Energy Conversion and Management 46 (2005) 2787–2802                                                                      ...
2788       A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802  Nomenclature  COP    ...
A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802        2789Table 1Composition of ...
2790          A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802Table 2aMartin–Hou e...
Tu
E Qo ¼      À 1 Á
Qo
ð5Þ             To       and EU is the exergy losses.                                   The fluxes Qo ; E Qo and EU are gra...
A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802                     2791Table 2bR...
2792        A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802Fig. 1. (a) Pressure–e...
A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802   2793superheating etc.), not onl...
2794                                         A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005...
A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802                                  ...
2796                                                        A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Mana...
A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802                                  ...
2798      A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802  The only exemption is ...
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Evaluation Of Mixtures Efficiency In Refrigerating Systems

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Evaluation Of Mixtures Efficiency In Refrigerating Systems

  1. 1. Energy Conversion and Management 46 (2005) 2787–2802 www.elsevier.com/locate/enconman Evaluation of mixtures efficiency in refrigerating systems A. Stegou-Sagia *, N. PaignigiannisSchool of Mechanical Engineering, Thermal Section, National Technical University of Athens, 9 Iroon Polytechniou Str., Zografou 15780, Athens, Greece Received 25 February 2004; received in revised form 29 July 2004; accepted 17 January 2005 Available online 3 March 2005Abstract The use of many common refrigerants is under restriction or phase out because of their high ODP (ozonedepletion potential) or GWP (global warming potential). The regulations on environmentally acceptablesubstances are different from country to country and are subject to frequent updates. In our article, the fol-lowing mixtures are under consideration: R-401B, R-401C, R-402A, R-404A, R-406A, R-408A, R-409A,R-410A, R-410B and R-507. Some of them do not have zero ODP, but they are in use due to their lowODP. We are focused on performance comparisons of these working fluids in vapor compression refrigeratingcycles. Our effort was conducted on the basis of exergy aspects. Various parameters of the cycles were chan-ged within a suitable range, and the results obtained were plotted in graphs of exergy efficiency factors orpresented in Grassmann diagrams and tables.Ó 2005 Elsevier Ltd. All rights reserved.Keywords: Vapor compression refrigerating systems; Refrigerant mixtures; Exergy * Corresponding author. Tel.: +30 210 7721255; fax: +30 210 7723976. E-mail address: asagia@central.ntua.gr (A. Stegou-Sagia).0196-8904/$ - see front matter Ó 2005 Elsevier Ltd. All rights reserved.doi:10.1016/j.enconman.2005.01.007
  2. 2. 2788 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 Nomenclature COP coefficient of performance cid Pmixt ideal gas mixture heat capacity E Qo cooling load exergy flux EU exergy losses GWP global warming potential m refrigerant mass flow rate nmotor compressor motor efficiency ODP ozone depletion potential p pressure pc critical pressure p0 reference pressure pr p/pc P power Qo cooling load R universal gas constant s entropy S0 reference entropy T temperature Tc critical temperature To temperature of cold space TÃ o evaporation temperature Tr T/Tc Tu ambient temperature V volume Vc critical volume Vr V/Vc q density qc critical density f exergy efficiency factors1. Introduction In this paper, different refrigerant mixtures have been chosen in order to observe their use invapor compression refrigerating cycles. Table 1 indicates their composition and the correspondingvalues for ozone depletion potential and global warming potential [1,2]. Refrigerating cycle modelling is very sensitive to the successful choice of the thermophysicalrefrigerant properties. In the literature, we have encountered various aspects such as: thermody-namic formulations [3,4], tables and equations for PVT data [5,6], ASHRAE information [7],NIST database [8] and Coolpack software [9]. In previous articles, the first author has presentedthermophysical property calculations in Refs. [10–12].
  3. 3. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2789Table 1Composition of the mixtures under considerationASHRAE No. Composition (% mass f.) NBPa (°C) (bubble/glide) GWPb (CO2 = 1) ReplacingR-404A R-125/143a/134a (44/52/4) (À46.5/0.8) 3700 R-502, R-22R-410A R-32/125 (50/50) (À52.7/0.1) 1900 R-22R-410B R-32/125 (45/55) (À51.8/0.1) 2000 R-22R-507 R-125/143a (50/50) À46.7 (azeo) 3800 R-502ASHRAE No. Composition (% mass f.) NBPa (°C) ODP (R-11 = 1) GWPb (CO2 = 1) Replacing (bubble/glide)R-401B R-22/152a/124 (61/11/28) (À34, 6/5, 9) 0.040 1200 R-12, R-500R-401C R-22/152a/124 (33/15/52) (À28, 3/4, 7) 0.030 850 R-12R-402A R-125/290/22 (60/2/38) (À48, 9/2, 0) 0.021 2600 R-502R-406A R-22/600a/142b (55/4/41) (À36, 0/9, 9) 0.057 1800 R-12R-408A R-125/143a/22 (7/46/47) (À44, 4/0, 7) 0.026 3000 R-502R-409A R-22/124/142b (60/25/15) (À34, 3/8, 5) 0.048 1400 R-12 a Boiling point or (bubble point/temperature glide) at 1 atm. Temperature glide: (Tdew À Tbubble). b ITH = 100 years. Consistent property values for this work have been deduced after careful treatment of thesources. We have tried to predict the necessary enthalpy and entropy values with equations thatare being proposed by thermodynamics theory [3,4] and appropriate fitting to tables given byRefs. [5,6]. An example of our methodology in the superheated vapor region is given for the mix-tures R-406A and R-404A. The Martin–Hou equation of state introduced by Refs. [5,13] is a selection for usage in enthalpyand entropy calculations: XT r X ðAi þ Bi T r þ C i eðÀKT r Þ Þ pr ¼ þ ð1Þ V r À B i¼1;5 ðV r À BÞiþ1The relevant coefficients are given in Table 2a. There is another equation of state proposed by Ref. [6] for R-404A (Peng–Robinson–Stryjek–Vera): p ¼ RT =ðV À bÞ À a=ðV 2 þ 2bV À b2 Þ ð2Þwhere p is in kPa, T is in K, V is in m3/mole and R = 0.008314 kJ/(mole) (K). More details for the coefficients are tabulated in Table 2b. The type of equation of state is ofcrucial importance; for example, the entropy values are deduced by [3] Z T id Z q # cPmixt qRT R 1 op s ¼ S0 þ dT À R ln þ À dq ð3Þ T0 T p0 0 q q2 oT q Emphasis must be noted that in a previous article [12], enthalpy and entropy correlations basedon the Peng–Robinson equation of state has been presented. A detailed description of our equa-tions on thermophysical property formulations will be included in a forthcoming paper.
  4. 4. 2790 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802Table 2aMartin–Hou equation of stateRefrigerant R-406A [13] R-404A [5]A1 0 À12.3651966130A2 À0.00143183269 9.9197094930A3 1.21069446EÀ06 À3.0486302819A4 4.76364975EÀ09 0A5 À6.44050254EÀ12 0B1 0 7.2676637470B2 1.1467238EÀ06 À6.3378844502B3 À2.46784358EÀ09 0B4 0 9.3213426646B5 2.14340749EÀ15 0C1 0 À11.815938601C2 À0.02994404977 À127.25315779C3 6.271493992EÀ05 0C4 0 À1108.5540675C5 0 0X 0.0009252221 3.8644416K 5.475 5.475B 8.14EÀ04 0Tc (K) 387.64 344.7pc (bar) 45.813 37.46qc (kg/m3) 455.52 493 2 3Ideal gas mixture heat capacity: cid id Pmixt ¼ A þ BT þ CT þ DT , T in K, cPmixt in kJ/kg K. R-406A [13]: A = 0.2026644,B = 0.2187572EÀ02, C = 0.008479702EÀ04, D = 0.0003858637EÀ06. R-404A [5]: A = À2.508661, B = 0.03347197,C = À1.1602EÀ04, D = 1.399253EÀ07.2. Basics from thermodynamics Fig. 1(a) illustrates a typical shape of the vapor compression cycle in a common single stagerefrigerating system. As is well known, the problem of refrigeration is to reduce the temperatureof the storage space (To) below the environmental temperature (Tu). The refrigerating cycleÕs performance is expressed as the exergy efficiency factor (f), i.e. the ratio[14]: E Qo f¼ ð4Þ E Qo þ E U The term E Qo is the cooling load Qo exergy flux,
  5. 5. Tu
  6. 6. E Qo ¼ À 1 Á
  7. 7. Qo
  8. 8. ð5Þ To and EU is the exergy losses. The fluxes Qo ; E Qo and EU are graphed in Fig. 1(b).
  9. 9. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2791Table 2bR-404A (Peng–Robinson–Stryjek–Vera, PRSV) [6]p = RT/(V À b) À a/(V2 + 2bV À b2)p is in kPa, T is in K, V is in m3/mole, R = 0.008314 kJ/(mole) (K) XX 3 3 X 3 a¼ xi xj aij ; b¼ xi bi ; i¼1 j¼1 i¼1where aij = (ai aj)0.5(1 À kij), bi = 0.077796RTci/pci, xi = mole fraction of component i, xj = mole fraction of component j,ai ¼ ð0:457235R2 T 2 =pci Þai , aj ¼ ð0:457235R2 T 2 =pcj Þaj , kij = binary interaction parameter for components i and j, ci cjai ¼ ½1 þ ji ð1 À T 0:5 ފ2 , ji ¼ j0i þ j1i ½ð1 þ T 0:5 Þð0:7 À T ri ފ, (ji = j0i for Tr 0.7), ri rij0i ¼ 0:378893 þ 1:4897153xi À 0:17131848x2 þ 0:0196554x3 , j1i = adjustable parameter for component i and Tri = i iTi/Tci for component i.Component Tci pci xi j1i xi1-6 HFC-125 (i = 1) 339.19 3595.0 0.3023 0.0310 0.35782HFC-143a (i = 2) 346.25 3758.1 0.2529 0.0450 0.60392HFC-134a (i = 3) 374.2 4056.0 0.3266 À0.0060 0.03826k11 = 0.00000 k12 = À0.0111 k13 = À0.0024k21 = À0.0111 k22 = 0.0000 k23 = 0.0013k31 = À0.0024 k32 = 0.0013 k33 = 0.0000Ideal gas mixture heat capacity X 3 cid ¼ Pmixt xi cidi P i¼1cid : Pi ideal gas heat capacity for each componentcid ¼ 4:184ðAi þ Bi T þ C i T 2 þ Di T 3 Þ PiA1 = 1.170144E+01 B1 = 0.216411EÀ01 C1 = 8.685258EÀ05A2 = 1.372849E+00 B2 = 0.750717EÀ01 C2 = À6.206979EÀ05A3 = 4.636855E+00 B3 = 0.617904EÀ01 C3 = À3.099070EÀ05D1 = À1.127756EÀ07D2 = 2.011233EÀ08D3 = 0.000000E+00 In an actual refrigerating system, a number of irreversibilities occur, resulting in exergy losses.The main losses are as follows: • Compression losses depend on the absolute pressure level, the pressure ratio for a given tem- perature lift and the thermal properties of the working medium: EU 12 ¼ m T u ðs2 À s1 Þ ð6ÞExergy losses due to the compressor motor (air cooled compressor) may be included for betteraccuracy. These losses are calculated as follows:
  10. 10. 2792 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802Fig. 1. (a) Pressure–enthalpy diagram for one stage vapor compression refrigeration cycle where: 1–2: compression(suction of no superheated vapor), 3–4: desuperheating, condensation. 4: no sub-cooling. 7–5: throttling, 5–6:evaporation. Pressure drops: 2–3 discharge line, 3–4 condenser, 4–7 liquid line, 5–6 evaporation, 6–1 suction line. (b) Fluxes Qo ; E Qo and EU . 1 À nmotor EU motor ¼ P compression ð7Þ nmotorSo the total amount would be: EU compression ¼ EU 12 þ EU motor ð8Þ• Condensation and desuperheating losses are: EU cd ¼ Qcondensation þ Qdesuperheating À m T u ðs3 À s4 Þ ð9Þ• Evaporation losses are: EU evaporation ¼ m T u ðs6 À s5 Þ À E Qo À Qo ð10Þ The condenser and evaporator losses are dependent on the specified actual temperature bound-ary of the application and the heat transfer properties of the medium. • Throttling losses are: EU throttling ¼ m T u ðs5 À s7 Þ ð11ÞFinally, for the total exergy losses, we have: EU ¼ EU compression þ EU evaporation þ EU cd þ EU throttling ð12Þ The main objective of our present project is to evaluate the influence of the operational param-eters of the refrigerating system by manipulating suitable modifications of the cycle (sub-cooling,
  11. 11. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2793superheating etc.), not only on the overall exergy losses but also on the exergy losses of its com-ponents. Furthermore, we want to examine a significant number of mixtures with computationaltechniques and deduce predictions for favourable alternative media. Most of the mixtures that we have taken into consideration are non-azeotropic. This meansthat the pressure–temperature relationship for the saturated liquid stated condition is differentfrom that of the saturated vapor in the same composition [3]. The steps taken in order to find our results are:• selection of evaporatorÕs outlet state against the desired cold space temperature;• condenser inlet and outlet temperatures should be sufficient to reject heat;• liquid enthalpy at the expansion device and related property data [5–9,13] can be used to get evaporatorÕs inlet temperature. In all other points, the fluid behaves normally. The accuracy of our results is based on the accuracy of the thermophysical property calcula-tions and the choice of usual operational conditions for the refrigerating cycles, as they are beingproposed in the literature and practical applications.3. General computer simulation results and comparisons3.1. Exergy efficiency diagrams Our exergy efficiency diagrams are drawn based on the assumptions stated below: The environmental temperature (Tu) is equal to 20 °C; the isentropic compression efficiency isequal to 0.75; the compressor motor efficiency (nmotor) is equal to 1; the pressure drop in the evap-orator and condenser is equivalent to 10 K; the suction line, discharge line and liquid line pressuredrops are equal to 0.2 bar. And the temperature of the cold space is 2 °C higher than that of theevaporation temperature ðT Ã Þ. o In our first attempt and for simplicity, we will assume that there is neither sub-cooling norsuperheating of the suction vapor. Two groups of exergy efficiency plots are given. In the first group (Figs. 2 and 3), we havethe variation with condensing temperature (25–60 °C) for a constant evaporating temperatureequal to À20 °C. The other group (Figs. 4 and 5) presents the influence of the evaporatingtemperature for a range À40 °C to À5 °C, while the condensing temperature is constant(30 °C). The following equations describe more specifically the exergy efficiency (f) [14]: Tu Tu EQo To À 1 j Qo j To À 1 Qo f¼ ¼ ¼ ð13Þ EQo þ EU P P Tu f¼ À 1 COP ð14Þ To TuAccording to theory [14], the term To Qo is called the cooling load anergy flux.
  12. 12. 2794 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 Comparison of exergy efficiency between R-401B and R-12 Comparison of exergy efficiency between R-404A and R-502 o o (Evaporating temperature = -20 C ) (Evaporating temperature = -20 C ) 0.55 0.55 0.50 0.50 0.45 Exergy efficiency Exergy efficiency 0.45 R-404A R-502 R-401B R-12 0.40 0.40 0.35 0.35 0.30 0.30 0.25 0.25 0.20 0.20 0.15 25 30 35 40 45 50 55 60 65 25 30 35 40 45 50 55 60 65 Condensing temperature (oC) Condensing temperature (oC) Comparison of exergy efficiency between R-401C and R-12 Comparison of exergy efficiency between R-406A and R-12 (Evaporating temperature = -20 o C) (Evaporating temperature = -20 oC) 0.55 0.55 0.50 0.50 Exergy efficiency 0.45 Exergy efficiency 0.45 R-401C R-12 R-406A R-12 0.40 0.40 0.35 0.35 0.30 0.30 0.25 0.25 0.20 0.20 25 30 35 40 45 50 55 60 65 25 30 35 40 45 50 55 60 65 Condensing temperature (oC) Condensing temperature (oC) Comparison of exergy efficiency between R-402A and R-502 Comparison of exergy efficiency between R-408A and R-502 o (Evaporating temperature = -20 C) (Evaporating temperature = -20 o C) 0.55 0.55 0.50 0.50 0.45 0.45 Exergy efficiency Exergy efficiency R-402A R-502 R-408A R-502 0.40 0.40 0.35 0.35 0.30 0.30 0.25 0.25 0.20 0.20 0.15 0.15 25 30 35 40 45 50 55 60 65 25 30 35 40 45 50 55 60 65 Condensing temperature (oC) Condensing temperature (oC)Fig. 2. Comparison of the exergy efficiency factors for the refrigerant mixtures R-401B, R-401C, R-402A, R-404A, R-406A, R-408A and their corresponding conventional ones (evaporating temperature: À20 °C, condensing temperatures:25–60 °C). Our findings (Figs. 2–5), for comparative reasons, are illustrated for the environmentallyfriendly refrigerant mixtures that, according to the international notion, can be substituted forthe conventional ones, which are indicated by the dashed lines.3.1.1. Constant evaporation temperature (T Ã ¼ 253 K) o When the evaporation temperature is constant, the same is true for the temperature of the coldroom. Consequently, f depends proportionally only on the coefficient of performance (COP)value.
  13. 13. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2795 Comparison of exergy efficiency between R-409A and R-12 Comparison of exergy efficiency between R-410B and R-22 (Evaporating temperature = -20oC) (Evaporating temperature = -20 oC) 0.55 0.55 0.50 0.50 0.45 Exergy efficiency Exergy efficiency R-409A R-12 0.45 0.40 R-410B R-22 0.40 0.35 0.30 0.35 0.25 0.30 0.20 0.25 0.15 0.20 25 30 35 40 45 50 55 60 65 25 30 35 40 45 50 55 60 65 Condensing temperature (oC) Condensing temperature (oC) Comparison of exergy efficiency between R-410A and R-22 Comparison of exergy efficiency between R-507 and R-502 (Evaporating temperature = -20 oC) (Evaporating temperature = -20 o C) 0.55 0.55 0.50 0.50 Exergy efficiency 0.45 Exergy efficiency 0.45 R-410A R-22 R-507 R-502 0.40 0.40 0.35 0.35 0.30 0.30 0.25 0.25 0.20 0.20 25 30 35 40 45 50 55 60 65 25 30 35 40 45 50 55 60 65 Condensing temperature (oC) Condensing temperature (oC)Fig. 3. R-409A, R-410A, R-410B, R-507 and the corresponding classical refrigerants: exergy efficiency as a function ofcondensing and evaporating temperatures. An increase in the condensation temperature results in a decrease of COP as the work neededfor compression is increased. As a result, the exergy efficiency is expected to decrease. This is ex-actly what we observe in Figs. 2 and 3. The maximum exergy efficiency is 51.11% (observed for the mixture R-410B at a condensationtemperature of 25 °C in Fig. 3), while the minimum value of exergy efficiency is 17.29% (observedfor the mixture R-408A at a condensation temperature of 60 °C in Fig. 2). As far as the examined azeotropic mixtures are concerned, their maximum exergy efficiency isvery close to the overall maximum exergy efficiency of 51.11% (50.22% and 50.37% for R-502 andR-507, respectively, at a condensation temperature of 25 °C, Fig. 3). Additionally, the difference between the maximum and minimum values of exergy efficiency inevery refrigerant mixture is quite high, rising from 24.73% (for the mixture R-408A) to 31.75%(for the mixture R-404A). Comparing the exergy efficiencies of the alternative refrigerant mixtures with those of the clas-sical refrigerants they replace (R-12, R-22 and R-502), we note that the exergy losses of the clas-sical refrigerants are lower. A big divergence (always in favour of the classical refrigerants) isobserved in the ‘‘pairs’’ R-409A/R-12 (Fig. 3) and R-408A/R-502 (Fig. 2). In the case of R-409A/R-12, this divergence can even reach 10% for certain temperatures. Also, there is a perfectmatch of the diagrams for the ‘‘pairs’’ R-507/R-502 (Fig. 3) and R-406A/R-12 (Fig. 2).
  14. 14. 2796 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 Comparison of exergy efficiency between R-404A and R-502 Comparison of exergy efficiency between R-401B and R-12 o (Condensing temperature = 30 C) o (Condensing temperature = 30 C) 0.45 0.45 0.44 0.44 Exergy efficiency Exergy efficiency 0.43 0.43 0.42 0.41 0.42 R-401B R-12 0.40 R-404A R-502 0.39 0.41 0.38 0.40 -40 -35 -30 -25 -20 -15 -10 -5 0 -40 -35 -30 -25 -20 -15 -10 -5 0 Evaporating temperature (oC) Evaporating temperature (oC) Comparison of exergy efficiency between R-401C and R-12 Comparison of exergy efficiency between R-406A and R-12 o (Condensing temperature = 30 oC) (Condensing temperature = 30 C) 0.45 0.45 0.44 0.44 Exergy efficiency Exergy efficiency 0.43 0.43 0.42 0.41 0.42 0.40 0.41 0.39 0.40 0.38 R-406A R-12 R-401C R-12 0.37 0.39 0.36 0.38 -40 -35 -30 -25 -20 -15 -10 -5 0 -40 -35 -30 -25 -20 -15 -10 -5 0 Evaporating temperature (oC) Evaporating temperature (oC) Comparison of exergy efficiency between R-402A and R-502 Comparison of exergy efficiency between R-408A and R- o (Condensing temperature = 30 C) 502 (Condensing temperature = 30 oC) 0.45 0.45 0.44 0.43 Exergy efficiency Exergy efficiency 0.43 0.41 0.42 0.39 0.41 R-408A R-502 0.40 0.37 R-402A R-502 0.39 0.35 0.38 0.33 -40 -35 -30 -25 -20 -15 -10 -5 0 -40 -35 -30 -25 -20 -15 -10 -5 0 Evaporating temperature (oC) Evaporating temperature (oC)Fig. 4. Comparison of the exergy efficiency factors for the refrigerant mixtures R-401B, R-401C, R-402A, R-404A, R-406A, R-408A and their corresponding conventional ones (condensing temperature: 30 °C, evaporating temperatures:À40 to À5 °C).3.1.2. Constant condensation temperature (Tcond = 303 K) In the case of constant condensation temperature, both T Ã and To change. As a result, the ovalue of the fraction T u is no longer constant, and the exergy efficiency depends not only on Tothe COP value but also on the value of T u . According to theory [3,14], for a constant conden- Tosation temperature, a decrease in the evaporation temperature results in an increase of the frac-tion T u and a decrease of COP. Hence, we cannot predict the exact form of the diagram Toconclusively.
  15. 15. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2797 Comparison of exergy efficiency between R-409A and R-12 Comparison of exergy efficiency between R-410A and R-22 (Condensing temperature = 30 o C ) (Condensing temperature = 30oC) 0.46 0.46 0.44 0.45 Exergy efficiency Exergy efficiency 0.42 0.44 0.40 R-409A R-12 R-410A R-22 0.38 0.43 0.36 0.42 0.34 0.41 0.32 0.30 0.40 -40 -35 -30 -25 -20 -15 -10 -5 0 -40 -35 -30 -25 -20 -15 -10 -5 0 Evaporating temperature (oC) Evaporating temperature (oC) Comparison of exergy efficiency between R-410B and R-22 Comparison of exergy efficiency between R-507 and R-502 (Condensing temperature = 30 o C) (Condensing temperature = 30 o C) 0.46 0.45 0.44 0.45 Exergy efficiency Exergy efficiency 0.43 0.44 0.42 0.43 R-507 R-502 0.41 R-410B R-22 0.42 0.40 0.41 0.39 0.38 0.40 -40 -35 -30 -25 -20 -15 -10 -5 0 -40 -35 -30 -25 -20 -15 -10 -5 0 Evaporating temperature (oC) Evaporating temperature (oC)Fig. 5. R-409A, R-410A, R-410B, R-507 and the corresponding classical refrigerants: exergy efficiency as a function ofcondensing and evaporating temperatures. The plotted lines have a non-symmetrical ‘‘bell shaped’’ form. The minimum values of exergyefficiency appear either at an evaporation temperature of À5 °C or À40 °C. All lines have anoverall maximum point. Most of the time, this overall maximum point appears at an evaporationtemperature of À20 °C with the exception of R-408A (Fig. 4) and R-410A, R-410B and R-507(Fig. 5), where the overall maximum point appears at an evaporation temperature of aboutÀ25 °C. Taking into account the previously mentioned assumptions, the maximum exergy efficiency is44.72% (observed for the mixtures R-410A and R-410B at an evaporation temperature of À25 °C,Fig. 5), while the minimum value of exergy efficiency is 31.59% (observed for the mixture R-409Aat an evaporation temperature of À40 °C, Fig. 5). Therefore, we note that the fluctuation betweenminimum and maximum values of exergy efficiency is fairly reduced now (constant condensationtemperature) compared to the previous case (constant evaporation temperature). The difference between the maximum and minimum values of exergy efficiency in every refrig-erant mixture has been significantly reduced, from 2.56% (for the mixture R-404A) to 5% (for themixture R-401C). It should be emphasized that the exergy losses of the classical refrigerants (R-12, R-22 andR-502) are significantly smaller.
  16. 16. 2798 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 The only exemption is the ‘‘pair’’ R-507/R-502, where R-502Õs efficiency is slightly less than thatof its substitute, R-507 (Fig. 5). The smallest divergence is less than 1%, while the largest, up to10%, is observed in the ‘‘pair’’ R-409A/R-12.3.2. Grassman diagrams For the Grassmann plots (Figs. 6 and 7), we have used the same basic assumptions as in exergyefficiency diagrams. The evaporation temperature has been chosen equally to À20 °C, the com-pressor motor efficiency is taken as 0.85 and the cooling load equals 100 kW.Fig. 6. Grassmann diagrams depicting the exergy losses with the use of R-401B, R-401C, R-402A, R-404A, R-406Aand R-408A.
  17. 17. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2799 Fig. 7. Grassmann diagrams depicting the exergy losses with the use of R-409A, R-410A, R-410B and R-507. From the Grassmann diagrams, we note that the compression exergy losses are the most signif-icant losses, increasing from 32.6% (for the mixture R-408A) to 37.1% (for the mixture R-406A),Fig. 6. Compression exergy losses are followed by the condensation losses, increasing from 12%(for the mixture R-401C) to 20.2% (for the mixture R-408A), Fig. 6. The third highest amount isthat of evaporation, followed by the throttling losses. The only exemption is the refrigerant mix-ture R-408A, where the throttling exergy losses are higher than those of evaporation. Morespecifically, the evaporation exergy losses increase from 5.7% (for the mixture R-406A, Fig. 6)to 11.8% (R-409A, Fig. 7), while the throttling losses increase from 3.4% (for the mixtureR-406A) to 8.4% (R-408A), Fig. 6. The mixture R-406A shows the highest value of exergy efficiency (see Appendix A) of 40.3%.Although this blend has the largest compression exergy losses (Fig. 6: 37.1%), its high value ofexergy efficiency stems from the fact that compared to all the other refrigerant mixtures, it hasthe smallest throttling and evaporation exergy losses. The lowest exergy efficiency value(31.1%) belongs to the mixture R-409A. Although R-409A demonstrates the second smallest com-pression exergy losses (32.7%), its condensation and evaporation exergy losses are very high (Fig.7: 18% and 11.8%, respectively). In all cases, the exergy losses are unavoidable, since all natural processes are irreversible. Yet,cutting down these losses is a feasible task. Of course, this reduction has its price, and wheneverexergy analysis is used, a specific exergy loss corresponding to a minimum operational cost isbeing sought. By modifying some of the systemÕs parameters, we can minimise exergy losses.The way changes of system parameters affect exergy losses is analysed below.
  18. 18. 2800 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–28024. Influence of operational parameters of the refrigeration system on exergy losses We give the greatest attention to two refrigerant mixtures, the R-404A and R-406A where wewill try to evaluate the influence of a variety of parameters. Below, the changes produced in theparameters under consideration are presented. The analytical results are tabulated in Appendix A,Tables 3 and 4.• Increase of the pressure drop in the evaporator and condenser. In both refrigerant mixtures, we note that the overall exergy efficiency has increased (by 2.2% and 1.7% for R-404A and R- 406A, respectively). We also observe a significant decrease of the throttling losses and an increase of the evaporation losses in both refrigerant mixtures.• Increase of the pressure drops in the suction, discharge and liquid line from 0.2 to 0.3 bar. We have a reduction in exergy efficiency (by 0.6% and 1.2% for R-404A and R-406A, respectively). This is due to the fact that an increase in the pressure drops of the refrigeration system results in an increased power consumption in order to overcome these losses. Apart from the increase of compression exergy losses, all the other exergy losses remain almost constant.• Increase of the isentropic compression efficiency from 0.75 to 0.8. The exergy efficiency is signif- icantly increased (by 2.5% and 1.7% for R-404A and R-406A, respectively). There is a signif- icant decrease of the compression exergy losses. All the other exergy losses change slightly. In the log P À h diagram, non-isentropic compression is always located on the right of isentro- pic compression. As a result, as the isentropic efficiency decreases, the end of compression cor- responds to a higher value of enthalpy in comparison to the enthalpy value of isentropic compression, and consequently, the compression power demand is increased. For an increased isentropic efficiency, we have a decreased compression power demand. Taking this into account and the fact that the cooling load does not change (the coordinates of points 5 and 6 in Fig. 1a remain unaltered), we conclude that the COP value increases. So, we can see from the definition of exergy efficiency that the exergy efficiency increases (the value of the fraction T u is constant). To• Use of sub-cooling of 5 K. A significant increase of efficiency, by 2.2% for R-404A and 3.1% for R-406A can be observed.The condensation exergy losses are slightly increased with R-404A but decreased significantly with R-406A. Additionally, the throttling and evaporation exergy losses decrease in both refrigerant mixtures. Conclusively, the use of sub-cooling results in an increase of the overall exergy efficiency. This is perfectly explainable as, with the use of sub-cooling, point 4 (Fig. 1a), which corresponds to the end of condensation, is moved to the left on the log P À h diagram. As a result of this movement and taking into account that the cooling load remains constant at 100 kW, the reflected value of cooling capacity is increased, and therefore, the re-circulating mass flow is reduced. Also, the coordinates of points 1 and 2 (Fig. 1a) remain invariable, and consequently, the compression power demand remains constant. Taking the aforesaid into consideration, the trends for the COP and the overall exergy efficiency values are determined.• Suction of superheated vapor of 5 K. The suction of superheated vapor results in a slight increase of the overall exergy efficiency. Nevertheless, this does not constitute a rule, as the suction of superheated vapor moves point 6 (Fig. 1a), which corresponds to the end of evaporation, to the right on the log P À h diagram. Therefore, the reflected value of cooling load capacity is increased, and the re-circulating mass flow is decreased (the cooling load remains constant).
  19. 19. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2801 At the same time, the compression power demand is increased, and we do not have a prior example to see if the COP will conclusively increase or decrease. Moreover, we cannot have a general view regarding the overall exergy efficiency.5. Conclusion This work enables us to acquire an assessment of the variation of the exergy efficiency factorand exergy flow related to the replacements of R-12, R-22 and R-502 with the mixtures under con-sideration (R-401B, R-401C, R-402A, R-404A, R-406A, R-408A, R-409A, R-410A, R-410B andR-507). Comments on the design procedure as a function of the thermodynamic characteristics ofthe different refrigerants are given. It is recognized that there will likely not be any major universalsubstitutes. Some fluids may be better suited for certain applications than others. The Parties tothe Montreal Protocol have taken decisive action to address the growing problem of ozone layerdepletion by enhancing the control provisions of the Protocol [15–17]. While an accelerated phaseout of controlled substances is technologically feasible for the majority of applications using con-trolled substances, there are some important applications for which acceptable alternatives havenot yet been developed or may not be available in time for the adjusted phase out date. Further-more, the developing countries sought and were granted exemptions from the control measures. Moreover, It is essential to balance all these concerns in order to make the best alternative deci-sions possible in the phase out process of ozone depletion substances. In synopsis, although plenty of work has taken part in reaching some decisions in this area, myco-author and I had to combine state conditions, a range of sources for enthalpy and entropy val-ues as well as a number of refrigerants. We hope that our paper will attract readers because onecan see and comprehend how a diversity of working fluids interact with the modification of certainsystem parameters.Appendix A Parameters and exergy in refrigerators are given in Tables 3 and 4.Table 3R-404A exergy behaviorExergy losses Initial (a) 15 K (b) 0.3 bar (c) 80% (d) Sub-cooling 5 K (e) Suction of conditions superheated vapor 5 KCompression (%) 36 36 36.6 32 36 35.6Condensation (%) 12.3 12.6 12.4 12.7 12.6 12.9Throttling (%) 6.4 2.9 6.4 6.9 4.7 6.3Evaporation (%) 7.5 8.5 7.4 8.1 6.7 7.3f (%) 37.8 40 37.2 40.3 40 37.9Initial conditions: Pressure drop in evaporator and condenser equivalent to 10 K, pressure drops in suction, dischargeand liquid line = 0.2 bar, isentropic compression efficiency 75%, no sub-cooling, no suction of superheated vapor used.Changes produced: (a) Increase of the pressure drop in the evaporator and condenser (equivalent to 15 K). (b) Pressuredrops in suction, discharge and liquid line 0.3 bar. (c) Isentropic compression efficiency 80%.
  20. 20. 2802 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802Table 4R-406A exergy behaviorExergy losses Initial (a) 15 K (b) 0.3 bar (c) 80% (d) Sub-cooling 5 K (e) Suction of conditions superheated vapor 5 KCompression (%) 37.1 37.3 38.3 33.1 38.6 36.7Condensation (%) 13.5 10 13.8 13.8 10.4 14.1Throttling (%) 3.4 1.3 3.3 3.6 2.2 3.3Evaporation (%) 5.7 9.4 5.5 6.1 5.4 5.5f (%) 40.3 42 39.1 42 43.4 40.4Initial conditions: Pressure drop in evaporator and condenser equivalent to 10 K, pressure drops in suction, dischargeand liquid line = 0.2 bar, isentropic compression efficiency 75%, no sub-cooling, no suction of superheated vapor used.Changes produced: (a) Increase of the pressure drop in the evaporator and condenser (equivalent to 15 K). (b) Pressuredrops in suction, discharge and liquid line 0.3 bar. (c) Isentropic compression efficiency 80%.References [1] Cavallini A. In: Proc of 19th International Congress of Refrigeration, IVa. International Institute of Refrigeration, 1995. p. 25–42. [2] Calm JM, Hourahan GC. Refrigerant data summary. Eng Syst 2001;18(11):74–88. [3] Smith JM, Van Ness HC. Introduction to chemical engineering thermodynamics. 3rd ed. Chemical engineering series. New York: McGraw-Hill; 1975. [4] Perry RH, Green DW. PerryÕs chemical engineering handbook. 6th ed. Singapore: McGraw-Hill; 1984. [5] INEOS Fluor, European Refrigeration, Refrigerant properties, 2004. [6] Du Pont SUVA Refrigerants, Technical Information, 2003. [7] ASHRAE, Fundamentals handbook, New York: ASHRAE, 2001. [8] NIST Standard Reference Database 23, NIST thermodynamic and transport properties of refrigerants and refrigerant mixtures, REFPROP, Version 6.01, 1998. [9] Coolpack Software, Denmark Technical University, Department of Mechanical Engineering, 2001.[10] Stegou-Sagia A. Thermodynamic property formulations and heat transfer aspects for replacement refrigerants R123 and R134a. Int J Energy Res 1997;21:871–84.[11] Stegou-Sagia A, Damanakis M. Thermophysical property formulations for R32/R134a mixtures. Int J Appl Thermodyn 1999;2(3):139–43.[12] Stegou-Sagia A, Damanakis M. Binary and ternary blends of R134a as alternative refrigerants to R-22. Int J Energy Convers Manage 2000;41:1345–59.[13] Doering R. Thermodynamic properties of a new refrigerant R-406A, Results of experimental and theoretical investigations, Solvay Fluor und Derivate GmbH, Technical Service—Refrigerants, 1995.[14] Baehr HD. Thermodynamik. siebente auflage. Berlin, Heidelberg: Springer-Verlag; 1989.[15] Blackmore R, Reddish A. Global environmental issues. 2nd ed. London: Hodder and Stougton; 1996.[16] Ozone Secretariat. Montreal protocol on substances that deplete the ozone layer, UNEP, 2000.[17] Calm JM. Responsible responses to refrigerant regulation. Eng Syst 2003;20(16):66–72.

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