The document evaluates the performance of various refrigerant mixtures in vapor compression refrigeration cycles based on exergy analysis. It considers mixtures including R-404A, R-410A, R-410B, R-507, R-401B, R-401C, R-402A, R-406A, R-408A, and R-409A. Thermophysical property calculations are presented for determining enthalpy and entropy values using equations of state. The refrigeration cycle is modeled and exergy efficiency is calculated based on the cooling load exergy and total exergy losses, which include compression, expansion and heat transfer losses. Performance comparisons of the mixtures are made through exergy efficiency factors and Grassmann
PYReco present at Cologne Expo Cologne 14 16th Feb 2012 = Final
Evaluating the efficiency of refrigerant mixtures in vapor compression systems
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. Paignigiannis
School 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 2005
Abstract
The use of many common refrigerants is under restriction or phase out because of their high ODP (ozone
depletion potential) or GWP (global warming potential). The regulations on environmentally acceptable
substances 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 low
ODP.
We are focused on performance comparisons of these working fluids in vapor compression refrigerating
cycles. 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 or
presented 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. 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 factors
1. Introduction
In this paper, different refrigerant mixtures have been chosen in order to observe their use in
vapor compression refrigerating cycles. Table 1 indicates their composition and the corresponding
values for ozone depletion potential and global warming potential [1,2].
Refrigerating cycle modelling is very sensitive to the successful choice of the thermophysical
refrigerant 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 presented
thermophysical property calculations in Refs. [10–12].
3. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2789
Table 1
Composition of the mixtures under consideration
ASHRAE No. Composition (% mass f.) NBPa (°C) (bubble/glide) GWPb (CO2 = 1) Replacing
R-404A R-125/143a/134a (44/52/4) (À46.5/0.8) 3700 R-502, R-22
R-410A R-32/125 (50/50) (À52.7/0.1) 1900 R-22
R-410B R-32/125 (45/55) (À51.8/0.1) 2000 R-22
R-507 R-125/143a (50/50) À46.7 (azeo) 3800 R-502
ASHRAE 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-500
R-401C R-22/152a/124 (33/15/52) (À28, 3/4, 7) 0.030 850 R-12
R-402A R-125/290/22 (60/2/38) (À48, 9/2, 0) 0.021 2600 R-502
R-406A R-22/600a/142b (55/4/41) (À36, 0/9, 9) 0.057 1800 R-12
R-408A R-125/143a/22 (7/46/47) (À44, 4/0, 7) 0.026 3000 R-502
R-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 the
sources. We have tried to predict the necessary enthalpy and entropy values with equations that
are being proposed by thermodynamics theory [3,4] and appropriate fitting to tables given by
Refs. [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 enthalpy
and 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 BÞiþ1
The 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 ð2Þ
where 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 of
crucial 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 based
on 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. 2790 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802
Table 2a
Martin–Hou equation of state
Refrigerant R-406A [13] R-404A [5]
A1 0 À12.3651966130
A2 À0.00143183269 9.9197094930
A3 1.21069446EÀ06 À3.0486302819
A4 4.76364975EÀ09 0
A5 À6.44050254EÀ12 0
B1 0 7.2676637470
B2 1.1467238EÀ06 À6.3378844502
B3 À2.46784358EÀ09 0
B4 0 9.3213426646
B5 2.14340749EÀ15 0
C1 0 À11.815938601
C2 À0.02994404977 À127.25315779
C3 6.271493992EÀ05 0
C4 0 À1108.5540675
C5 0 0
X 0.0009252221 3.8644416
K 5.475 5.475
B 8.14EÀ04 0
Tc (K) 387.64 344.7
pc (bar) 45.813 37.46
qc (kg/m3) 455.52 493
2 3
Ideal 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 stage
refrigerating system. As is well known, the problem of refrigeration is to reduce the temperature
of 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,
9. ð5Þ
To
and EU is the exergy losses.
The fluxes Qo ; E Qo and EU are graphed in Fig. 1(b).
10. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2791
Table 2b
R-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 i¼1
where 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 cj
ai ¼ ½1 þ ji ð1 À T 0:5 ÞŠ2 , ji ¼ j0i þ j1i ½ð1 þ T 0:5 Þð0:7 À T ri ÞŠ, (ji = j0i for Tr 0.7),
ri ri
j0i ¼ 0:378893 þ 1:4897153xi À 0:17131848x2 þ 0:0196554x3 , j1i = adjustable parameter for component i and Tri =
i i
Ti/Tci for component i.
Component Tci pci xi j1i xi
1-6 HFC-125 (i = 1) 339.19 3595.0 0.3023 0.0310 0.35782
HFC-143a (i = 2) 346.25 3758.1 0.2529 0.0450 0.60392
HFC-134a (i = 3) 374.2 4056.0 0.3266 À0.0060 0.03826
k11 = 0.00000 k12 = À0.0111 k13 = À0.0024
k21 = À0.0111 k22 = 0.0000 k23 = 0.0013
k31 = À0.0024 k32 = 0.0013 k33 = 0.0000
Ideal gas mixture heat capacity
X
3
cid ¼
Pmixt xi cidi
P
i¼1
cid :
Pi ideal gas heat capacity for each component
cid ¼ 4:184ðAi þ Bi T þ C i T 2 þ Di T 3 Þ
Pi
A1 = 1.170144E+01 B1 = 0.216411EÀ01 C1 = 8.685258EÀ05
A2 = 1.372849E+00 B2 = 0.750717EÀ01 C2 = À6.206979EÀ05
A3 = 4.636855E+00 B3 = 0.617904EÀ01 C3 = À3.099070EÀ05
D1 = À1.127756EÀ07
D2 = 2.011233EÀ08
D3 = 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 ð6Þ
Exergy losses due to the compressor motor (air cooled compressor) may be included for better
accuracy. These losses are calculated as follows:
11. 2792 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802
Fig. 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
nmotor
So 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 ð11Þ
Finally, 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,
12. A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802 2793
superheating 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 computational
techniques and deduce predictions for favourable alternative media.
Most of the mixtures that we have taken into consideration are non-azeotropic. This means
that the pressure–temperature relationship for the saturated liquid stated condition is different
from 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 being
proposed in the literature and practical applications.
3. General computer simulation results and comparisons
3.1. 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 is
equal 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 pressure
drops are equal to 0.2 bar. And the temperature of the cold space is 2 °C higher than that of the
evaporation temperature ðT Ã Þ.
o
In our first attempt and for simplicity, we will assume that there is neither sub-cooling nor
superheating of the suction vapor.
Two groups of exergy efficiency plots are given. In the first group (Figs. 2 and 3), we have
the variation with condensing temperature (25–60 °C) for a constant evaporating temperature
equal to À20 °C. The other group (Figs. 4 and 5) presents the influence of the evaporating
temperature 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
Tu
According to theory [14], the term To
Qo is called the cooling load anergy flux.
13. 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 environmentally
friendly refrigerant mixtures that, according to the international notion, can be substituted for
the 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 cold
room. Consequently, f depends proportionally only on the coefficient of performance (COP)
value.
14. 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 of
condensing and evaporating temperatures.
An increase in the condensation temperature results in a decrease of COP as the work needed
for 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 condensation
temperature of 25 °C in Fig. 3), while the minimum value of exergy efficiency is 17.29% (observed
for 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 is
very close to the overall maximum exergy efficiency of 51.11% (50.22% and 50.37% for R-502 and
R-507, respectively, at a condensation temperature of 25 °C, Fig. 3).
Additionally, the difference between the maximum and minimum values of exergy efficiency in
every 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) is
observed 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 perfect
match of the diagrams for the ‘‘pairs’’ R-507/R-502 (Fig. 3) and R-406A/R-12 (Fig. 2).
15. 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
o
value of the fraction T u is no longer constant, and the exergy efficiency depends not only on
To
the COP value but also on the value of T u . According to theory [3,14], for a constant conden-
To
sation 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
To
conclusively.
16. 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 of
condensing and evaporating temperatures.
The plotted lines have a non-symmetrical ‘‘bell shaped’’ form. The minimum values of exergy
efficiency appear either at an evaporation temperature of À5 °C or À40 °C. All lines have an
overall maximum point. Most of the time, this overall maximum point appears at an evaporation
temperature 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 about
À25 °C.
Taking into account the previously mentioned assumptions, the maximum exergy efficiency is
44.72% (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-409A
at an evaporation temperature of À40 °C, Fig. 5). Therefore, we note that the fluctuation between
minimum and maximum values of exergy efficiency is fairly reduced now (constant condensation
temperature) 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 the
mixture R-401C).
It should be emphasized that the exergy losses of the classical refrigerants (R-12, R-22 and
R-502) are significantly smaller.
17. 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 that
of its substitute, R-507 (Fig. 5). The smallest divergence is less than 1%, while the largest, up to
10%, 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 exergy
efficiency 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-406A
and R-408A.
18. 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 is
that 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. More
specifically, 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 mixture
R-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 of
exergy efficiency stems from the fact that compared to all the other refrigerant mixtures, it has
the 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 whenever
exergy analysis is used, a specific exergy loss corresponding to a minimum operational cost is
being 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.
19. 2800 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802
4. 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 we
will try to evaluate the influence of a variety of parameters. Below, the changes produced in the
parameters 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).
20. 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 factor
and 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 and
R-507). Comments on the design procedure as a function of the thermodynamic characteristics of
the different refrigerants are given. It is recognized that there will likely not be any major universal
substitutes. Some fluids may be better suited for certain applications than others. The Parties to
the Montreal Protocol have taken decisive action to address the growing problem of ozone layer
depletion by enhancing the control provisions of the Protocol [15–17]. While an accelerated phase
out of controlled substances is technologically feasible for the majority of applications using con-
trolled substances, there are some important applications for which acceptable alternatives have
not 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, my
co-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 one
can see and comprehend how a diversity of working fluids interact with the modification of certain
system parameters.
Appendix A
Parameters and exergy in refrigerators are given in Tables 3 and 4.
Table 3
R-404A exergy behavior
Exergy losses Initial (a) 15 K (b) 0.3 bar (c) 80% (d) Sub-cooling 5 K (e) Suction of
conditions superheated vapor 5 K
Compression (%) 36 36 36.6 32 36 35.6
Condensation (%) 12.3 12.6 12.4 12.7 12.6 12.9
Throttling (%) 6.4 2.9 6.4 6.9 4.7 6.3
Evaporation (%) 7.5 8.5 7.4 8.1 6.7 7.3
f (%) 37.8 40 37.2 40.3 40 37.9
Initial conditions: Pressure drop in evaporator and condenser equivalent to 10 K, pressure drops in suction, discharge
and 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) Pressure
drops in suction, discharge and liquid line 0.3 bar. (c) Isentropic compression efficiency 80%.
21. 2802 A. Stegou-Sagia, N. Paignigiannis / Energy Conversion and Management 46 (2005) 2787–2802
Table 4
R-406A exergy behavior
Exergy losses Initial (a) 15 K (b) 0.3 bar (c) 80% (d) Sub-cooling 5 K (e) Suction of
conditions superheated vapor 5 K
Compression (%) 37.1 37.3 38.3 33.1 38.6 36.7
Condensation (%) 13.5 10 13.8 13.8 10.4 14.1
Throttling (%) 3.4 1.3 3.3 3.6 2.2 3.3
Evaporation (%) 5.7 9.4 5.5 6.1 5.4 5.5
f (%) 40.3 42 39.1 42 43.4 40.4
Initial conditions: Pressure drop in evaporator and condenser equivalent to 10 K, pressure drops in suction, discharge
and 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) Pressure
drops 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.