance. This improvement can be obtained by decreasing the circulation ratio f (Case 1) or by changing the governing cycle temperatures as follows: lowering the refrigeration temperature, i.e., the evaporator temperature Te (Case 2); lowering the heat source temperature, i.e., the generator temperature Tg (Case 3); or raising the cooling water temperature, i.e., the condenser and absorber temperature Tw (Case 4), as summarized in Table 1. Implementation of the four cases was explained and analyzed by Jelinek et al. [12].
3. System analysis and discussion A combination of the computerized simulation programs, one for the conventional singlestage double-pressure level (DPL) absorption cycle and the other for the jet ejector, facilitated evaluation of the influence of the jet ejector on the performance of the TPL single-stage absorption cycle. In all cases, calculation of the performances (COP, f, Qhs/Qe, which is defined as heat transferred by the solution heat exchanger in relation to the cooling capacity, and dPrec, which is defined as the pressure recovered by the ejector related to the pressure difference between nozzle inlet and outlet) of the TPL and DPL cycles was carried out for the refrigerant R125 with the following absorbents: Dimethylether of tetraethyleneglycol (DMETEG), Nmethyl-2-pyrrolidone (NMP), N-methyl-e-caprolactam (MCL), N-N0 dimethylacetamide (DMAC), Dimethylethylenurea (DMEU) and Dimethylpropylenurea (DMPU). 3.1. Comparison between the TPL absorption cycle and the DPL absorption cycle The calculations for the DPL absorption cycle (absorber pressure ¼ evaporator pressure) and the TPL absorption cycle (absorber pressure ¼ evaporator pressure plus the recovered pressure) were carried out under the following operating conditions: cooling water temperature v v v 25 C (condenser temperature 32 C and absorber temperature 28 C), generator temperature in v v the range of 70–140 C, and evaporator temperatures of 5 C. For both cycles, the pressure Table 1 Comparison between the possibilities provided by the TPL vs. the common single-stage (DPL) absorption cycle in terms of the outside temperatures and the circulation ratio v
Case 1 Case 2 Case 3 Case 4 a
v
v
Tg ( C)
Te ( C)
Tw ( C)
a
a
a
Lower
a
Lower
a
a
Lower
a
a
a
a
a
Higher
a
Equal to the value of the common single-stage absorption cycle.
f
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drop along the cycle was taken as zero and the outlet temperature of the poor solution from the v solution heat exchanger is 5 C above the absorber outlet temperature. For the same governing temperatures of both the DPL and TPL cycles, the higher absorber pressure (by the amount of the pressure recovery) for the TPL cycle and the same absorber outlet temperature (line D–ATPL) results in a higher weight fraction of the refrigerant in the solution (point ATPL) as can be seen in Fig. 3. This increase in the weight fraction at the absorber outlet leads to a reduction in f, i.e., a reduction in the mass flow rate of the strong solution in the pump, and consequently to a reduction in the amount of heat to be transferred at the solution heat exchanger. The calculated COP, f and Qhs/Qe for the same evaporator temperature in TPL and DPL absorption cycles with the following working fluids R125-DMETEG, R125NMP, R125-MCL, R125-DMAC, R125-DMEU and R125-DMPU are shown in Figs. 4–8. The calculated pressure recovery in the ejector, as expressed by the term dPrec (in %), increases as the generator temperature decreases. Increases of dPrec up to 8% were obtained, as v can be seen in Fig. 4. For example, at generator temperature of 90 C, the values of dPrec for the investigated working fluids are in the range of 5–6.5%. The values of f in the TPL cycle are lower than those of the DPL cycle (Fig. 5). At the same operating conditions, the reduction in the values of f can reach 40%. For example, at generator v temperature of 90 C, the values of f in the DPL cycle for the investigated working fluids are in the range of 4–5.8 while in the TPL cycle the values are in the range of 3.2–4.2. The reduction in f is in the range of 20–28%. The values of PWR in the TPL cycle are lower than those in the DPL cycle as shown in Fig. 6. At high generator temperature the values are similar, but as the generator temperature v decreases the differences are increased. For example, at generator temperature of 90 C, the
Fig. 4. Variation of the calculated pressure recovery (%) in the jet ejector with generator temperature for evaporator v temperature of 5 C as calculated for various absorbents.
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Fig. 5. Comparison between the calculated f in TPL absorption cycle (a) and in DPL absorption cycle (b) as a funcv tion of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents.
values of PWR in the DPL cycle for the investigated working fluids are in the range of 1.7–2.6% while in the TPL cycle the values are in the range of 1.3–1.8%. The reduction in PWR is in the range of 24–31%. The lower values of PWR in the TPL cycle are due to the lower f at the TPL cycle.
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Fig. 6. Comparison between the calculated PWR in TPL absorption cycle (a) and in DPL absorption cycle (b) as a v function of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents.
The relative size of the solution heat exchanger can be expressed in terms of Qhs/Qe. The values of Qhs/Qe in the TPL cycle are lower than those in the DPL cycle (Fig. 7). At the same operating conditions, the reduction in Qhs/Qe can reach 40% of the Qhs/Qe in the DPL cycle.
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Fig. 7. Comparison between the calculated Qhs/Qe in TPL absorption cycle (a) and in DPL absorption cycle (b) as a v function of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents. v
For example, at generator temperature of 90 C, the values of Qhs/Qe in the DPL cycle for the investigated working fluids are in the range of 2.8–5.2 while in the TPL cycle the values are in the range of 2–3.5. The reduction in Qhs/Qe is in the range of 29–33%. Fig. 8 present a comparison between the calculated COP in TPL absorption cycle (Fig. 8a) and in DPL absorption cycle (Fig. 8b) as a function of the generator temperature for evaporv ator temperature of 5 C as calculated for various absorbents. As can be seen, higher values
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Fig. 8. Comparison between the calculated COP in TPL absorption cycle (a) in DPL absorption cycle (b) and the difv ference in COP in percent (c) as a function of the generator temperature for evaporator temperature of 5 C as calculated for various absorbents.
of the COP were obtained by the TPL cycle in comparison with those obtained by the DPL cycle. It can also be seen that these higher COP were achieved at lower generator temperatures. The optimal operating conditions, for a constant evaporator temperature, are those that achieve the highest COP. The values of the generator temperature and f for these maximum values of the COP are given in Table 2 for the mentioned working fluids. Table 2 illustrates the influence of pressure recovery on the performance of the TPL cycle in comparison with the DPL cycle. At the same evaporator temperatures, same cooling water temperature and the highest COP, the main benefit of the TPL cycle is evident in the reduction of the generator temperav ture by 9–12 C, the increasing of the maximum COP’s by 5–6% while the circulation ratios remained almost the same ( 3%). Since the LMTD of the solution heat exchanger becomes lower and the circulation ratios remained practically constant, the size of the solution heat exchanger decreases up to 19%.
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Table 2 Tg, f and Qhs/Qe at maximum calculated COP for various absorbents with R125 in both cycles and the differences between them DPL
DMETEG NMP MCL DMAC DMEU DMPU
TPL
(TPL–DPL)
COP
Tg v ( C)
f
Qhs/Qe
COP
Tg
f
Qhs/Qe d(COP) %
dTg v ( C)
d(f) %
d(Qhs/Qe) %
0.46 0.54 0.53 0.54 0.56 0.52
102 96 102 102 100 101
4.0 3.3 3.9 3.3 4.2 3.8
4.3 2.4 3.6 2.9 3.0 2.9
0.49 0.57 0.56 0.57 0.59 0.54
90 87 92 93 88 92
3.9 3.4 3.9 3.4 4.2 3.8
3.5 2.2 3.1 2.6 2.5 2.5
12 9 10 9 12 9
2.5 3.0 0 3.0 0 0
18.6 8.3 13.9 10.3 16.7 13.8
6.1 5.0 5.9 5.0 6.1 5.4
The results of this analysis for the various working fluids in both cycles while the generator v temperature is 100 C are shown in Table 3. It can be seen from Table 3 that the COP’s were increased in the range of 3–5%, the circulation ratios decreases in the range of 13–21% and the Qhs/Qe’s decreased in the range of 21–35%. It should be pointed out that since the circulation ratio decreases while the cycle temperatures remained constant, a reduction in the size of the solution heat exchanger was obtained. 3.2. Comparison between the working fluids performances The best working fluid can be considered as the one with the highest COP while the circulation ratio f, the generator temperature and the value of the heat transferred by the solution heat exchanger divided by the heat rejected by the evaporator, Qhs/Qe, are as low as possible. Lower circulation ratio f and lower value of Qhs/Qe leads to a smaller size of the absorption machine. As can be seen in Table 2, the maximum COP in the TPL cycle was obtained by the solution R125-DMEU (0.59) followed by R125-NMP (0.57), R125-DMAC (0.57), R125-MCL (0.56), R125-DMPU (0.54) and R125-DMETEG (0.49). It should be pointed out that although R125v DMEU has the highest COP at low generator temperature (88 C), its circulation ratio is the Table 3 v COP, f and Qhs/Qe for various absorbents with R125 in both cycles at generator temperature of 100 C and evaporv ator temperature of 5 C DPL
DMETEG NMP MCL DMAC DMEU DMPU
TPL
(TPL–DPL)
COP
f
Qhs/Qe
COP
f
Qhs/Qe
d(COP) %
d(f) %
d(Qhs/Qe) %
0.46 0.54 0.53 0.54 0.56 0.52
4.2 3.0 4.1 3.4 4.2 3.9
4.4 2.6 3.7 3.0 3.0 2.9
0.49 0.56 0.55 0.56 0.59 0.54
3.35 2.6 3.35 2.9 3.3 3.25
3.3 1.7 2.8 2.35 2.2 2.3
2.9 3.3 4.8 4.1 4.8 4.9
20.2 13.3 18.3 14.7 21.4 16.7
25.0 34.6 24.3 21.7 26.7 20.7
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highest (4.2) and the Qhs/Qe value is moderate in comparison with the other solutions. Therefore, the second and third solutions have to be taken into consideration. Since the COP, the circulation ratio and the value of Qhs/Qe of R125-NMP and R125-DMAC are almost the same, R125-NMP is preferable among them because it can be operated under lower generator temv v perature, 87 C, compared to 93 C with R125-DMAC. This order was obtained also with the DPL cycle. When operating conditions of the cycle were known and the generator temperature was taken v as 100 C (see Table 3), the highest COP was achieved in the following order: R125-DMEU (0.59) followed by R125-DMAC (0.56), R125-NMP (0.56), R125-MCL (0.55), R125-DMPU (0.54) and R125-DMETEG (0.49). It should be pointed out that although R125-DMEU has the highest COP, its circulation ratio and the value of Qhs/Qe are moderate in comparison with the other solutions. Therefore, the second and third solutions have to be taken into consideration. Although, the COP of R125-DMAC is a little bit higher than the COP of R125-NMP, the later is preferable because its lower f (2.6 compared to 2.9 with R125-DMAC) and the value of Qhs/ Qe is lower (1.7 compared to 2.35 with R125-DMAC).
4. Conclusions Integrating a specially designed jet ejector at the absorber inlet resulted in an advanced TPL single-stage absorption cycle. The jet ejector facilitated pressure recovery and improved the mixing between the weak solution and the refrigerant vapor coming from the evaporator. The influence of the jet ejector on the performance of the TPL absorption cycle was evaluated by comparing the performance of the TPL absorption cycle with that of the DPL cycle under the same operating conditions. The analysis was made with six environmentally acceptable working fluids; R125-DMETEG, R125-NMP, R125-MCL, R125-DMAC, R125-DMEU and R125DMPU. Among the four options of improvements in the TPL absorption cycle performances [12], the ability to reduce the circulation ratio f was investigated. From this study, the improvement in the performance of the TPL cycle in comparison to the DPL cycle with the mentioned working fluids is evident. The TPL cycle can be operated by v lower heat source (9–12 C) and yet with higher COP’s of 5–6% and required smaller solution heat exchanger up to 19% (see Table 2). However, for the same operation condition, the TPL cycle provided higher COP’s of 3–5%, lower circulation ratios of 13–21% and required smaller solution heat exchanger of 22–35% (see Table 3). It should be noted that these improvements are functions of the operating conditions and the thermophysical properties of the working fluids. The interdependencies between the different parameters are quite complex. It is obvious, that at the end of the day cost functions for the different apparatuses have to be incorporated into the optimisation, so that an economic optimum can be found. Concerning the working fluids based on the refrigerant R125 and the six absorbents, it was found that the solution R125-DMEU showed the best performances followed by R125-NMP, R125-DMAC, R125-MCL, R125-DMPU and R125-DMETEG. The calculated COP’s in the TPL cycle were in the range of 0.45–0.59 and the circulation ratios were in the range of 2–5.
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Acknowledgements This work was partially supported by the German–Israeli Scientific Cooperation Unit, Ministry of Science, under the program BMBF-MOS, Grant-GR-1718.
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