- Email: [email protected]
water diffusion absorption machine
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Available online at www.sciencedirect.com
ScienceDirect w w w . i i fi i r . o r g
journal homepage: www.elsevier.com/locate/ijrefrig
Experimental and thermodynamic investigation of an ammonia/water diffusion absorption machine Souha Mazouz*, Rami Mansouri, Ahmed Bellagi U.R. Thermique and Thermodynamique des Procedes Industriels, Ecole Nationale d'Ingenieurs de Monastir, ENIM, University of Monastir, Tunisia
article info
abstract
Article history:
In this paper the results of experimental and thermodynamic investigations on a com-
Received 19 February 2014
mercial absorption diffusion device are presented. Two alternative experimental methods,
Received in revised form
steady state and dynamic method, are used to evaluate the characteristics and the cooling
18 May 2014
capacity of the machine. All essential features of the machine are determined specially the
Accepted 4 June 2014
overall heat transfer coefficients (UA)ext and (UA)int, respectively, 0.43 W K1 and 0.21 W K1.
Available online 12 June 2014
The tests are performed under practical interior ambient air conditions and variable heat loads. For a heat supply of 42 W and a generator temperature of 185 C, a COP of 0.12 is found. © 2014 Elsevier Ltd and IIR. All rights reserved.
Keywords: Diffusion absorption Refrigeration Ammoniaewater Performance
rimentale et thermodynamique d'une machine a Etude expe diffusion d'ammoniac/eau absorption a diffusion ; Re frige ration ; Ammoniac-eau ; Performance Mots cles : Absorption a
1.
Introduction
In a perspective of sustainable development and a more rational use of energy, the research challenge is to find innovative technologies that balance the demands of efficient cold chain and minimized environmental impacts. This is the reason behind the renewed interest for absorption technique for production of cold. This technique, although old for over a century, has several advantages. By using heat as driving force, it bypasses electricity and makes possible the * Corresponding author. Tel.: þ216 98 826 181. E-mail address: [email protected] (S. Mazouz). http://dx.doi.org/10.1016/j.ijrefrig.2014.06.002 0140-7007/© 2014 Elsevier Ltd and IIR. All rights reserved.
valorization of waste heat of intermediate temperatures (150e200 C) and the use of solar thermal energy provided by solar collectors such as evacuated-tube collectors. Diffusion absorption machines (DAR) have another major advantage: they operate at a uniform pressure and therefore have no moving part. The absence of any mechanical work input allows a silent and very reliable machine suitable for hotel rooms, offices and camping cars. Many papers dealing with the performance and the optimization of the DAR cycle have been published. Comparison
84
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Nomenclature a, b, c, d, e, a, b regression parameters COP coefficient of performance heat capacity, J K1 cp Q_ heat transfer, W T temperature, K ~ T mean evaporator temperature, K t time, s (UA) global heat transfer coefficient, W K1 Subscripts 0 initial time a ambient abs absorber cond condenser elec electric evap evaporator ext exterior gen generator i refrigerator interior id COP achieved with Carnot thermodynamic cycle i,0 interior temperature at the beginning of the tests int interior rec rectifier w water w,0 water temperature at the beginning of the tests
between the working fluid systems NH3eH2O, NH3eLiNO3 and NH3eNaSCN absorption systems is found in the papers of Bourseau and Bugarel (1986). Solar operated absorption diffusion refrigerating systems are reported in Valizadeh and Ashrafi (1996), Gutierrez (1988) and Wang (2012). Al Shemmeri and Wang (2003) developed a mathematical model for analyzing the performance and operating parameters of this type of machine. The results show that the influence of the gasegas exchanger is critical. Substitution of the pressure equalizing gas hydrogen by helium has been also investigated by Herold (1996), Prakash-Maiya (2003). Recently, Wang et al. studied a diffusion absorption refrigerator in which the systems R23/R134a DMF and helium are used as, respectively, as refrigerants, absorbent and diffusion gas. With a generator temperature between 110 and 160 C and ambient temperature in the range 10e28 C, a very low evaporating temperature has been reached 40 C, but the COP was less than 0.07 (Wang et al., 2011). The performance of the TFEeTEGDME and NH3eH2O DAR cycles was compared in terms of the COP and circulation ratio by Zhen et al. (2013). They concluded that the TFEeTEGDME mixture is a good working fluid for the DAR cycle. In 1995, Chen et al. (1996) developed a simulation model for a 15 kW cooling capacity machine. The basic configuration reached a COP of 0.4. They also conceived a new generator which reuses the rejected heat in the rectifier. The new configuration has slightly improved the COP (5%) compared to the original one. Akam et al. (1999) conducted an experimental study of a refrigeration absorptionediffusion loop. Tests are carried out for two heating modes: electricity and butane
combustion gas. They concluded that in both modes of heating, operation of the machine is not a problem but the COP values are higher in the case of electric heating. Another study of a diffusioneabsorption machine, using helium as inert gas, has been published by Srikhirin and Aphornratana (2002). Mass and energy balances are applied to each element of the cycle. For the modeling of the bubble pump, a correlation established from an air-lift is used. They noted that the bubble pump must be designed so that the rates of refrigerant flow and weak solution are equal. If too much solution is pumped, then heat loss occurs. They found a COP ranging from 0.09 to 0.15 for a cooling capacity between 100 and 180 W. Zohar et al. (2005) have developed a thermodynamic model for the simulation of an absorptionediffusion refrigeration cycle. Simulations were performed using the solver EES (Engineering Equation Solver) (EES, 2003). It is found that the largest COP is reached for a concentration of the rich solution ranging between 0.25 and 0.3 and a driving heat temperature in the range 195e205 C. The recommended values for the concentrations of ammonia rich and poor solutions are, respectively, 0.3 and 0.1. Helium has been found preferable to hydrogen as inert gas leading to a higher COP. On the basis of a thermodynamic model of the machine, Zohar et al. (2007) tried two configurations of a machine with and without sub-cooling of the condensate before it enters the evaporator. This study shows that the COP of the cycle without sub-cooling of the refrigerant liquid is 14e20% higher and the best performance is obtained when the mass fraction of ammonia-rich solution varies in the range from 0.25 to 0.4. They also showed that better the degree of rectification the higher the COP. Starace and De Pascalis (2012, 2013) elaborated a thermodynamic model of the DAR cycle without any assumption about the purity of the refrigerant ammonia exiting the rectifier. This model was compared with that of Zohar et al. and the results show a higher accuracy in predicting the real operation of the system. The model was experimentally validated using a prototype with a bubble pump coupled to a domestic magnetron to reduce the starting transient of the circuit. Jakob et al. (2008) studied both experimentally and theoretically a 2.5 kW diffusion machine performed by solar energy and designed for air-conditioning. With a generator temperature of 115 C and an evaporator temperature of 5 C, an experimental COP of 0.38 is found. The objective of the present paper is to present a methodology for the evaluation of the cooling capacity of a commercial absorption diffusion machine in both steady state and transient mode. First, the coefficients of heat transfer between the refrigerator interior and ambient air and between the refrigerator interior and the evaporator (UA)ext and (UA)int, are determined. The evolution of the COP with supplied generator heat is also evaluated and discussed. To this purpose, a series of tests in transient and steady mode is performed with heating powers ranging from 10 to 70 W by room temperature between 20 and 30 C.
2. Operating principle and experimental procedure The machine is a small capacity refrigerator designed for hotel rooms and camping cars. It runs on electrical power and
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
operates according to the PlateneMunters thermodynamic cycle (Von Platen and Munters, 1928) with three components working fluid system: ammoniaewaterehydrogen. As shown in Fig. 1, the machine consists of a rectifier, a condenser, an absorber, an evaporator and a generator in the form of two externally heated coaxial tubes. The inner tube is the bubble pump and receives heat indirectly through the solution flowing in the annular space acting as a boiler. The heat Q_ gen supplied to the generator will allow degassing a fraction of the ammonia contained in the rich solution flowing from the absorber. The exiting ammonia vapors are further purified in the rectifier by partial condensation of the rest of water vapors. The heat rejected in the rectifier is noted Q_ rec . The almost pure ammonia condenses thereafter in the condenser, by rejecting the heat Q_ cond to the environment. The condensate is further sub-cooled in the gas heat exchanger before evaporating at low temperature in the evaporator by lowering its partial pressure through the addition of the inert gas returning from the absorber, thus producing the useful cooling capacity Q_ evap . The gas mixture exiting the evaporator is fed to the bottom of the absorber where it flows upwards, counter-currently to the weak solution coming from the generator and fed at its top. The exothermic absorption of ammonia in the aqueous solution is associated with a rejection of the heat Q_ abs to the ambient. The machine is equipped with K-type thermocouples placed at the entrance and exit of each of its components and connected via a data acquisition unit (34970A AGILENT) to a computer (Fig. 2). The electric heating element is connected to a power controller. Ambient temperature Ta is also continuously measured and registered. The experimental tests are carried out by first adjusting the heating power to the generator, then starting the measurements and storing all the experimental data by monitoring the evolution of the various temperatures until reaching the steady state.
3.
85
Measurement results
Basic tests were conducted with maximum generator power. Monitoring different temperature profiles (Figs. 3e5) allows us to determine the minimum time required for the machine to stabilize. These results show that at least 2 h is necessary to reach the steady state and it is the mean evaporator temperature that takes the longest time to stabilize. In the following, this temperature will be taken as guide to assess whether the steady state is reached or not when evaluating tests in steady state mode. Furthermore, it is noted that a minimal heat power (roughly 20 W) has to be supplied to the generator to ensure the functioning and the stability of the refrigerator (Fig. 6). The COP of the machine is defined as COP ¼
Q_ evap Q_ gen
(1)
While Q_ gen is easily determined, the cooling capacity, Q_ evap is deduced from measured temperatures and heat transfer models. The internal and external overall heat transfer coefficients of the DAR, (UA)ext and (UA)int are evaluated in separate experiments. In steady state, they are correlated by the equation ~evap Q_ evap ¼ ðUAÞext ðTa Ti Þ ¼ ðUAÞint Ti T
4.
(2)
Determination of (UA)ext
To determine the overall heat transfer coefficient (UA)ext of the refrigerator in steady state mode, the standard method consists of heating the interior of the refrigerator by electric power Q_ elec supplied to a resistor placed inside and to measure the final internal temperature Ti reached. By assuming that
Fig. 1 e Schematic view of the refrigerator.
86
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 2 e Experimental set-up.
Fig. 3 e Evolution of generator, condenser and ambient temperatures.
87
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 4 e Evolution of gas heat exchanger, evaporator and interior temperatures. the enclosure is completely sealed and does not undergo openings during use, thus heat exchange is taking place only through the walls, the overall heat transfer coefficient (UA)ext is then
mathematical expression of the temperature Ti. Assuming that the overall heat transfer coefficient (UA)ext is constant and a uniform refrigerator temperature, energy balance for the refrigerator cabin writes:
Q_ elec ¼ ðUAÞext ðTi Ta Þ
cp
(3)
Tests are conducted at ambient temperatures between 23 and 25 C with a maximum heating power of 25 W. Fig. 7 shows an example of measurement in the case of Q_ elec ¼ 5:8 W. (UA)ext is determined by floating Q_ elec vs. (Ti Ta) for various electric heat powers supplied to the resistor (Fig. 8). It is found for our refrigerator ðUAÞext ¼ 0:43 W=K
(4)
An alternative method for the determination of (UA)ext in dynamic mode is also tested. The first step is to develop a descriptive model of the measured internal temperature Ti. It is found that for all tests the real part of the function 2 0 1e 31 B 6 Bt þ dðln2Þ1e cC 7C B 6 C C7 Ti ¼ a þ bB1 exp6 B C A7 @ 4 @ 5A d
dTi ¼ Q_ elec ðUAÞext ðTa Ti Þ dt
(6)
The solution of the ODE writes: Q_ elec þ Ta a 1 exp ðUAÞext ðt t0 Þ
Ti ¼ Ti;0 exp ðUAÞext ðt t0 Þ þ
! (7)
where a ¼ ðUAÞext =cp As third step finally nonlinear regression procedure is applied to determine both parameters, a and (UA)ext, of this model. As shown in Fig. 10, the internal temperature is well described with this model. It is found in this particular case that (UA)ext ¼ 0.43 W K1.
0
(5)
where a, b, c, d and e are regression parameters, represents with good accuracy the experimental data (Fig. 9). In a second step, a simple heat transfer model of the refrigerator interior is developed and integrated to find a
5. Experimental determination of (UA)int and Q_ evap The procedure is to load the refrigerator with different quantities of water. The cooling capacity is determined by considering the heat exchange between the tubes of the evaporator on the one hand and the water load on the other.
88
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 5 e Evolution of absorber temperatures.
Neglecting the amount of heat transferred to the atmosphere of the refrigerator in comparison to that transferred to the water, the evolution of the water temperature Tw as a function of the time t, allows us to calculate the cooling capacity Q_ evap . In steady state, the overall heat transfer coefficients are correlated by the equation ~evap Q_ evap ¼ ðUAÞext ðTa Tw Þ ¼ ðUAÞint Tw T
(8)
As the value of (UA)ext has been already determined (0.43 W K1), the internal overall heat transfer coefficient between evaporator and water, (UA)int, can then readily be deduced, 0.21 W K1. In transient state, the energy balance writes in analogy to Eq. (6): cp
dTw ~evap ¼ ðUAÞext ðTa Tw Þ ðUAÞint Tw T dt
(9)
Using the Laplace transform and after simplification the expression of Tw becomes: b ð1 exp½ ðaþ bÞðt t0 ÞÞ aþb
Fig. 6 e Oscillating temperatures of generator (red) and evaporator (blue) when the machine is supplied with insufficient heat power (17.5 W). (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
Tw ¼ Tw0 exp½ ðaþbÞðt t0 Þþ Ta Zt þaexp½ ðaþbÞðt t0 Þ
~evap ðtÞexp½ðaþbÞtdt T
t0
(10)
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 7 e Evolution of inside (red) and ambient temperatures (orange) during the test for Qelec ¼ 5.8 W. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
A nonlinear regression procedure is used to determine the parameters a and b of this model with: f ðUAÞint ¼ b ðUAÞext
(11)
Fig. 8 e Determination of (UA)ext in the stationary mode.
89
Fig. 9 e Model (red) and data of the inside temperature (blue). (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
As shown in Fig. 11, the water temperature is well predicted by this model. The values of a and b were found to be 0.0067 and 0.0129 thus leading to (UA)int ¼ 0.20 W K1 in good agreement with the value found in steady mode.
Fig. 10 e Heat transfer model of the internal temperature (red) and data (blue) for Qelec ¼ 5.8 W. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
90
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 11 e Heat transfer model of the water temperature (red) and data (green) for determination of (UA)int in dynamic mode. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
6.
Thermodynamic analysis
Once the internal and external heat transfer coefficients are determined, the cooling capacity and the COP can be calculated. The evolution of these machine characteristics vs the generator heat input are illustrated in Figs. 12 and 13. A simultaneous increase in both the cooling capacity and the COP is observed when the heat input is increased from 39 to 42 W. As the driving heat power increases, more cooling capacity is provided until a maximum value is reached. An increase in the generator input means more refrigerant vapor generated thus extracting more heat in the evaporator. Further heating at the generator is not necessary beyond the optimal conditions because the excess heat added is rejected as waste heat to the cooling air in the rectifier and the condenser. As a consequence, the COP beyond this limit gets lower. The most efficient NH3eH2O absorption cycle is the GAX cycle (Herold et al., 1996). In order to compare the performance of our diffusion machine with this chiller, it is interesting to calculate the theoretical upper limit of the COP, COPid achieved with the Carnot thermodynamic cycle and depending only on the temperatures Tgen, Tcond, Tabs and Tevap COPid ¼
Tgen Tabs Tevap Tcond Tevap Tgen
Fig. 12 e COP vs. generator heat supply.
lower performance in comparison with the theoretical expected COP is due mainly to the absorber (El May et al., 2011) which is the most critical component of the machine, due to a non-equilibrium state of solution in the absorber. Comparing this performance of 0.5 with the experimental value of 0.12 found for our diffusion machine under similar conditions, the experimental COP is very low. This is due to the slow intensity of heat and mass transfer in the various components of the machine where the fluids are circulating by natural convection (gas) and gravity (liquid solution). Furthermore the circulation of the liquid solution is limited by the low
(12)
For the case depicted in Figs. 3e5 (Tgen ¼ 185 C, Tcond ¼ Tabs ¼ 40 C, Tevap ¼ 14 C) the value of the theoretical COP is 1.5. In practice the COP of the absorption GAX cycle operating at the same temperatures is about 0.5 (El May et al., 2011). This
Fig. 13 e Cooling capacity vs. generator heat supply.
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
performance of the bubble pump. It must be noticed also that a fraction of the produced cold is used to cool the auxiliary gas (hydrogen) flowing in the evaporator.
7.
Conclusion
In this study, an experimental investigation of a 20 W diffusion refrigerating machine has been conducted. A series of tests in both transient and steady state was achieved with heating powers ranging from 10 to 70 W and ambiences between 20 and 30 C. All the features of the machine were determined, specially the coefficients (UA)ext and (UA)int. In steady state method, a heating power of 25 W was used to heat the interior of the refrigerator. The (UA)ext found is 0.43 W K1. In the dynamic method, using a regression model for the internal temperature Ti and a simple heat transfer model of the refrigerator interior, the (UA)ext is found equal to 0.43 W K1. A nonlinear regression procedure is also used to determine the (UA)int leading to a value of 0.21 W K1. The best performance of the refrigerator is obtained experimentally with a heat supply of 42 W by a generator temperature of 185 C. The COP of the machine is found equal to 0.12. A more detailed thermodynamic analysis using an exergy balance may help explain how the COP is degraded in the various components of the refrigerator and propose means to improve system efficiency.
references
Akam, S.A., Said, N., Ouchicha, Z., Bellal, B., 1999. rimentation d'une boucle a absorption NH3eH2O. Rev. Expe Energ. Ren. Valorisation, 17e22. Al Shemmeri, T., Wang, Y., 2003. Theoretical investigation and parameter study of a diffusion absorption refrigeration system. In: Proceedings of the 21st IIR International Congress of Refrigeration, Washington, USA. Bourseau, P., Bugarel, R., 1986. Absorption diffusion machines: comparison of the performances of NH3-H2O and NH3-NaSCN. Int. J. Refrigeration 9 (n 4), 206e214. Chen, J., Kim, K.J., Herold, K.E., 1996. Performance enhancement of a diffusion absorption refrigerator. Int. J. Refrigeration 19 (n 3), 208e218.
91
EES e Engineering Equation Solver, Academic Version 6.563, 2003. F-Chart Software. www.fchart.com. El May, S., Boukholda, I., Bellagi, A., 2011. Energetic and exergetic analysis of a commercial ammonia-water. Int. J. Exergy 8 (No.1), 33e50. Gutierrez, F., 1988. Behavior of a house hold absorptionediffusion refrigerator adapted to autonomous solar operation. Sol. Energy 40 (n .1), 17e23. Herold, K.E., 1996. Diffusion-Absorption Heat Pump. Final Reports for Gas Research Institute, GRI-96/0271. Gas Research Institute, USA. Herold, K.E., Radermacher, R., Klein, S.A., 1996. Absorption Chillers and Heat Pumps. CRC Press, Boca Raton, FL. Jakob, U., Eicker, U., Schneider, D., Taki, A.H., Cook, M.J., 2008. Simulation and experimental investigation into diffusion absorption cooling machines for air conditioning applications. Appl. Therm. Eng. 28 (n 10), 1138e1150. Prakash-Maiya, M., 2003. Studies on gas circuit of diffusion absorption refrigerator. In: Proceedings of the 21st IIR International Congress of Refrigeration, Washington, USA. Srikhirin, P., Aphornratana, S., 2002. Investigation of a diffusion absorption refrigerator. Appl. Therm. Eng. 22 (n 11), 1181e1193. Starace, G., De Pascalis, L., 2012. An advanced analytical model of the diffusion absorption refrigerator cycle. Int. J. Refrigeration 35, 605e612. Starace, G., De Pascalis, L., 2013. An enhanced model for the design of diffusion absorption refrigerators. Int. J. Refrigeration 36, 1495e1503. Valizadeh, H., Ashrafi, N., 1996. A continuous cycle solar thermal refrigeration system. Renew. Energy 1e4, 632e640. Von Platen, B.C., Munters, C.G., 1928. Refrigerator US Patent 1 685 764, pp. 685e764. Wang, H., 2012. A new style solar-driven diffusion absorption refrigerator and its operating characteristics. Energy Proc. 18, 681e692. Wang, Q., Gong, L., Wang, J.P., Sun, T.F., Cui, K., Chen, G.M., 2011. A numerical investigation of a diffusion absorption refrigerator operating with the binary refrigerant for low temperature applications. Appl. Therm. Eng. 31, 1763e1769. Zhen, L., Yong, L., Huashan, L., Xianbiao, B., Weibin, M., 2013. Performance analysis of a diffusion absorption refrigeration cycle working with TFEeTEGDME mixture. Energy Build. 58, 86e92. Zohar, A., Jelinek, M., Levy, A., Borde, I., 2005. Numerical investigation of a diffusion absorption refrigeration cycle. Int. J. Refrigeration 28 (n 4), 515e525. Zohar, Jelinek, M., Levy, A., Borde, I., 2007. The influence of diffusion absorption refrigeration cycle configuration on the performance. Appl. Therm. Eng. 27 (n 13), 2213e2219.
Available online at www.sciencedirect.com
ScienceDirect w w w . i i fi i r . o r g
journal homepage: www.elsevier.com/locate/ijrefrig
Experimental and thermodynamic investigation of an ammonia/water diffusion absorption machine Souha Mazouz*, Rami Mansouri, Ahmed Bellagi U.R. Thermique and Thermodynamique des Procedes Industriels, Ecole Nationale d'Ingenieurs de Monastir, ENIM, University of Monastir, Tunisia
article info
abstract
Article history:
In this paper the results of experimental and thermodynamic investigations on a com-
Received 19 February 2014
mercial absorption diffusion device are presented. Two alternative experimental methods,
Received in revised form
steady state and dynamic method, are used to evaluate the characteristics and the cooling
18 May 2014
capacity of the machine. All essential features of the machine are determined specially the
Accepted 4 June 2014
overall heat transfer coefficients (UA)ext and (UA)int, respectively, 0.43 W K1 and 0.21 W K1.
Available online 12 June 2014
The tests are performed under practical interior ambient air conditions and variable heat loads. For a heat supply of 42 W and a generator temperature of 185 C, a COP of 0.12 is found. © 2014 Elsevier Ltd and IIR. All rights reserved.
Keywords: Diffusion absorption Refrigeration Ammoniaewater Performance
rimentale et thermodynamique d'une machine a Etude expe diffusion d'ammoniac/eau absorption a diffusion ; Re frige ration ; Ammoniac-eau ; Performance Mots cles : Absorption a
1.
Introduction
In a perspective of sustainable development and a more rational use of energy, the research challenge is to find innovative technologies that balance the demands of efficient cold chain and minimized environmental impacts. This is the reason behind the renewed interest for absorption technique for production of cold. This technique, although old for over a century, has several advantages. By using heat as driving force, it bypasses electricity and makes possible the * Corresponding author. Tel.: þ216 98 826 181. E-mail address: [email protected] (S. Mazouz). http://dx.doi.org/10.1016/j.ijrefrig.2014.06.002 0140-7007/© 2014 Elsevier Ltd and IIR. All rights reserved.
valorization of waste heat of intermediate temperatures (150e200 C) and the use of solar thermal energy provided by solar collectors such as evacuated-tube collectors. Diffusion absorption machines (DAR) have another major advantage: they operate at a uniform pressure and therefore have no moving part. The absence of any mechanical work input allows a silent and very reliable machine suitable for hotel rooms, offices and camping cars. Many papers dealing with the performance and the optimization of the DAR cycle have been published. Comparison
84
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Nomenclature a, b, c, d, e, a, b regression parameters COP coefficient of performance heat capacity, J K1 cp Q_ heat transfer, W T temperature, K ~ T mean evaporator temperature, K t time, s (UA) global heat transfer coefficient, W K1 Subscripts 0 initial time a ambient abs absorber cond condenser elec electric evap evaporator ext exterior gen generator i refrigerator interior id COP achieved with Carnot thermodynamic cycle i,0 interior temperature at the beginning of the tests int interior rec rectifier w water w,0 water temperature at the beginning of the tests
between the working fluid systems NH3eH2O, NH3eLiNO3 and NH3eNaSCN absorption systems is found in the papers of Bourseau and Bugarel (1986). Solar operated absorption diffusion refrigerating systems are reported in Valizadeh and Ashrafi (1996), Gutierrez (1988) and Wang (2012). Al Shemmeri and Wang (2003) developed a mathematical model for analyzing the performance and operating parameters of this type of machine. The results show that the influence of the gasegas exchanger is critical. Substitution of the pressure equalizing gas hydrogen by helium has been also investigated by Herold (1996), Prakash-Maiya (2003). Recently, Wang et al. studied a diffusion absorption refrigerator in which the systems R23/R134a DMF and helium are used as, respectively, as refrigerants, absorbent and diffusion gas. With a generator temperature between 110 and 160 C and ambient temperature in the range 10e28 C, a very low evaporating temperature has been reached 40 C, but the COP was less than 0.07 (Wang et al., 2011). The performance of the TFEeTEGDME and NH3eH2O DAR cycles was compared in terms of the COP and circulation ratio by Zhen et al. (2013). They concluded that the TFEeTEGDME mixture is a good working fluid for the DAR cycle. In 1995, Chen et al. (1996) developed a simulation model for a 15 kW cooling capacity machine. The basic configuration reached a COP of 0.4. They also conceived a new generator which reuses the rejected heat in the rectifier. The new configuration has slightly improved the COP (5%) compared to the original one. Akam et al. (1999) conducted an experimental study of a refrigeration absorptionediffusion loop. Tests are carried out for two heating modes: electricity and butane
combustion gas. They concluded that in both modes of heating, operation of the machine is not a problem but the COP values are higher in the case of electric heating. Another study of a diffusioneabsorption machine, using helium as inert gas, has been published by Srikhirin and Aphornratana (2002). Mass and energy balances are applied to each element of the cycle. For the modeling of the bubble pump, a correlation established from an air-lift is used. They noted that the bubble pump must be designed so that the rates of refrigerant flow and weak solution are equal. If too much solution is pumped, then heat loss occurs. They found a COP ranging from 0.09 to 0.15 for a cooling capacity between 100 and 180 W. Zohar et al. (2005) have developed a thermodynamic model for the simulation of an absorptionediffusion refrigeration cycle. Simulations were performed using the solver EES (Engineering Equation Solver) (EES, 2003). It is found that the largest COP is reached for a concentration of the rich solution ranging between 0.25 and 0.3 and a driving heat temperature in the range 195e205 C. The recommended values for the concentrations of ammonia rich and poor solutions are, respectively, 0.3 and 0.1. Helium has been found preferable to hydrogen as inert gas leading to a higher COP. On the basis of a thermodynamic model of the machine, Zohar et al. (2007) tried two configurations of a machine with and without sub-cooling of the condensate before it enters the evaporator. This study shows that the COP of the cycle without sub-cooling of the refrigerant liquid is 14e20% higher and the best performance is obtained when the mass fraction of ammonia-rich solution varies in the range from 0.25 to 0.4. They also showed that better the degree of rectification the higher the COP. Starace and De Pascalis (2012, 2013) elaborated a thermodynamic model of the DAR cycle without any assumption about the purity of the refrigerant ammonia exiting the rectifier. This model was compared with that of Zohar et al. and the results show a higher accuracy in predicting the real operation of the system. The model was experimentally validated using a prototype with a bubble pump coupled to a domestic magnetron to reduce the starting transient of the circuit. Jakob et al. (2008) studied both experimentally and theoretically a 2.5 kW diffusion machine performed by solar energy and designed for air-conditioning. With a generator temperature of 115 C and an evaporator temperature of 5 C, an experimental COP of 0.38 is found. The objective of the present paper is to present a methodology for the evaluation of the cooling capacity of a commercial absorption diffusion machine in both steady state and transient mode. First, the coefficients of heat transfer between the refrigerator interior and ambient air and between the refrigerator interior and the evaporator (UA)ext and (UA)int, are determined. The evolution of the COP with supplied generator heat is also evaluated and discussed. To this purpose, a series of tests in transient and steady mode is performed with heating powers ranging from 10 to 70 W by room temperature between 20 and 30 C.
2. Operating principle and experimental procedure The machine is a small capacity refrigerator designed for hotel rooms and camping cars. It runs on electrical power and
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
operates according to the PlateneMunters thermodynamic cycle (Von Platen and Munters, 1928) with three components working fluid system: ammoniaewaterehydrogen. As shown in Fig. 1, the machine consists of a rectifier, a condenser, an absorber, an evaporator and a generator in the form of two externally heated coaxial tubes. The inner tube is the bubble pump and receives heat indirectly through the solution flowing in the annular space acting as a boiler. The heat Q_ gen supplied to the generator will allow degassing a fraction of the ammonia contained in the rich solution flowing from the absorber. The exiting ammonia vapors are further purified in the rectifier by partial condensation of the rest of water vapors. The heat rejected in the rectifier is noted Q_ rec . The almost pure ammonia condenses thereafter in the condenser, by rejecting the heat Q_ cond to the environment. The condensate is further sub-cooled in the gas heat exchanger before evaporating at low temperature in the evaporator by lowering its partial pressure through the addition of the inert gas returning from the absorber, thus producing the useful cooling capacity Q_ evap . The gas mixture exiting the evaporator is fed to the bottom of the absorber where it flows upwards, counter-currently to the weak solution coming from the generator and fed at its top. The exothermic absorption of ammonia in the aqueous solution is associated with a rejection of the heat Q_ abs to the ambient. The machine is equipped with K-type thermocouples placed at the entrance and exit of each of its components and connected via a data acquisition unit (34970A AGILENT) to a computer (Fig. 2). The electric heating element is connected to a power controller. Ambient temperature Ta is also continuously measured and registered. The experimental tests are carried out by first adjusting the heating power to the generator, then starting the measurements and storing all the experimental data by monitoring the evolution of the various temperatures until reaching the steady state.
3.
85
Measurement results
Basic tests were conducted with maximum generator power. Monitoring different temperature profiles (Figs. 3e5) allows us to determine the minimum time required for the machine to stabilize. These results show that at least 2 h is necessary to reach the steady state and it is the mean evaporator temperature that takes the longest time to stabilize. In the following, this temperature will be taken as guide to assess whether the steady state is reached or not when evaluating tests in steady state mode. Furthermore, it is noted that a minimal heat power (roughly 20 W) has to be supplied to the generator to ensure the functioning and the stability of the refrigerator (Fig. 6). The COP of the machine is defined as COP ¼
Q_ evap Q_ gen
(1)
While Q_ gen is easily determined, the cooling capacity, Q_ evap is deduced from measured temperatures and heat transfer models. The internal and external overall heat transfer coefficients of the DAR, (UA)ext and (UA)int are evaluated in separate experiments. In steady state, they are correlated by the equation ~evap Q_ evap ¼ ðUAÞext ðTa Ti Þ ¼ ðUAÞint Ti T
4.
(2)
Determination of (UA)ext
To determine the overall heat transfer coefficient (UA)ext of the refrigerator in steady state mode, the standard method consists of heating the interior of the refrigerator by electric power Q_ elec supplied to a resistor placed inside and to measure the final internal temperature Ti reached. By assuming that
Fig. 1 e Schematic view of the refrigerator.
86
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 2 e Experimental set-up.
Fig. 3 e Evolution of generator, condenser and ambient temperatures.
87
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 4 e Evolution of gas heat exchanger, evaporator and interior temperatures. the enclosure is completely sealed and does not undergo openings during use, thus heat exchange is taking place only through the walls, the overall heat transfer coefficient (UA)ext is then
mathematical expression of the temperature Ti. Assuming that the overall heat transfer coefficient (UA)ext is constant and a uniform refrigerator temperature, energy balance for the refrigerator cabin writes:
Q_ elec ¼ ðUAÞext ðTi Ta Þ
cp
(3)
Tests are conducted at ambient temperatures between 23 and 25 C with a maximum heating power of 25 W. Fig. 7 shows an example of measurement in the case of Q_ elec ¼ 5:8 W. (UA)ext is determined by floating Q_ elec vs. (Ti Ta) for various electric heat powers supplied to the resistor (Fig. 8). It is found for our refrigerator ðUAÞext ¼ 0:43 W=K
(4)
An alternative method for the determination of (UA)ext in dynamic mode is also tested. The first step is to develop a descriptive model of the measured internal temperature Ti. It is found that for all tests the real part of the function 2 0 1e 31 B 6 Bt þ dðln2Þ1e cC 7C B 6 C C7 Ti ¼ a þ bB1 exp6 B C A7 @ 4 @ 5A d
dTi ¼ Q_ elec ðUAÞext ðTa Ti Þ dt
(6)
The solution of the ODE writes: Q_ elec þ Ta a 1 exp ðUAÞext ðt t0 Þ
Ti ¼ Ti;0 exp ðUAÞext ðt t0 Þ þ
! (7)
where a ¼ ðUAÞext =cp As third step finally nonlinear regression procedure is applied to determine both parameters, a and (UA)ext, of this model. As shown in Fig. 10, the internal temperature is well described with this model. It is found in this particular case that (UA)ext ¼ 0.43 W K1.
0
(5)
where a, b, c, d and e are regression parameters, represents with good accuracy the experimental data (Fig. 9). In a second step, a simple heat transfer model of the refrigerator interior is developed and integrated to find a
5. Experimental determination of (UA)int and Q_ evap The procedure is to load the refrigerator with different quantities of water. The cooling capacity is determined by considering the heat exchange between the tubes of the evaporator on the one hand and the water load on the other.
88
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 5 e Evolution of absorber temperatures.
Neglecting the amount of heat transferred to the atmosphere of the refrigerator in comparison to that transferred to the water, the evolution of the water temperature Tw as a function of the time t, allows us to calculate the cooling capacity Q_ evap . In steady state, the overall heat transfer coefficients are correlated by the equation ~evap Q_ evap ¼ ðUAÞext ðTa Tw Þ ¼ ðUAÞint Tw T
(8)
As the value of (UA)ext has been already determined (0.43 W K1), the internal overall heat transfer coefficient between evaporator and water, (UA)int, can then readily be deduced, 0.21 W K1. In transient state, the energy balance writes in analogy to Eq. (6): cp
dTw ~evap ¼ ðUAÞext ðTa Tw Þ ðUAÞint Tw T dt
(9)
Using the Laplace transform and after simplification the expression of Tw becomes: b ð1 exp½ ðaþ bÞðt t0 ÞÞ aþb
Fig. 6 e Oscillating temperatures of generator (red) and evaporator (blue) when the machine is supplied with insufficient heat power (17.5 W). (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
Tw ¼ Tw0 exp½ ðaþbÞðt t0 Þþ Ta Zt þaexp½ ðaþbÞðt t0 Þ
~evap ðtÞexp½ðaþbÞtdt T
t0
(10)
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 7 e Evolution of inside (red) and ambient temperatures (orange) during the test for Qelec ¼ 5.8 W. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
A nonlinear regression procedure is used to determine the parameters a and b of this model with: f ðUAÞint ¼ b ðUAÞext
(11)
Fig. 8 e Determination of (UA)ext in the stationary mode.
89
Fig. 9 e Model (red) and data of the inside temperature (blue). (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
As shown in Fig. 11, the water temperature is well predicted by this model. The values of a and b were found to be 0.0067 and 0.0129 thus leading to (UA)int ¼ 0.20 W K1 in good agreement with the value found in steady mode.
Fig. 10 e Heat transfer model of the internal temperature (red) and data (blue) for Qelec ¼ 5.8 W. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
90
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
Fig. 11 e Heat transfer model of the water temperature (red) and data (green) for determination of (UA)int in dynamic mode. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
6.
Thermodynamic analysis
Once the internal and external heat transfer coefficients are determined, the cooling capacity and the COP can be calculated. The evolution of these machine characteristics vs the generator heat input are illustrated in Figs. 12 and 13. A simultaneous increase in both the cooling capacity and the COP is observed when the heat input is increased from 39 to 42 W. As the driving heat power increases, more cooling capacity is provided until a maximum value is reached. An increase in the generator input means more refrigerant vapor generated thus extracting more heat in the evaporator. Further heating at the generator is not necessary beyond the optimal conditions because the excess heat added is rejected as waste heat to the cooling air in the rectifier and the condenser. As a consequence, the COP beyond this limit gets lower. The most efficient NH3eH2O absorption cycle is the GAX cycle (Herold et al., 1996). In order to compare the performance of our diffusion machine with this chiller, it is interesting to calculate the theoretical upper limit of the COP, COPid achieved with the Carnot thermodynamic cycle and depending only on the temperatures Tgen, Tcond, Tabs and Tevap COPid ¼
Tgen Tabs Tevap Tcond Tevap Tgen
Fig. 12 e COP vs. generator heat supply.
lower performance in comparison with the theoretical expected COP is due mainly to the absorber (El May et al., 2011) which is the most critical component of the machine, due to a non-equilibrium state of solution in the absorber. Comparing this performance of 0.5 with the experimental value of 0.12 found for our diffusion machine under similar conditions, the experimental COP is very low. This is due to the slow intensity of heat and mass transfer in the various components of the machine where the fluids are circulating by natural convection (gas) and gravity (liquid solution). Furthermore the circulation of the liquid solution is limited by the low
(12)
For the case depicted in Figs. 3e5 (Tgen ¼ 185 C, Tcond ¼ Tabs ¼ 40 C, Tevap ¼ 14 C) the value of the theoretical COP is 1.5. In practice the COP of the absorption GAX cycle operating at the same temperatures is about 0.5 (El May et al., 2011). This
Fig. 13 e Cooling capacity vs. generator heat supply.
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 8 3 e9 1
performance of the bubble pump. It must be noticed also that a fraction of the produced cold is used to cool the auxiliary gas (hydrogen) flowing in the evaporator.
7.
Conclusion
In this study, an experimental investigation of a 20 W diffusion refrigerating machine has been conducted. A series of tests in both transient and steady state was achieved with heating powers ranging from 10 to 70 W and ambiences between 20 and 30 C. All the features of the machine were determined, specially the coefficients (UA)ext and (UA)int. In steady state method, a heating power of 25 W was used to heat the interior of the refrigerator. The (UA)ext found is 0.43 W K1. In the dynamic method, using a regression model for the internal temperature Ti and a simple heat transfer model of the refrigerator interior, the (UA)ext is found equal to 0.43 W K1. A nonlinear regression procedure is also used to determine the (UA)int leading to a value of 0.21 W K1. The best performance of the refrigerator is obtained experimentally with a heat supply of 42 W by a generator temperature of 185 C. The COP of the machine is found equal to 0.12. A more detailed thermodynamic analysis using an exergy balance may help explain how the COP is degraded in the various components of the refrigerator and propose means to improve system efficiency.
references
Akam, S.A., Said, N., Ouchicha, Z., Bellal, B., 1999. rimentation d'une boucle a absorption NH3eH2O. Rev. Expe Energ. Ren. Valorisation, 17e22. Al Shemmeri, T., Wang, Y., 2003. Theoretical investigation and parameter study of a diffusion absorption refrigeration system. In: Proceedings of the 21st IIR International Congress of Refrigeration, Washington, USA. Bourseau, P., Bugarel, R., 1986. Absorption diffusion machines: comparison of the performances of NH3-H2O and NH3-NaSCN. Int. J. Refrigeration 9 (n 4), 206e214. Chen, J., Kim, K.J., Herold, K.E., 1996. Performance enhancement of a diffusion absorption refrigerator. Int. J. Refrigeration 19 (n 3), 208e218.
91
EES e Engineering Equation Solver, Academic Version 6.563, 2003. F-Chart Software. www.fchart.com. El May, S., Boukholda, I., Bellagi, A., 2011. Energetic and exergetic analysis of a commercial ammonia-water. Int. J. Exergy 8 (No.1), 33e50. Gutierrez, F., 1988. Behavior of a house hold absorptionediffusion refrigerator adapted to autonomous solar operation. Sol. Energy 40 (n .1), 17e23. Herold, K.E., 1996. Diffusion-Absorption Heat Pump. Final Reports for Gas Research Institute, GRI-96/0271. Gas Research Institute, USA. Herold, K.E., Radermacher, R., Klein, S.A., 1996. Absorption Chillers and Heat Pumps. CRC Press, Boca Raton, FL. Jakob, U., Eicker, U., Schneider, D., Taki, A.H., Cook, M.J., 2008. Simulation and experimental investigation into diffusion absorption cooling machines for air conditioning applications. Appl. Therm. Eng. 28 (n 10), 1138e1150. Prakash-Maiya, M., 2003. Studies on gas circuit of diffusion absorption refrigerator. In: Proceedings of the 21st IIR International Congress of Refrigeration, Washington, USA. Srikhirin, P., Aphornratana, S., 2002. Investigation of a diffusion absorption refrigerator. Appl. Therm. Eng. 22 (n 11), 1181e1193. Starace, G., De Pascalis, L., 2012. An advanced analytical model of the diffusion absorption refrigerator cycle. Int. J. Refrigeration 35, 605e612. Starace, G., De Pascalis, L., 2013. An enhanced model for the design of diffusion absorption refrigerators. Int. J. Refrigeration 36, 1495e1503. Valizadeh, H., Ashrafi, N., 1996. A continuous cycle solar thermal refrigeration system. Renew. Energy 1e4, 632e640. Von Platen, B.C., Munters, C.G., 1928. Refrigerator US Patent 1 685 764, pp. 685e764. Wang, H., 2012. A new style solar-driven diffusion absorption refrigerator and its operating characteristics. Energy Proc. 18, 681e692. Wang, Q., Gong, L., Wang, J.P., Sun, T.F., Cui, K., Chen, G.M., 2011. A numerical investigation of a diffusion absorption refrigerator operating with the binary refrigerant for low temperature applications. Appl. Therm. Eng. 31, 1763e1769. Zhen, L., Yong, L., Huashan, L., Xianbiao, B., Weibin, M., 2013. Performance analysis of a diffusion absorption refrigeration cycle working with TFEeTEGDME mixture. Energy Build. 58, 86e92. Zohar, A., Jelinek, M., Levy, A., Borde, I., 2005. Numerical investigation of a diffusion absorption refrigeration cycle. Int. J. Refrigeration 28 (n 4), 515e525. Zohar, Jelinek, M., Levy, A., Borde, I., 2007. The influence of diffusion absorption refrigeration cycle configuration on the performance. Appl. Therm. Eng. 27 (n 13), 2213e2219.