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Heat-transfer assessment of the low GWP substitutes for traditional HFC refrigerants
International Journal of Heat and Mass Transfer 139 (2019) 31–38
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Heat-transfer assessment of the low GWP substitutes for traditional HFC refrigerants Giovanni A. Longo ⇑, Giulia Righetti, Claudio Zilio University of Padova, Department of Management and Engineering, I-36100 Vicenza, Italy
a r t i c l e
i n f o
Article history: Received 21 January 2019 Received in revised form 26 April 2019 Accepted 29 April 2019 Available online 7 May 2019 Keywords: Assessment GWP Brazed Plate Heat Exchangers BPHE Performance Evaluation Criteria PEC Total Temperature Penalisation TTP
a b s t r a c t This paper presents the heat-transfer assessment of different low GWP substitutes for traditional HydroFluoroCarbon (HFC) refrigerants obtained by applying both an experimental analysis based on the direct measurement of thermal (heat transfer coefficients) and hydraulic (pressure drops) performances and a theoretical analysis based on a specific Performance Evaluation Criteria (PEC), the Total Temperature Penalisation (TTP), in the specific case of boiling and condensation in a commercial Brazed Plate Heat Exchanger (BPHE). The results of both the experimental and the theoretical assessment confirm that the refrigerant R32 has comparable heat transfer characteristics to R410A. Similarly the refrigerants R290 and R1270 exhibit superior heat transfer performance to R404A; the refrigerants R152a, R1234yf and R1234ze(E) show heat transfer characteristics comparable to R134a and R1234ze (Z) exhibits superior heat transfer performance to R236fa. The coupling of experimental and theoretical assessments provides a sound procedure for the heat-transfer assessment of the low GWP alternatives for traditional HFC refrigerants. Ó 2019 Elsevier Ltd. All rights reserved.
1. Introduction The assessment of a refrigerant must take into account several constrains including thermodynamic and thermophysical properties, environmental impact such as Ozone Depletion Potential (ODP) and Global Warming Potential (GWP), safety issues such as flammability and toxicity, chemical characteristics such as stability and compatibility with the materials, availability and cost. The first operating fluids used in refrigeration technology (1830–1930s) were in general toxic, flammable, and highly reactive and they were eliminated in 1930s by the introduction of the ChloroFluoroCarbon (CFC) and HydroChloroFluoroCarbon (HCFC) refrigerants characterised by good thermodynamic properties and chemical stability, absence of toxicity and flammability, good compatibility with material and reduced cost [1]. The phase-out of CFC and HCFC refrigerants in 1990–2010s and their progressive substitution with HydroFluoroCarbon (HFC) refrigerants was due to an environmental issue, their unacceptable ODP [2]. Nowadays traditional HFC refrigerants are subjected to a progressive reduction in use or a complete phase-out due to their high GWP according to an international protocol [3].
⇑ Corresponding author. E-mail address: [email protected] (G.A. Longo). https://doi.org/10.1016/j.ijheatmasstransfer.2019.04.144 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.
McLinden et al. in 2017 [4] investigated 184,000 small molecules for identifying the possible substitutes for traditional HFC refrigerants applying thermodynamic properties, GWP, flammability, stability, and toxicity constraints. Only 138 molecules exhibit a GWP lower than 1000, a critical temperature between 320 and 420 K, no or mild flammability, no toxicity and sufficient chemical stability. Most of the 138 molecules are relatively «new» fluids that have need of a sound assessment before the extended application in refrigeration and air conditioning. Currently there is no consolidated and comprehensive method or procedure for the heattransfer assessment of a refrigerant valid for the different heat transfer processes (single-phase, boiling, condensation) and the different type of heat exchangers (tubular, micro-channel, plate heat exchanger, etc.). This paper presents the heat-transfer assessment of different low GWP fluids that have been proposed as possible substitutes for traditional HFC refrigerants. The assessment is obtained through the application of both an experimental analysis based on the direct measurement of thermal (heat transfer coefficients) and hydraulic (pressure drops) performances and a theoretical analysis based on a specific Performance Evaluation Criteria (PEC), the Total Temperature Penalisation (TTP). This approach is applied in the specific case of refrigerant two-phase heat transfer inside Brazed Plate Heat Exchanger (BPHE) in forced convection regime.
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Nomenclature A b cp dh g G h J KE/V L m nch p p* P Pr q Q Ra Rp Reeq S s T TTP W X
area of the plate (m2) height of the corrugation (m) specific heat capacity (J kg1 K1) hydraulic diameter, dh = 2b (m) gravity acceleration (m s2) mass flux, G = m/(nch W b) (kg m2 s1) heat transfer coefficient (W m2 K1) specific enthalpy (J kg1) specific kinetic energy (J m3) flow length of the plate (m) mass flow rate (kg s1) number of channels pressure (Pa) reduced pressure corrugation pitch (m) Prandtl number, Pr = l cp/k heat flux, q = Q/S (W m2) heat flow rate (W) arithmetic mean roughness (ISO4271/1) (lm) roughness (DIN 4762/1) (lm) equivalent Reynolds number, Reeq = G [(1 X) + X (qL/ qG)1/2] dh/lL heat transfer area (m2) plate wall thickness (m) temperature (K) Total Temperature Penalisation (K) width of the plate (m) vapour quality, X = (J JL)/DJLG
Greek symbols b inclination angle of the corrugation D difference
DJLG U k
l q
n
latent heat of vaporisation/condensation (J kg1) enlargement factor thermal conductivity (W m1 K1) viscosity (kg m1 s1) density (kg m3) frictional pressure drop coefficient
Subscripts a momentum b boiling ave average c manifold and port cb convective boiling cn condensation d driving eq equivalent f frictional g gravity G vapour phase in inlet L liquid phase LG liquid gas phase change m average value nb nucleate boiling out outlet r refrigerant s saturation sf secondary fluid t total w water
2. Experimental assessment The experimental heat-transfer assessment of a refrigerant involves the combined evaluation of thermal and hydraulic performances in a specific application. The authors of the present paper had carried out in the past an extensive analysis of refrigerants condensation and vaporisation inside a commercial Brazed Plate Heat Exchanger (BPHE) [5,6]. The traditional HFC refrigerants (R134a, R410A, R404A, R236fa), HydroCarbon (HC) refrigerants (R600a, R290, R1270), low GWP HFC refrigerants (R152a, R32), HydroFluoroOlefin (HFO) refrigerants (R1234yf, R1234ze(E), R1234ze(Z)) and also a HydroChloroFluoroOlefin (HCFO) refrigerant (R1233zd(E)) were considered. The tested BPHE consists of 10 plates, 72 mm in width and 310 mm in length, which present a macro-scale herringbone corrugation with an inclination angle of 65° and a corrugation amplitude of 2 mm: Fig. 1 and Table 1 give the main geometrical characteristics of the tested BPHE. The thermal and hydraulic performances of a refrigerant may be evaluated by considering two temperature differences: the saturation temperature drop DTs.r, due to the total refrigerant pressure drop Dpt, and the driving temperature difference DTd.r, inverse to the heat transfer coefficient hr. Both DTs.r and DTd.r reduce the overall refrigeration system energetic and exergetic efficiencies by increasing the required compressor power for achieving the same refrigeration effect. The saturation temperature drop DTs.r, equal to the difference between the saturation temperature of the refrigerant at the inlet
Fig. 1. Schematic view of a plate of the commercial BPHE.
G.A. Longo et al. / International Journal of Heat and Mass Transfer 139 (2019) 31–38
The driving temperature difference DTd.r, equal to the average difference between the wall temperature and the saturation temperature, is related to the refrigerant heat transfer coefficient hr
Table 1 Geometrical characteristics of the commercial BPHE. Parameter
Measure/Type
Fluid flow plate length L (mm) Plate width W (mm) Area of the plate A (m2) Corrugation type Angle of the corrugation (°) Corrugation amplitude b (mm) Corrugation pitch P (mm) Plate roughness Ra (m) Plate roughness Rp (m) Number of plates Channels on refrigerant side nch.r Channels on water side nch.w
278.0 72.0 0.020 Herringbone 65 2.0 8.0 0.4 1.0 10 4 5
33
DT d:r ¼ T w:ave T s:ave ¼ ðmr JLG Þ=ðShr Þ ¼ q=hr
and at the outlet of the heat exchanger, is related to the total refrigerant pressure drop Dpt
DT s:r ¼ T s ðpin Þ T s ðpout Þ
ð1Þ
pout ¼ pin Dpt
ð2Þ
Dpt ¼ Dpf þ Dpa þ Dpg þ Dpc
ð3Þ
where Dpf is the frictional pressure drop, Dpa the momentum pressure drop, Dpg the gravity pressure drop and Dpc the manifold and ports local pressure drop. The condensation and boiling total pressure drops enhance the pressure difference between the evaporator and the condenser increasing the compression work. In addition, they cause a reduction of the saturation temperature in the evaporator and the condenser with negative effects on the driving temperature differences. Therefore condensation and boiling total pressure drops produce a double negative effect on the global performance of a refrigeration system.
ð4Þ
where mr is the refrigerant mass flow rate, DJLG the refrigerant latent heat of vaporisation/condensation, S the heat transfer area and q the heat flux. High condensation and boiling heat transfer coefficients involve lower driving temperature differences with positive effects on the global performance of a refrigeration system. Fig. 2 shows the temperature profiles in a boiling process inside a counter-flow heat exchanger highlighting the two characteristic temperature differences DTs.r and DTd.r. Therefore, in order to evaluate experimentally the thermal and hydraulic performances of the different refrigerants in two phase heat transfer inside a commercial BPHE the experimental data points are plotted showing the driving temperature difference DTd.r, inverse to the refrigerant heat transfer coefficients, versus the saturation temperature drops DTs.r caused by the total pressure drops. The more effective refrigerants appear in the lower left portion of the diagram (higher heat transfer coefficients and lower saturation temperature drops), while the thermal and hydraulic performances decreases moving toward the higher right portion of the diagram (lower heat transfer coefficients and higher saturation temperature drops). Fig. 3 shows the driving temperature difference DTd.r versus the saturation temperature drops DTs.r measured in the boiling tests at 20 °C of saturation temperature inside the commercial BPHE illustrated in Fig. 1 and Table 1 with 13 different refrigerants. The boiling tests were carried out under similar operating conditions for all the refrigerants tested with an inlet vapour quality around 0.2–0.3, an outlet vapour quality around 0.8–1.0, an heat flux in the range 2–23 kW m2 and a refrigerant mass velocity in the range 5–40 kg m2 s1. The refrigerant boiling inside the tested BPHE was controlled by the combined effects of nucleate boiling and forced convection boiling. In general, nucleate boiling is the dominant heat transfer regime for
Fig. 2. Temperature profiles in a boiling process inside a counter-flow heat exchanger.
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Fig. 3. Driving temperature difference vs. saturation temperature drop in boiling inside a commercial BPHE. (Experimental Data).
high and medium pressure refrigerants, while forced convection boiling is the dominant heat transfer regime for low pressure refrigerants. In Fig. 3, it is easy to distinguish the high pressure refrigerants (round symbol), from the medium pressure refrigerants (triangular symbol) and the low pressure refrigerants (quadratic symbol). Within the high pressure fluids, the traditional refrigerant R404A and R410A exhibit boiling performances slightly higher than those of their possible low GWP substitutes R32, R290 and R1270. Considering the medium pressure fluids, the traditional refrigerant R134a shows boiling performances very similar to the low GWP substitute R1234yf, slightly higher than those of the low GWP substitute R152a and slightly lower than those of the low GWP substitute R1234ze(E). Finally, taking into account the low pressure fluids, the traditional refrigerant R236fa exhibits
boiling performances higher than those of all its low GWP substitutes R600a, R1234ze(Z) and R1233zd(E). Fig. 4 shows the driving temperature difference DTd.r, versus the saturation temperature drops DTs.r measured during the condensation tests at 30 °C of saturation temperature inside the commercial BPHE illustrated in Fig. 1 and Table 1 with 12 different refrigerants in forced convection condensation regime (Reeq > 1600). The condensation tests were carried out under similar operating conditions for all the refrigerants tested with an inlet vapour quality around 0.9–1.0, an outlet vapour quality around 0.0–0.1, an heat flux in the range 9–34 kW m2 and a refrigerant mass velocity in the range 15–43 kg m-2s1. The refrigerant condensation inside the BPHE tested was gravity controlled at low refrigerant mass velocity (G < 15–20 kg m2 s1), while forced convection conden-
Fig. 4. Driving temperature difference vs. saturation temperature drop in forced convection condensation inside a commercial BPHE. (Experimental Data).
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sation occurs at higher refrigerant mass velocity (G > 15– 20 kg m2 s1). The transition point from gravity controlled to forced convection condensation was found for an equivalent Reynolds number Reeq around 1600 for all the refrigerants tested. Also in this case it is easy to distinguish the high pressure refrigerants (round symbol), from the medium pressure refrigerants (triangular symbol) and the low pressure refrigerants (quadratic symbol). The low GWP refrigerant R32 exhibits condensation performances very similar to those of the traditional refrigerant R410A, while the low GWP refrigerants R290 and R1270 shows condensation performances very similar to those of the traditional refrigerant R404A. Refrigerant R134a exhibits condensation performances very similar to those of R1234yf, slightly higher than those of R152a and slightly lower than those of R1234ze(E). Refrigerant R1234ze(Z) exhibits condensation performances higher both than those of the traditional refrigerant R236fa and also of the alternative low GWP substitute R600a. The results of this experimental assessment confirm that the refrigerant R32 has comparable heat transfer characteristics to R410A, while the refrigerants R290 and R1270 exhibit superior heat transfer performance to R404A; refrigerants R152a, R1234yf and R1234ze(E) show heat transfer characteristics comparable to R134 and R1234ze(Z) exhibits superior heat transfer performance to R236fa.
3. Theoretical assessment The theoretical heat-transfer ranking of refrigerants requires the coupled analysis of thermal (heat transfer coefficients) and hydraulic (pressure drops) performances in a specific operating case: Cavallini et al. [7] carried out an exhaustive analysis on the Performance Evaluation Criteria (PEC) for refrigerant two-phase heat transfer inside tubes. In 2013 Brown et al. [8] presented the combined thermal and hydraulic analysis of refrigerant flow boiling inside plain round tubes by coupling the saturation temperature drop DTs.r and the driving temperature difference DTd.r in a single Performance Evaluation Criteria (PEC) called Total Temperature Penalisation (TTP) equal to
TTP ¼ DT d:r þ 1=2T s:r
ð5Þ
The above analysis can be extended also to refrigerant boiling and condensation inside a BPHE by using proper correlations for computing the different terms. In particular the refrigerant heat transfer coefficient hr and the different refrigerant pressure drop components are calculated by the following correlations for BPHE [9,10]:
( hr ¼
ðhnb > hcb Þ
hnb 2 ðhcb
þ
2 1=2 hnb Þ
ðhnb > hcb Þ
ðboilingÞ
ð6aÞ
1/3 hcb = 0.122 U (kL/dh) Re0.8 eq PrL
hnb = 0.580 U h0 CRa F(p*) (q/q0)0.467 Reeq = G [(1 X) + X (qL/qG)1/2] dh/lL PrL = lL cpL/kL
hr ¼ 1:875UðkL =dh Þ Re0:445 Pr1=3 eq L
ðcondensationÞ
ð6bÞ
Dpf ¼ nb KE=V ¼ nb G2 =ð2qm Þ ðboilingÞ
ð7aÞ
Dpf ¼ ncn KE=V ¼ ncn G2 =ð2qm Þ ðcondensationÞ
ð7bÞ
Dpa ¼ G2 ð1=qG 1=qL ÞDX
ð8Þ
Dpg ¼ g qm L
ð9Þ
Dpc ¼ 1:5G2 =ð2qm Þ
ð10Þ
U is the enlargement factor of the corrugated plates (ratio between the actual and the projected area of the plates); h0 is the reference value (p*0 = 0.1, q0 = 20,000 W m2, Ra0 = 0.4 lm) of the nucleate boiling heat transfer coefficient, specific for each refrigerant. CRa ¼ ðRa = 0:4lmÞ0:1333
ð11Þ
accounts for the effect in nucleate boiling of the arithmetic mean roughness Ra (lm) of the plates as defined in ISO4287/1.
Fðp Þ ¼ 1:2p0:27 þ ½2:5 þ 1 =ð1 p Þp
ð12Þ
*
accounts for reduced pressure p effect in nucleate boiling. cp, q, k, l are the refrigerant heat capacity, density, thermal conductivity and dynamic viscosity; DX is the vapour quality change between inlet and outlet; qm is the average two-phase density between inlet and outlet calculated by the homogeneous model at the average vapour quality Xm between inlet and outlet; nb and ncn are the frictional pressure drop coefficients in boiling and condensation, specific for each BPHE and refrigerant. Table 2 gives the reference heat transfer coefficients h0 and the frictional pressure drop coefficients nb and ncn for the BPHE illustrated in Fig. 1 and Table 1. The refrigerant properties are evaluated by the NIST Standard Reference Database REFPROP 10.0 [11].
Table 2 Reference heat transfer coefficients h0 and frictional pressure drop coefficients nb and ncn. Refrigerant ASHRAE Classification
Refrigerant Chemical Name
ho (W m2 K1)
nb
ncn
R134a R410A
1,1,1,2-Tetrafluoroethane Difluoromethane (50%) Pentafluoroethane (50%) 1,1,1,3,3,3-Hexafluoropropane Pentafluoroethane (44%) 1,1,1-Trifluoroethane (52%) 1,1,1,2-Tetrafluoroethane (4%) Difluoromethane 1,1-Difluoroethane Isobutane Propane Propylene 2,3,3,3- tetrafluoroprop-1-ene trans-1,3,3,3-Tetrafluoropropene 1(Z)-1,3,3,3-Tetrafluoropropene trans-1-Chloro-3,3,3-trifluoro-1-propene
3500 4400
1425 1553
1835 2050
2850 3300
1490 1667
2000 1800
4250 3800 3200 4000 4370 3100 3200 3000 2500
1666 1465 1525 1525 1525 1460 1666 1300 1300
2000 1800 1900 1900 1900 1875 1800 1800 1800
R236fa R404A
R32 R152a R600a R290 R1270 R1234yf R1234ze(E) R1234ze(Z) R1233zd(E)
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Fig. 5. Driving temperature difference vs. saturation temperature drop in boiling inside a commercial BPHE. (Theoretical Analysis).
Fig. 6. Driving temperature difference vs. saturation temperature drop in forced convection condensation inside a commercial BPHE. (Theoretical Analysis).
Figs. 5 and 6 shows the driving temperature difference DTd.r, versus the saturation temperature drops DTs.r calculated for boiling at 20 °C and condensation at 30 °C of 13 different refrigerants inside the commercial BPHE, respectively. The analysis was carried out under the same operating conditions of the experimental boiling and condensation tests inside the commercial BPHE. The comparison between Fig. 3 and Fig. 5 and between Fig. 4 and Fig. 6, showing a fair agreement with the experimental data, confirms the robustness of the theoretical assessment based on the saturation temperature drop DTs.r, the driving temperature difference DTd.r and the Total Temperature Penalisation (TTP). Figs. 7 and 8 shows the heat transfer coefficients hr versus the Total Temperature Penalisation (TTP) calculated for boiling at 20 °C and condensation at 30 °C of 13 different refrigerants inside the commercial BPHE. The more effective refrigerants appear in
the higher left portion of the diagram (higher heat transfer coefficients and lower TTP), while the thermal and hydraulic performances decreases moving toward the lower right portion of the diagram (lower heat transfer coefficients and higher TTP). Considering boiling inside BPHE, refrigerant R410A shows performances slightly higher than those of its low GWP substitute R32, while refrigerant R404A exhibits performances identical to those of R290 and lower than those of R1270. Refrigerant R134a shows boiling performances very similar to those of all its low GWP substitutes (R1234yf, R1234ze(E) and R152a), while refrigerant R236fa exhibits boiling performances similar to those of R1234ze(Z), lower than those of R600a and higher than those of R1233zd(E). With reference to condensation inside BPHE, refrigerants R32 is more effective than R410A, while refrigerant R1270 is more
G.A. Longo et al. / International Journal of Heat and Mass Transfer 139 (2019) 31–38
37
Fig. 7. Heat transfer coefficients vs. Total Temperature Penalisation (TTP) in boiling inside a commercial BPHE. (Theoretical Analysis).
Fig. 8. Heat transfer coefficients vs. Total Temperature Penalisation (TTP) in forced convection condensation inside a commercial BPHE. (Theoretical Analysis).
effective than both R290 and R404A. Refrigerant R134a shows condensation performances similar to R152a and higher than those of R1234ze(E) and R1234yf. Refrigerant R600a exhibits condensation performances similar to those of R1234ze(Z) and higher than those of both R236fa and R1233zd(E). The results of the theoretical assessment confirm those of the experimental assessment, providing a sound procedure for the heat-transfer assessment of the low GWP alternatives for traditional HFC refrigerants.
fer characteristics to R410A. Similarly the refrigerants R290 and R1270 exhibit superior heat transfer performance to R404A and the refrigerants R152a, R1234yf and R1234ze(E) show heat transfer characteristics comparable to R134a. Finally, R1234ze(Z) exhibits superior heat transfer performance to R236fa. The coupling of experimental and theoretical assessment provides a sound procedure for the heat-transfer selection of the low GWP alternatives for traditional HFC refrigerants.
4. Conclusions Declaration of Competing Interest The results of both the experimental and the theoretical assessment confirm that the refrigerant R32 has comparable heat trans-
The authors declare that there is no conflict of interest.
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Acknowledgement This research project was partially funded by: CariVerona Foundation, Verona, Italy, Ricerca Scientifica e Tecnologica 2016-2019: ‘‘Sostenibilità e autenticazione nutrizionale di filiere lattierocasearie a tutela del consumatore”.
[4]
[5]
[6]
Appendix A. Supplementary material
[7]
Supplementary data to this article can be found online at https://doi.org/10.1016/j.ijheatmasstransfer.2019.04.144.
[8]
[9]
References [1] J.M. Calm, The next generation of refrigerants – Historical review, considerations, and outlook, Int. J. Refrig. 31 (2008) 1123–1133. [2] United Nations. Montreal Protocol on Substances That Deplete the Ozone Layer. United Nations Environment Program, Ozone Secretariat, New York, NY, USA, 1987. [3] Unep, 38th Meeting of the Open-ended Working Group (OEWG 38) of the Parties to the Montreal Protocol on Substances that Deplete the Ozone Layer
[10]
[11]
and 28th Meeting of the Parties to the Montreal Protocol (MOP 28), Kigali, 8, 10-14 October 2016. M.O. McLinden, J.S. Brown, R. Brignoli, A.F. Kazakov, P.A. Domanski, Limited options for low-global-warming-potential refrigerants, Nat. Commun. 8 (2017) 14476. G.A. Longo, A. Gasparella, Heat transfer and pressure drop during HFC refrigerant vaporisation inside a brazed plate heat exchanger, Int. J. Heat Mass Transf. 50 (2007) 5194–5203. G.A. Longo, Heat transfer and pressure drop during HFC refrigerant saturated vapour condensation inside a brazed plate heat exchanger, Int. J. Heat Mass Transf. 53 (2010) 1079–1087. A. Cavallini, J.S. Brown, D. Del Col, C. Zilio, In-tube condensation performance of refrigerants considering penalization terms (exergy losses) for heat transfer and pressure drop, Int. J. Heat Mass Transf. 53 (2010) 2886–2896. J.S. Brown, C. Zilio, R. Brignoli, A. Cavallini, Heat transfer and pressure drop penalization terms (exergy losses) during flow boiling of refrigerants, Int. J. Energy Res. 37 (2013) 1669–1679. G.A. Longo, C. Zilio, G. Righetti, A new computational procedure for refrigerant condensation inside herringbone-type brazed plate heat exchangers, Int. J. Heat Mass Transf. 82 (2015) 530–536. G.A. Longo, S. Mancin, G. Righetti, C. Zilio, A new model for refrigerant boiling inside brazed plate heat exchangers (BPHEs), Int. J. Heat Mass Transf. 91 (2015) 144–149. E.W. Lemmon, I.H. Bell, M.L. Huber, M.O. McLinden, NIST Reference Fluid Thermodynamic and Transport Properties Database - REFPROP, Version 10.0, National Institute of Standards and Technology, Standard Reference Data Program, Gaithersburg, 2017.
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Heat-transfer assessment of the low GWP substitutes for traditional HFC refrigerants Giovanni A. Longo ⇑, Giulia Righetti, Claudio Zilio University of Padova, Department of Management and Engineering, I-36100 Vicenza, Italy
a r t i c l e
i n f o
Article history: Received 21 January 2019 Received in revised form 26 April 2019 Accepted 29 April 2019 Available online 7 May 2019 Keywords: Assessment GWP Brazed Plate Heat Exchangers BPHE Performance Evaluation Criteria PEC Total Temperature Penalisation TTP
a b s t r a c t This paper presents the heat-transfer assessment of different low GWP substitutes for traditional HydroFluoroCarbon (HFC) refrigerants obtained by applying both an experimental analysis based on the direct measurement of thermal (heat transfer coefficients) and hydraulic (pressure drops) performances and a theoretical analysis based on a specific Performance Evaluation Criteria (PEC), the Total Temperature Penalisation (TTP), in the specific case of boiling and condensation in a commercial Brazed Plate Heat Exchanger (BPHE). The results of both the experimental and the theoretical assessment confirm that the refrigerant R32 has comparable heat transfer characteristics to R410A. Similarly the refrigerants R290 and R1270 exhibit superior heat transfer performance to R404A; the refrigerants R152a, R1234yf and R1234ze(E) show heat transfer characteristics comparable to R134a and R1234ze (Z) exhibits superior heat transfer performance to R236fa. The coupling of experimental and theoretical assessments provides a sound procedure for the heat-transfer assessment of the low GWP alternatives for traditional HFC refrigerants. Ó 2019 Elsevier Ltd. All rights reserved.
1. Introduction The assessment of a refrigerant must take into account several constrains including thermodynamic and thermophysical properties, environmental impact such as Ozone Depletion Potential (ODP) and Global Warming Potential (GWP), safety issues such as flammability and toxicity, chemical characteristics such as stability and compatibility with the materials, availability and cost. The first operating fluids used in refrigeration technology (1830–1930s) were in general toxic, flammable, and highly reactive and they were eliminated in 1930s by the introduction of the ChloroFluoroCarbon (CFC) and HydroChloroFluoroCarbon (HCFC) refrigerants characterised by good thermodynamic properties and chemical stability, absence of toxicity and flammability, good compatibility with material and reduced cost [1]. The phase-out of CFC and HCFC refrigerants in 1990–2010s and their progressive substitution with HydroFluoroCarbon (HFC) refrigerants was due to an environmental issue, their unacceptable ODP [2]. Nowadays traditional HFC refrigerants are subjected to a progressive reduction in use or a complete phase-out due to their high GWP according to an international protocol [3].
⇑ Corresponding author. E-mail address: [email protected] (G.A. Longo). https://doi.org/10.1016/j.ijheatmasstransfer.2019.04.144 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.
McLinden et al. in 2017 [4] investigated 184,000 small molecules for identifying the possible substitutes for traditional HFC refrigerants applying thermodynamic properties, GWP, flammability, stability, and toxicity constraints. Only 138 molecules exhibit a GWP lower than 1000, a critical temperature between 320 and 420 K, no or mild flammability, no toxicity and sufficient chemical stability. Most of the 138 molecules are relatively «new» fluids that have need of a sound assessment before the extended application in refrigeration and air conditioning. Currently there is no consolidated and comprehensive method or procedure for the heattransfer assessment of a refrigerant valid for the different heat transfer processes (single-phase, boiling, condensation) and the different type of heat exchangers (tubular, micro-channel, plate heat exchanger, etc.). This paper presents the heat-transfer assessment of different low GWP fluids that have been proposed as possible substitutes for traditional HFC refrigerants. The assessment is obtained through the application of both an experimental analysis based on the direct measurement of thermal (heat transfer coefficients) and hydraulic (pressure drops) performances and a theoretical analysis based on a specific Performance Evaluation Criteria (PEC), the Total Temperature Penalisation (TTP). This approach is applied in the specific case of refrigerant two-phase heat transfer inside Brazed Plate Heat Exchanger (BPHE) in forced convection regime.
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G.A. Longo et al. / International Journal of Heat and Mass Transfer 139 (2019) 31–38
Nomenclature A b cp dh g G h J KE/V L m nch p p* P Pr q Q Ra Rp Reeq S s T TTP W X
area of the plate (m2) height of the corrugation (m) specific heat capacity (J kg1 K1) hydraulic diameter, dh = 2b (m) gravity acceleration (m s2) mass flux, G = m/(nch W b) (kg m2 s1) heat transfer coefficient (W m2 K1) specific enthalpy (J kg1) specific kinetic energy (J m3) flow length of the plate (m) mass flow rate (kg s1) number of channels pressure (Pa) reduced pressure corrugation pitch (m) Prandtl number, Pr = l cp/k heat flux, q = Q/S (W m2) heat flow rate (W) arithmetic mean roughness (ISO4271/1) (lm) roughness (DIN 4762/1) (lm) equivalent Reynolds number, Reeq = G [(1 X) + X (qL/ qG)1/2] dh/lL heat transfer area (m2) plate wall thickness (m) temperature (K) Total Temperature Penalisation (K) width of the plate (m) vapour quality, X = (J JL)/DJLG
Greek symbols b inclination angle of the corrugation D difference
DJLG U k
l q
n
latent heat of vaporisation/condensation (J kg1) enlargement factor thermal conductivity (W m1 K1) viscosity (kg m1 s1) density (kg m3) frictional pressure drop coefficient
Subscripts a momentum b boiling ave average c manifold and port cb convective boiling cn condensation d driving eq equivalent f frictional g gravity G vapour phase in inlet L liquid phase LG liquid gas phase change m average value nb nucleate boiling out outlet r refrigerant s saturation sf secondary fluid t total w water
2. Experimental assessment The experimental heat-transfer assessment of a refrigerant involves the combined evaluation of thermal and hydraulic performances in a specific application. The authors of the present paper had carried out in the past an extensive analysis of refrigerants condensation and vaporisation inside a commercial Brazed Plate Heat Exchanger (BPHE) [5,6]. The traditional HFC refrigerants (R134a, R410A, R404A, R236fa), HydroCarbon (HC) refrigerants (R600a, R290, R1270), low GWP HFC refrigerants (R152a, R32), HydroFluoroOlefin (HFO) refrigerants (R1234yf, R1234ze(E), R1234ze(Z)) and also a HydroChloroFluoroOlefin (HCFO) refrigerant (R1233zd(E)) were considered. The tested BPHE consists of 10 plates, 72 mm in width and 310 mm in length, which present a macro-scale herringbone corrugation with an inclination angle of 65° and a corrugation amplitude of 2 mm: Fig. 1 and Table 1 give the main geometrical characteristics of the tested BPHE. The thermal and hydraulic performances of a refrigerant may be evaluated by considering two temperature differences: the saturation temperature drop DTs.r, due to the total refrigerant pressure drop Dpt, and the driving temperature difference DTd.r, inverse to the heat transfer coefficient hr. Both DTs.r and DTd.r reduce the overall refrigeration system energetic and exergetic efficiencies by increasing the required compressor power for achieving the same refrigeration effect. The saturation temperature drop DTs.r, equal to the difference between the saturation temperature of the refrigerant at the inlet
Fig. 1. Schematic view of a plate of the commercial BPHE.
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The driving temperature difference DTd.r, equal to the average difference between the wall temperature and the saturation temperature, is related to the refrigerant heat transfer coefficient hr
Table 1 Geometrical characteristics of the commercial BPHE. Parameter
Measure/Type
Fluid flow plate length L (mm) Plate width W (mm) Area of the plate A (m2) Corrugation type Angle of the corrugation (°) Corrugation amplitude b (mm) Corrugation pitch P (mm) Plate roughness Ra (m) Plate roughness Rp (m) Number of plates Channels on refrigerant side nch.r Channels on water side nch.w
278.0 72.0 0.020 Herringbone 65 2.0 8.0 0.4 1.0 10 4 5
33
DT d:r ¼ T w:ave T s:ave ¼ ðmr JLG Þ=ðShr Þ ¼ q=hr
and at the outlet of the heat exchanger, is related to the total refrigerant pressure drop Dpt
DT s:r ¼ T s ðpin Þ T s ðpout Þ
ð1Þ
pout ¼ pin Dpt
ð2Þ
Dpt ¼ Dpf þ Dpa þ Dpg þ Dpc
ð3Þ
where Dpf is the frictional pressure drop, Dpa the momentum pressure drop, Dpg the gravity pressure drop and Dpc the manifold and ports local pressure drop. The condensation and boiling total pressure drops enhance the pressure difference between the evaporator and the condenser increasing the compression work. In addition, they cause a reduction of the saturation temperature in the evaporator and the condenser with negative effects on the driving temperature differences. Therefore condensation and boiling total pressure drops produce a double negative effect on the global performance of a refrigeration system.
ð4Þ
where mr is the refrigerant mass flow rate, DJLG the refrigerant latent heat of vaporisation/condensation, S the heat transfer area and q the heat flux. High condensation and boiling heat transfer coefficients involve lower driving temperature differences with positive effects on the global performance of a refrigeration system. Fig. 2 shows the temperature profiles in a boiling process inside a counter-flow heat exchanger highlighting the two characteristic temperature differences DTs.r and DTd.r. Therefore, in order to evaluate experimentally the thermal and hydraulic performances of the different refrigerants in two phase heat transfer inside a commercial BPHE the experimental data points are plotted showing the driving temperature difference DTd.r, inverse to the refrigerant heat transfer coefficients, versus the saturation temperature drops DTs.r caused by the total pressure drops. The more effective refrigerants appear in the lower left portion of the diagram (higher heat transfer coefficients and lower saturation temperature drops), while the thermal and hydraulic performances decreases moving toward the higher right portion of the diagram (lower heat transfer coefficients and higher saturation temperature drops). Fig. 3 shows the driving temperature difference DTd.r versus the saturation temperature drops DTs.r measured in the boiling tests at 20 °C of saturation temperature inside the commercial BPHE illustrated in Fig. 1 and Table 1 with 13 different refrigerants. The boiling tests were carried out under similar operating conditions for all the refrigerants tested with an inlet vapour quality around 0.2–0.3, an outlet vapour quality around 0.8–1.0, an heat flux in the range 2–23 kW m2 and a refrigerant mass velocity in the range 5–40 kg m2 s1. The refrigerant boiling inside the tested BPHE was controlled by the combined effects of nucleate boiling and forced convection boiling. In general, nucleate boiling is the dominant heat transfer regime for
Fig. 2. Temperature profiles in a boiling process inside a counter-flow heat exchanger.
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Fig. 3. Driving temperature difference vs. saturation temperature drop in boiling inside a commercial BPHE. (Experimental Data).
high and medium pressure refrigerants, while forced convection boiling is the dominant heat transfer regime for low pressure refrigerants. In Fig. 3, it is easy to distinguish the high pressure refrigerants (round symbol), from the medium pressure refrigerants (triangular symbol) and the low pressure refrigerants (quadratic symbol). Within the high pressure fluids, the traditional refrigerant R404A and R410A exhibit boiling performances slightly higher than those of their possible low GWP substitutes R32, R290 and R1270. Considering the medium pressure fluids, the traditional refrigerant R134a shows boiling performances very similar to the low GWP substitute R1234yf, slightly higher than those of the low GWP substitute R152a and slightly lower than those of the low GWP substitute R1234ze(E). Finally, taking into account the low pressure fluids, the traditional refrigerant R236fa exhibits
boiling performances higher than those of all its low GWP substitutes R600a, R1234ze(Z) and R1233zd(E). Fig. 4 shows the driving temperature difference DTd.r, versus the saturation temperature drops DTs.r measured during the condensation tests at 30 °C of saturation temperature inside the commercial BPHE illustrated in Fig. 1 and Table 1 with 12 different refrigerants in forced convection condensation regime (Reeq > 1600). The condensation tests were carried out under similar operating conditions for all the refrigerants tested with an inlet vapour quality around 0.9–1.0, an outlet vapour quality around 0.0–0.1, an heat flux in the range 9–34 kW m2 and a refrigerant mass velocity in the range 15–43 kg m-2s1. The refrigerant condensation inside the BPHE tested was gravity controlled at low refrigerant mass velocity (G < 15–20 kg m2 s1), while forced convection conden-
Fig. 4. Driving temperature difference vs. saturation temperature drop in forced convection condensation inside a commercial BPHE. (Experimental Data).
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sation occurs at higher refrigerant mass velocity (G > 15– 20 kg m2 s1). The transition point from gravity controlled to forced convection condensation was found for an equivalent Reynolds number Reeq around 1600 for all the refrigerants tested. Also in this case it is easy to distinguish the high pressure refrigerants (round symbol), from the medium pressure refrigerants (triangular symbol) and the low pressure refrigerants (quadratic symbol). The low GWP refrigerant R32 exhibits condensation performances very similar to those of the traditional refrigerant R410A, while the low GWP refrigerants R290 and R1270 shows condensation performances very similar to those of the traditional refrigerant R404A. Refrigerant R134a exhibits condensation performances very similar to those of R1234yf, slightly higher than those of R152a and slightly lower than those of R1234ze(E). Refrigerant R1234ze(Z) exhibits condensation performances higher both than those of the traditional refrigerant R236fa and also of the alternative low GWP substitute R600a. The results of this experimental assessment confirm that the refrigerant R32 has comparable heat transfer characteristics to R410A, while the refrigerants R290 and R1270 exhibit superior heat transfer performance to R404A; refrigerants R152a, R1234yf and R1234ze(E) show heat transfer characteristics comparable to R134 and R1234ze(Z) exhibits superior heat transfer performance to R236fa.
3. Theoretical assessment The theoretical heat-transfer ranking of refrigerants requires the coupled analysis of thermal (heat transfer coefficients) and hydraulic (pressure drops) performances in a specific operating case: Cavallini et al. [7] carried out an exhaustive analysis on the Performance Evaluation Criteria (PEC) for refrigerant two-phase heat transfer inside tubes. In 2013 Brown et al. [8] presented the combined thermal and hydraulic analysis of refrigerant flow boiling inside plain round tubes by coupling the saturation temperature drop DTs.r and the driving temperature difference DTd.r in a single Performance Evaluation Criteria (PEC) called Total Temperature Penalisation (TTP) equal to
TTP ¼ DT d:r þ 1=2T s:r
ð5Þ
The above analysis can be extended also to refrigerant boiling and condensation inside a BPHE by using proper correlations for computing the different terms. In particular the refrigerant heat transfer coefficient hr and the different refrigerant pressure drop components are calculated by the following correlations for BPHE [9,10]:
( hr ¼
ðhnb > hcb Þ
hnb 2 ðhcb
þ
2 1=2 hnb Þ
ðhnb > hcb Þ
ðboilingÞ
ð6aÞ
1/3 hcb = 0.122 U (kL/dh) Re0.8 eq PrL
hnb = 0.580 U h0 CRa F(p*) (q/q0)0.467 Reeq = G [(1 X) + X (qL/qG)1/2] dh/lL PrL = lL cpL/kL
hr ¼ 1:875UðkL =dh Þ Re0:445 Pr1=3 eq L
ðcondensationÞ
ð6bÞ
Dpf ¼ nb KE=V ¼ nb G2 =ð2qm Þ ðboilingÞ
ð7aÞ
Dpf ¼ ncn KE=V ¼ ncn G2 =ð2qm Þ ðcondensationÞ
ð7bÞ
Dpa ¼ G2 ð1=qG 1=qL ÞDX
ð8Þ
Dpg ¼ g qm L
ð9Þ
Dpc ¼ 1:5G2 =ð2qm Þ
ð10Þ
U is the enlargement factor of the corrugated plates (ratio between the actual and the projected area of the plates); h0 is the reference value (p*0 = 0.1, q0 = 20,000 W m2, Ra0 = 0.4 lm) of the nucleate boiling heat transfer coefficient, specific for each refrigerant. CRa ¼ ðRa = 0:4lmÞ0:1333
ð11Þ
accounts for the effect in nucleate boiling of the arithmetic mean roughness Ra (lm) of the plates as defined in ISO4287/1.
Fðp Þ ¼ 1:2p0:27 þ ½2:5 þ 1 =ð1 p Þp
ð12Þ
*
accounts for reduced pressure p effect in nucleate boiling. cp, q, k, l are the refrigerant heat capacity, density, thermal conductivity and dynamic viscosity; DX is the vapour quality change between inlet and outlet; qm is the average two-phase density between inlet and outlet calculated by the homogeneous model at the average vapour quality Xm between inlet and outlet; nb and ncn are the frictional pressure drop coefficients in boiling and condensation, specific for each BPHE and refrigerant. Table 2 gives the reference heat transfer coefficients h0 and the frictional pressure drop coefficients nb and ncn for the BPHE illustrated in Fig. 1 and Table 1. The refrigerant properties are evaluated by the NIST Standard Reference Database REFPROP 10.0 [11].
Table 2 Reference heat transfer coefficients h0 and frictional pressure drop coefficients nb and ncn. Refrigerant ASHRAE Classification
Refrigerant Chemical Name
ho (W m2 K1)
nb
ncn
R134a R410A
1,1,1,2-Tetrafluoroethane Difluoromethane (50%) Pentafluoroethane (50%) 1,1,1,3,3,3-Hexafluoropropane Pentafluoroethane (44%) 1,1,1-Trifluoroethane (52%) 1,1,1,2-Tetrafluoroethane (4%) Difluoromethane 1,1-Difluoroethane Isobutane Propane Propylene 2,3,3,3- tetrafluoroprop-1-ene trans-1,3,3,3-Tetrafluoropropene 1(Z)-1,3,3,3-Tetrafluoropropene trans-1-Chloro-3,3,3-trifluoro-1-propene
3500 4400
1425 1553
1835 2050
2850 3300
1490 1667
2000 1800
4250 3800 3200 4000 4370 3100 3200 3000 2500
1666 1465 1525 1525 1525 1460 1666 1300 1300
2000 1800 1900 1900 1900 1875 1800 1800 1800
R236fa R404A
R32 R152a R600a R290 R1270 R1234yf R1234ze(E) R1234ze(Z) R1233zd(E)
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Fig. 5. Driving temperature difference vs. saturation temperature drop in boiling inside a commercial BPHE. (Theoretical Analysis).
Fig. 6. Driving temperature difference vs. saturation temperature drop in forced convection condensation inside a commercial BPHE. (Theoretical Analysis).
Figs. 5 and 6 shows the driving temperature difference DTd.r, versus the saturation temperature drops DTs.r calculated for boiling at 20 °C and condensation at 30 °C of 13 different refrigerants inside the commercial BPHE, respectively. The analysis was carried out under the same operating conditions of the experimental boiling and condensation tests inside the commercial BPHE. The comparison between Fig. 3 and Fig. 5 and between Fig. 4 and Fig. 6, showing a fair agreement with the experimental data, confirms the robustness of the theoretical assessment based on the saturation temperature drop DTs.r, the driving temperature difference DTd.r and the Total Temperature Penalisation (TTP). Figs. 7 and 8 shows the heat transfer coefficients hr versus the Total Temperature Penalisation (TTP) calculated for boiling at 20 °C and condensation at 30 °C of 13 different refrigerants inside the commercial BPHE. The more effective refrigerants appear in
the higher left portion of the diagram (higher heat transfer coefficients and lower TTP), while the thermal and hydraulic performances decreases moving toward the lower right portion of the diagram (lower heat transfer coefficients and higher TTP). Considering boiling inside BPHE, refrigerant R410A shows performances slightly higher than those of its low GWP substitute R32, while refrigerant R404A exhibits performances identical to those of R290 and lower than those of R1270. Refrigerant R134a shows boiling performances very similar to those of all its low GWP substitutes (R1234yf, R1234ze(E) and R152a), while refrigerant R236fa exhibits boiling performances similar to those of R1234ze(Z), lower than those of R600a and higher than those of R1233zd(E). With reference to condensation inside BPHE, refrigerants R32 is more effective than R410A, while refrigerant R1270 is more
G.A. Longo et al. / International Journal of Heat and Mass Transfer 139 (2019) 31–38
37
Fig. 7. Heat transfer coefficients vs. Total Temperature Penalisation (TTP) in boiling inside a commercial BPHE. (Theoretical Analysis).
Fig. 8. Heat transfer coefficients vs. Total Temperature Penalisation (TTP) in forced convection condensation inside a commercial BPHE. (Theoretical Analysis).
effective than both R290 and R404A. Refrigerant R134a shows condensation performances similar to R152a and higher than those of R1234ze(E) and R1234yf. Refrigerant R600a exhibits condensation performances similar to those of R1234ze(Z) and higher than those of both R236fa and R1233zd(E). The results of the theoretical assessment confirm those of the experimental assessment, providing a sound procedure for the heat-transfer assessment of the low GWP alternatives for traditional HFC refrigerants.
fer characteristics to R410A. Similarly the refrigerants R290 and R1270 exhibit superior heat transfer performance to R404A and the refrigerants R152a, R1234yf and R1234ze(E) show heat transfer characteristics comparable to R134a. Finally, R1234ze(Z) exhibits superior heat transfer performance to R236fa. The coupling of experimental and theoretical assessment provides a sound procedure for the heat-transfer selection of the low GWP alternatives for traditional HFC refrigerants.
4. Conclusions Declaration of Competing Interest The results of both the experimental and the theoretical assessment confirm that the refrigerant R32 has comparable heat trans-
The authors declare that there is no conflict of interest.
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Acknowledgement This research project was partially funded by: CariVerona Foundation, Verona, Italy, Ricerca Scientifica e Tecnologica 2016-2019: ‘‘Sostenibilità e autenticazione nutrizionale di filiere lattierocasearie a tutela del consumatore”.
[4]
[5]
[6]
Appendix A. Supplementary material
[7]
Supplementary data to this article can be found online at https://doi.org/10.1016/j.ijheatmasstransfer.2019.04.144.
[8]
[9]
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