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2012 Study on a Novel Absorption Referigeration System at Low Cooling Temperatures Yijian He [email protected]

Zuwen Zhu

Xu Gao

Guangming Chen

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He, Yijian; Zhu, Zuwen; Gao, Xu; and Chen, Guangming, "Study on a Novel Absorption Referigeration System at Low Cooling Temperatures" (2012). International Refrigeration and Air Conditioning Conference. Paper 1267. http://docs.lib.purdue.edu/iracc/1267

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Study on a Novel Absorption Refrigeration System at Low Cooling Temperatures

Yi Jian HE*, Zu Wen ZHU, Xu GAO, Guang Ming CHEN

Institute of Refrigeration and Cryogenics, Zhejiang University, Hangzhou, China (Phone/Fax:+86 571 87952464, [email protected])

* Corresponding Author

ABSTRACT

A novel absorption refrigeration system is proposed in this context using R134a+R23 mixed refrigerants and DMF solvent. The new system uses two-staged absorber in series to reduce the evaporation pressure and can obtain a lower refrigeration temperature by the low-grade thermal energy. From thermodynamic analysis, the key factors are discussed on the performances of the new system. Under the same working conditions, comparisons on the performances are also conducted between the new one and an auto-cascade absorption refrigeration system with single-absorber. Experiments are conducted to investigate the key effecting factors on the performances of the new system and a single-absorber auto-cascade absorption refrigeration system. The experimental results show that the new system can achieve a lower refrigeration temperature under the same working condition. The lowest refrigeration temperature of the new system reaches to -62.3℃ with a COP of 0.023 under a generation temperature of 184.4℃, condensing temperature of 18.8℃, and absorption temperatures of 26.4℃ at the low-pressure absorber, and 28.5℃ at the high-pressure absorber.

1. INTRODUCTION

An absorption refrigeration cycle can utilize low-grade thermal energy, and shows potentials of saving energy and relieving trends of global warming. It also holds advantages of using environmental–acceptable substances as its working fluids, Water/Lithium Bromide (H2O/LiBr) solution as well as Ammonia/Water (NH3/H2O) solution. As a result, the researchers in the word-wide area focus great attentions on the development of new absorption refrigeration technologies, and try to extend its applications in the industry. There are a lot of deep freezing demands lower than -20℃ in the industrial processes, such as chemical engineering, food industry as well as pharmaceutical one (Berlinck et al., 1997; José et al., 1998, 2002; Bassols et al., 2002; Bruno et al., 2002; Colonna and Gabrielli, 2003). However, it is formidable challenges for a conventional absorption refrigeration system to get a refrigeration temperature below -20℃, for example, absorption refrigeration systems using H2O/LiBr solution as its working fluids, as well as NH3/H2O solution (Rogdakis and Antonopoulos, 1992; Thioye, 1997; Wang et al., 2000; Venegas et al., 2002). The water (R718) as a natural refrigerant shows excellent performances, such as a great latent heat of vaporization, good thermal stability and appropriate viscosity (Misra et al., 2005; Kilicarslan and Muller, 2005). However, its tri-point is near to 0℃, and a H2O/LiBr absorption refrigeration system can not be used for cooling less than 0℃. Such a system must work under a vacuum condition - besides, its working fluids are greatly corrosive to steel (Anderko and Young, 2000). A NH3/H2O absorption refrigeration system can be used for cooling below 0℃, but generally in cases of a refrigeration temperature higher

than -20℃ (Anderko and Young, 2000; Srikhirin et al., 2001; Colibri, 2012; Apte, 2012). The NH3/H2O absorption refrigeration system generally works under a high pressure, and requires heat sources with rather high temperature as its driving energy. The natural refrigerant R717 shows excellent thermo-physical properties. It is a toxic, strongly irritant and flammable substance, and is also destructively corrosive to copper. Some working fluids based on Freon

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2335, Page 2 refrigerants and organic absorbents show excellent overall performances and can almost completely overcome serious drawbacks of NH3/H2O and H2O/LiBr working fluids (Borde et al., 1995a, 1995b, 1997; Jelinek et al., 1998, 1999, 2002; Levy et al., 2004). They are not toxic and corrosive to carbon steel and copper, and also show good stability. Moreover, a few of such working fluids can not result in ozone depletion (ODP=0) and hardly do global warming (GWP is acceptable). Consequently, these working fluids are environmental-acceptable. In the interest of lowering refrigeration temperature of absorption refrigeration systems, an auto-cascade absorption refrigeration system (ACAR cycle) is proposed, which uses working fluids of a non-azeotropic Freon refrigerant and an organic absorbent (Chen, 2004). There is an auto-cascade cell consisting of a separator, valve and condenser- evaporator, which indicates the differences between an ACAR system and a conventional one, and the mass and energy coupling of refrigerant are conducted in the auto-cascade cell. The refrigerant after condensing is separated into a liquid stream and a vapor stream at the separator of auto-cascade cell. The vapor stream before throttling is condensed by the throttled liquid stream at the condenser-evaporator of auto-cascade cell. The investigations on ACAR systems are carried out and some good progresses are achieved (He and Chen, 2004, 2007a, 2007b; He et al., 2005). On the other hand, it is still an attracting and important challenge to improve performances of absorption refrigeration systems, such as enhancing efficiency of energy conversion and degrading temperatures of driving thermal energy. In fact, it is found that the performances of an ACAR system are mainly determined by the mass flow rates and thermodynamic states of refrigerant before throttling (He and Chen, 2004, 2007a, 2007b; Zhong and Chen, 2004; He et al., 2005). The mass flow rates and thermodynamic states of refrigerant before throttling are dependent on the mass and energy coupling in the auto-cascade cell. Therefore, the mass and energy coupling of auto-cascade processes has key effects on the performances of an ACAR system. From the literatures (He and Chen, 2004; Zhong and Chen, 2004), it is clearly seen that a refrigeration temperature rises up and its COP drops down for the lack of cooling capacity produced by the separated liquid stream at the condenser-evaporator of auto-cascade cell. Hence, the performances of an ACAR system might be improved by optimizing mass and energy coupling of auto-cascade processes. Berlitz et al. (1998) reported that the COP of a NH3/H2O absorption refrigerator for combined heating and cooling cabe significantly improved by incorporating a second absorber. Its second absorber is placed between a condenser and an evaporator, and a stream of refrigerant is absorbed at a medium pressure between the condensing pressure and the evaporating pressure. As a result, the absorbing processes are intensified and the concentration of solution is further increased. The performances are improved for this beneficial intensification of absorbing processes. As we know, the absorption characteristics between a class of substances and absorbents are greatly dependent on its normal boiling point. In general, the solubility of a high-boiling point substance is superior to that of a low-boiling point substance under the same temperature and pressure. Consequently, the components of a non-azeotropic refrigerant show different absorption characteristics in the same absorbent. If the components of non-azeotropic refrigerant are independently absorbed, their absorption processes might be intensified, respectively. From the aforementioned contents and our previous investigations (He and Chen, 2004, 2007a, 2007b; Zhong and Chen, 2004; He et al., 2005), we develop a novel auto-cascade absorption refrigeration system consisting of two absorbers. It is not only expected to optimize its mass and energy coupling of auto-cascade processes by intensifying its absorbing processes, but also avoiding mixing irreversibility(He et al., 2012a, 2012b). The new system is aimed at high efficient deep-freezing as low as -60℃. It uses environmental-acceptable non-azeotropic mixtures of Trifluoromethane (R23) and 1,1,1,2-Tetrafluoroethane (R134a) as a refrigerant, and N, N-Dimethylformamide (DMF) as an absorbent. In this context, theoretical and experimental analyses are undertaken on the new system and an ACAR system with single-absorber.

2. DESCRIPTIONS OF THE NEW SYSTEM

The auto-cascade absorption refrigeration system with double-absorber consists of two different circulations: refrigerant circulation and solution circulation. Firstly, the refrigerant circulation is introduced here. When heating stream is inputted into the generator, the vapor mixture of refrigerant bubbles up from strong solution (DMF as absorbent), and flows into the condenser. After the vapor is cooled at the condenser by a coolant, it goes into the separator. Then, it is separated into one vapor stream and one liquid stream, which are named as S1 and S2, respectively. The main component of S1 is a low-boiling point substance (R23) and flows out from the top of separator. Otherwise, the main component of S2 is a high-boiling point substance (R134a), and flows out from the bottom of separator. The stream S2 flows into the low-pressure side of condenser-evaporator via throttling valve 3, where S2 vaporizes and condenses S1. Finally, the stream S1 flows into the evaporator via throttling valve 1, and produces refrigeration. As to the solution circulation, the stream S1 from the evaporator enters into the absorber 1,

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2335, Page 3 and is firstly absorbed by weak solution (DMF as absorbent), which comes from throttling valve 2. After the solution becomes strong in the component of R23, it absorbs S2 from the condenser-evaporator via pump 1, and becomes strong in the component of R134a in the absorber 2. The strong solution is pumped into the high-pressure side of cycle, and is pre-heated in the solution heat exchanger by weak solution from the generator. While the pre- heated strong solution flows into the generator, the refrigeration circulation starts again. The schematic diagram of auto-cascade absorption refrigeration cycle with single-absorber (It is called as cycle 1, here), as well as auto- cascade absorption refrigeration cycle with double-absorber (It is called as cycle 2), are illustrated in Figure 1.

Figure 1 Schematics of an auto-cascade absorption refrigeration system with single-absorber and double-absorber

3. MATHEMATICAL AND PHYSICAL MODEL OF THE NEW SYSTEM

A mathematical and physical model is developed to analyze performances of the new system. The mathematical and physical model is set up on basis of energy and mass balance equations. (a) Circulating ratio of solution According to a definition of circulating ratio of solution for the traditional absorption refrigeration cycle, the circulating ratio of solution for ACAR cycle is defined as

1− ∑ xws, i Ra = i (1) ∑∑xxssi,,− wsi ii where Ra is a circulating ratio of solution; x is a mole fraction of a refrigerant component in the absorbent; the subscript i denotes the i- th component of mixture refrigerant; ss as well as ws are respectively strong solution and weak solution. (b) Generating heat load at the generator

qg = m ⋅[ h2 + ( Ra −⋅− 1) h1 Ra ⋅ h17 ] (2) where qg is the generating heat load ( W/mol ); m is the mass flow rate of refrigerant; if it is set as 1, the term of

m can be omitted in the equation ( mol/s ); h1 , h17 and h2 respectively represent a specific enthalpy of weak solution, strong solution, as well as refrigerant vapor exiting generator ( J/mol ). (c) Condensing heat load at the condenser

qhhk =2 − 3 (3) where qk is the heat discharged during the condensing process ( W/mol ); h3 represents a specific enthalpy of refrigerant exiting condenser ( J/mol ). (d) Separating process at the separator

mm=lv + m  (4)

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mh⋅=3 mvl ⋅+ h 45 m  ⋅ h (5) where m l is the mass flow rate of separated liquid refrigerant (S2) ( mol/s ); m v represents the mass flow rate of separated vapor refrigerant (S1) ( mol/s ); h4 is a specific enthalpy of S1 ( J/mol ); h5 is a specific enthalpy of S2 ( J/mol ). (e) Condensing heat load at the condenser- evaporator

qce,=⋅− mh l () 76 h (6)

h84= h − qmce v (7) where qce, is the exchanging heat load during the condensing-evaporating process ( W/mol ); h6 represents a specific enthalpy of the throttled S2, and it is equal to h5 ( J/mol ); h7 is a specific enthalpy of the throttled S2 exiting condenser-evaporator ( J/mol ); h8 is a specific enthalpy of S1 exiting condenser-evaporator ( J/mol ). (f) Refrigerating capacity at the evaporator

qev=⋅− mh ()10 h9 (8) where qe is the refrigerating capacity during the evaporating process ( W/mol ); h10 is a specific enthalpy of the throttled S1 and equals to h8 ( J/mol ); h10 represents a specific enthalpy of the throttled S1 exiting evaporator ( J/mol ). (g) Absorption process at the absorber

qa,1 =( Ra −⋅ 1) h12 + m lv ⋅ h7 −() Ra − m ⋅ h13 (9)

qa,2 =() Ra − m vv ⋅+⋅−⋅ h14 m h10 Ra h15 (10) where qa,1 and qa,2 is the discharged heat during the absorbing process ( W/mol ); h12 represents a specific

enthalpy of weak solution from the generator and equals to h11 ( J/mol ); h13 represents a specific enthalpy of

strong solution at the absorber 1 ( J/mol ); h14 is a specific enthalpy of weak solution from the absorber 1 and

equals to h13 ( J/mol ); h15 is a specific enthalpy of strong solution at the absorber 2 ( J/mol ). (h) Exchanging heat load at the solution heat exchanger

qHe =( Ra −⋅ 1) ( h1 − h 11 ) = Ra ⋅ ( h17 − h 16 ) (11)

where qHe is the exchanging heat at the solution exchanger ( W/mol ); h16 represents a specific enthalpy of strong solution pumped from the absorber 2 and equals to h15 ( J/mol ); h17 is a specific enthalpy of strong solution exiting solution heat exchanger ( J/mol ). (i) Coefficient of performance

COP= qeg q (12) where COP is a coefficient of performance. To develop the mathematical physical model, other assumptions have been made as following: (1)Both the generating process and the absorbing process start/end at an equilibrium state; (2)There is no pressure drop by friction in the cycle; (3)The composition of solution is not affected by throttling; (4)The absorption process is an isothermal process; (5)The refrigeration is achieved by J-T throttling; (6)The power consumed by pump is negligible. It is necessary for the accurate calculation of the mathematic physical model to obtain accurate thermodynamic properties of working fluids, including refrigerant and solution. It is required to obtain accurate equilibrium data of working fluids; however, there are no available VLE data of such working fluids up to now. As a result, the thermodynamic properties of working fluids are estimated by an equation of state (EOS). There are a lot of different formulations of EOS, such as a class of cubic equation of state. A cubic equation of state is cubic in volume and thus

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2335, Page 5 very easy to handle, for example, a very popular P-R equation of state (Peng and Robinson, 1976). The P-R equation of state is chosen in this text. It has the following form RT a P = − (13) v− b vv( ++ 2) b b2 where ‘R’ is the universal gas constant, ‘a’ is a function of temperature, and b as well as c are constants about critical temperature ( Tc ) and critical pressure of substance ( Pc ). Under the constraints of thermodynamics, the constants ‘a’, b and c have the following form 0.457235RT22 2 = c +−0.5 (14) a 11fTω ( r ) pc Where 2 fω =+−0.37464 1.54226ωω 0.26992 (15) where ‘ω’ is an acentric factor of a substance, ‘Tr ’ is the relative temperature. 0.077796RT b = c (16) pc To calculate thermodynamic properties of mixture refrigerant, the constants ‘a’ and ‘ b ’ in the equation of state should be replaced by the mixture constants ‘ am ’ and ‘ bm ’, which are as follows

am = Σ Σ xi x j aij (17) i j

bm = Σ Σ xi x jbij (18) i j Where

aij = ai a j (1− kij ) (19)

bi + b j b = (20) ij 2

The binary interaction coefficient kij in equation (19) is generally evaluated from experimental data of binary VLE.

The subscript i or j denote the component i or j of working fluids. The kij between the same substance is 0,

and it is -0.077 between R23 and R134a. Besides, kij is also set 0 between the absorbent and R23 or R134a.

4. THEORETICAL ANALYSES

Figure 2 COP comparisons between cycle 1 and 2 Figure 3 Cycle 2’s relationships (1) between COP and generating pressure

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Under the same working conditions, the same generating temperatures, condensing temperatures, evaporating temperatures and absorption temperatures, COP comparisons are illustrated in Fig. 2 between cycle 1 and cycle 2. , The results show COP of Cycle 2 is more 15~35% than that of cycle 1. The theoretical analysis and experiments show that the coupling between stream 1 and stream 2 at the condenser-evaporator has key effects on the performance of cycle [25~28]. The insufficient of cooling capacity of stream 2 is the greatest constraint to improvement of COP. However, it becomes possible to increase the cooling capacity of stream at the condenser- evaporator in the ACAR cycle with double-absorber. It become possible that the refrigerant streams are absorbed under different absorption pressures in the cycle 2, and the increase of absorption pressure (evaporating pressure) results in an increase of refrigeration capacity. Therefore, more mass of stream 1 can be condensed by stream 2 at the condenser-evaporator. The refrigeration capacity of cycle also increases, and a greater COP can be obtained. The relationships between COP and generating pressures are illustrated in Fig.3~5 for cycle 2. The results show that there is an optimal generating pressure under different working conditions. The main reasons are that the generating pressure affects not only the components of refrigerants but also its state before throttling. These two factors will produce great effects on the energy and mass coupling at the auto-cascade unit. The final results is that there is an optimal generating pressure to achieve a biggest COP of the new system.

Figure 4 Cycle 2’s relationships (2) between COP Figure 5 Cycle 2’s relationships (3) between COP and generating pressure and generating pressure

5. EXPERIMENTAL RESULTS

Table 1 Comparison on performances between Cycle 1 and 2 (Generating Temperature) Experimental parameters Cycle 1 Cycle 2 R23 Mole ratio 0.30 0.30 Generating Temperature (℃) 150.3 167.7 182.7 151.4 171.4 184.4 Condensing Temperature (℃) 21.3 19.4 21.5 19.8 19.3 18.8 27.0① 28.1① 26.4① Absorption Temperature (℃) 27.7 31.2 27.3 30.6② 29.2② 28.5② Evaporating Temperature (℃) -44.3 -52.5 -57.4 -49.8 -59.5 -62.3 Generating Pressure (kPa) 1080 1110 1140 1060 1100 1130 170① 150① 140① Absorption Pressure (kPa) 180 170 160 180② 160② 150② COP 0.011 0.015 0.018 0.013 0.020 0.023 （Denoting: ①and ② are respectivel y absorption pressures at absorption process of the low pressure and that of the high pressure for cycle 2）

When mole ratio of R23:R134a equals 0.3:0.7, the experimental comparisons are illustrated in Tab.1 between cycle 1 and cycle 2. Under the different working conditions, the lowest refrigeration temperatures ranges from -44.3 ~ - 57.4℃ for cycle 2 and can achieve a COP of 0.011~0.018. These results are similar with that reported by He and Chen (2005). However, for cycle 2, the better performances are achieved under the similar working conditions with that of cycle 1, for example, the lower refrigeration temperatures and greater COP. The refrigeration temperatures of

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2335, Page 7 cycle 2 is about lower 5~8℃ than those of cycle 1 under the same working conditions, and its COP is about greater 18~27.8% than that of cycle 1. The experimental results agree with theoretical one well. Under the same working conditions, the falling curves of refrigeration temperatures are illustrated in Fig.6 and 7 to compare performances for both cycle 1 and cycle 2. In addition, the lowest refrigeration temperature of -62.3℃ is achieved for cycle 2 with a COP of 0.023 under a generation temperature of 184.4℃, condensing temperature of 18.8℃, and absorption temperatures of 26.4℃ at the low-pressure absorber, and 28.5℃ at the high-pressure absorber.

Figure 6 Cycle 1’s falling curve of cooling temperatures Figure 7 Cycle 2’s falling curve of cooling temperatures

6. CONCLUSIONS

It is an interesting object for an absorption refrigeration system to achieve low refrigeration temperatures. A novel absorption refrigeration system is proposed, and theoretical and experimental investigations are conducted in context. Some interesting results are achieved, i.e. a lower refrigeration temperature and higher COP. In summary, the follows are concluded: (1) The novel system shows better performances by theoretical and experimental investigation conducted in this context; (2) It shows feasibility that an ACAR system with double-absorber achieves low refrigeration temperatures by utilizing low-grade thermal energy; (3) It is effective to optimize the coupling of energy and mass at auto-cascade unit; (4) It becomes important and necessary to find new working fluids which can improve performances of ACAR systems effectively.

ACKNOWLEDGEMENT

This work is supported by the Fundamental Research Funds for the Central Universities (Project No. 2010QNA40 78), as well as supported by Major Program of National Natural Science Foundation of China (Project No. 50890184) and National Basic Research Program of China (Project No. 2010CB227304).

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