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Prediction of evaporation heat transfer coefficient and pressure drop of refrigerant mixtures in horizontal tubes
Prediction of evaporation heat transfer coefficient and pressure drop of refrigerant mixtures in horizontal tubes D. Jung Department of Mechanical Engineering, Inha University, Inchon, Korea R. R a d e r m a c h e r Department of Mechanical Engineering, University of Maryland, College Park, MD 20742, USA Received 30 January 1992; revised 26 August 1992
A study on the prediction of heat transfer coefficient and pressure drop of refrigerant mixtures is reported. Heat transfer coefficients and pressure drops of prospective mixtures to replace R12 and R22 are predicted on the same cooling capacity basis assuming evaporation in horizontal tubes. Results indicate that nucleate boiling is suppressed at qualities greater than 20% for all mixtures, and evaporation becomes the main heat transfer mechanism. For the same capacity, some mixtures containing R32 and R 152a show 8 10% increase in heat transfer coefficients. Some mixtures with large volatility difference exhibit as much as 55% reduction compared to RI2 and R22, caused by mass transfer resistance and property degradation due to mixing (32%) and reduced mass flow rates (23%). Other mixtures with moderate volatility difference exhibit 2030% degradation due mainly to reduced mass flow rates. The overall impact of heat transfer degradation, however, is insignificant if major heat transfer resistance exists in the heat transfer fluid side (air system). If the resistance in the heat transfer fluid side is of the same order of magnitude as that on the refrigerant side (water system), considerable reduction in overall heat transfer coefficient of up to 20% is expected. A study of the effect of uncertainties in transport properties on heat transfer shows that transport properties of liquid affect heat transfer more than other properties. Uncertainty of 10% in transport properties causes a change of less than 6% in heat transfer prediction. (Keywords:heat transfer coefficientprediction;alternativerefrigerants;refrigerantmixtures)
Prrvision du coefficient de transfert de chaleur et de la perte de charge lors de l'rvaporation de mrlanges de frigorigrnes dans des tubes horizontaux Cette ktude porte sur l'estimation du coefficient de transfert de chaleur et de la perte de charge pour des mklanges de Jkigorigbnes, substituts potentiels du RI2 et du R22 pour la mdme capacitb fr~gor(fique et bvaporation en tubes horizontaux. Les rbsultats indiquent que l'bbullition nuclkke est supprimbe pour tousles mblanges prbsentant des concentrations supbrieures h 20% et que l'bvaporation devient alors le mbcanisme principal du transfert de chaleur. A puissance bgale, quelques mklanges contenant du R32 et du R152a montrent u n e augmentation des coefficients de transfert de chaleur de 8~10%. Pour quelques mdlanges de produits gt volatilitbs trbs d,iffkrentes, on observe une rbduction de 55% par rapport au R12 et au R22, causde par la rbsistance au transfert de masse et la dbgradation des propribtbs dues au mHange (32%) et gl la rbduction des dbbits masse (23%). D 'autres mblanges, ayant une moindre difference de volatilitk, montrent une dbgradation de 20-30% due surtout h la rbduction des dbbits masse. Nbanmoins, l'effet global de la rbduction du tramfert de chaleur est insignifiant si la rbsistance au transfert de chaleur est plus forte du c6tb du fluide qui transJOre la chaleur (systbme h air). Si la rdsistance du cdtb du fluide de transfert de chaleur est du m~me ordre de grandeur que celle du crtb du frigorig~ne (systkme h eau), on pr~voit une rbduction considbrable du coefficient de tramfert de chaleur global s'Nevant h 2(t%. Une ktude sur l'effet des incertitudes clans les propribtks de transfert sur le transfert de chaleur montre que les propribtbs de transfert du liquide influencent plus Jbrtement le tran~fert de chaleur que d'autres propribtbs. Une incertitude de 10% dans les propribtbs de transfert modifie de moins de 6% les prbvisions du transfert de chaleur. (Mots clrs: prbvision du coefficient de transfert de chaleur; frigorigrne de remplacement; mrlange de frigorigrnes)
Fully halogenated chlorofluorocarbons (CFC) are responsible for destroying the stratospheric ozone layer and increasing global warming 1. At present, research is being undertaken at many institutions to replace fully halogenated CFCs such as R l l , R12, R I 3 B I , Rl13, R l I 4 and R115, of which trade and production are regulated by the Montreal Protocol z. Thanks to extensive research during the past years, new ozone-safe fluids such as 0140-7007/93/030201~)9 ~ 1993 Butterworth Heinemann Ltd and IIR
R134a, R141b, R142b, R143a, R123 and R124 have been developed. These new fluids, however, are to be tested for certain requirements such as toxicity and material compatibility. In replacing ozone-depleting substances, one of the basic properties to be satisfied by any substitute fluids is vapour pressure. As vapour pressure changes, so does volumetric capacity. A change in volumetric capacity
Rev. Int. Froid 1993 Vol 16 No 3
201
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher
Nomenclature
G
D F .!'
G h k L rn N P q T X X
Y U
specific heat (J kg ~K- ~) tube diameter (m) heat transfer enhancement factor friction factor mass flux (kg m 2s-t) heat transfer coefficient (W m 2 K ~) latent heat of vaporization (kJ kg-t) thermal conductivity (W m-~ K 1) tube length (m) mass flow rate (g s-~) factor due to nucleate boiling pressure (kPa) heat flux (Win 2) temperature (K or °C) liquid phase concentration quality vapour phase concentration overall heat transfer coefficient (W m-2 oC 1)
implies a change in compressor, a key component in HVAC equipment. Newly developed ozone-safe fluids may not provide the desired vapour pressure of the fluids to be replaced. An easy way of manipulating vapour pressure to obtain desired pressure is using mixtures. For instance, vapour pressure of R12 can be obtained with R22/R142b mixture, of which R22 and R142b are more and less volatile than the target fluid, R12, respectively. Non-azeotropic refrigerant mixtures (NARMs) are not only able to meet the vapour pressure requirement, but also increase the system performance by utilizing a gliding temperature effect, as evidenced by some studies 3,4. Since the introduction of N A R M S in the HVAC area, there has been an unresolved issue of degradation of heat transfer coefficients (HTC) associated with NARMs. For instance, Stoecker 5 tested a prototype refrigerator with R22/R114 mixtures. In his experiments, the heat transfer fluid employed was not air but a mixture of ethylene glycol and water. Stoecker has observed a noticeable decrease in HTCs and degradation of system performance. Another report, however, revealed that the degradation of heat transfer coefficients associated with R22/R114 mixtures is not significant enough to measure its effect on refrigerant-to-air heat exchange in an evaporator, which was based on actual evaporator test dataL Since N A R M s are prospective substitutes for ozonedepleting fluids, their heat transfer characteristics should be clarified to help engineers design proper heat exchangers. Experimental investigation would be the most reliable means to fulfil this goal, even if it would be cost- and time-consuming. As an interim solution toward better understanding of heat transfer with mixtures, a study on the prediction of heat transfer coefficients and pressure drops, based on reliable experimental data, might be useful. This report succeeds our previous study on the prediction of heat transfer coefficients of newly marketed pure refrigerants v. The goal of this paper is to predict and compare heat transfer coefficients of refrigerant mixtures 202
Int. J. Refrig. 1993 Vo116 No 3
p cr /z q5
density (kg m ~) surface tension (N m ~1 viscosity (Pa s) pressure drop multiplier
Dimensionless numbers Bo Pr Re ~u
Boiling number, q/Gh/~. Prandtl number of liquid, ~)~t/k Reynolds number, GD/p Martinelli's parameter in Table 1
Subscripts cec fo I Io nbc tp v
convective evaporation contribution total flow assumed as liquid liquid liquid only nucleate boiling contribution two-phase vapour
during evaporation, and to determine the impact of heat transfer degradation with mixtures on overall heat transfer coefficients. Pressure drop will also be predicted. Finally, the effect of uncertainties of transport properties on heat transfer prediction will be examined.
Analysis Selection of correlations To accomplish the objectives, correlations developed by Jung et al) -1° are used. Jung et al. measured more than 3000 local heat transfer coefficients and 400 pressure drops of R22, RI14, R12, R152a and their mixtures at various concentrations during evaporation. Their test section was made of a stainless steel tube of 9. I mm id and 8.0 m long. The test section was heated by passing a direct current to obtain data within a desired quality range. Data were taken at reduced pressure of 0.08 to 0.16 at the exit of the test section. Applied heat flux was from 10.0 to 45.0 kW m -z, while mass flow rate of refrigerants varied from 16.0 to 46.0 g s -1, equivalent to 250720 kg m 2 s-~. The exit quality varied from 40 to 95%. The overall compositions were 0.00, 0.21, 0.60, 0.89 and 1.00 mole fraction R12 for RI2/RI52a mixtures, and 0.00, 0.23, 0.47, 0.77 and 1.00 mole fraction R22 for R22/ R114 mixtures. These ranges and physical dimension of the test section are typical of air-source residential heat pumps of 3-5 kW capacity. Figure 1 illustrates local HTCs obtained at various heat fluxes for a given mass flow rate. Heat transfer coefficients are strongly dependent upon heat flux at low qualities. This heat transfer regime is termed a partial boiling regime, which is characterized by the formation of bubbles on the wall. As quality increases further, the annular liquid film becomes thinner so that heat transfer becomes more effective. At last, the wall superheat (Twall--Tsar) becomes smaller than the threshold wall superheat to sustain nucleation on the heating surface. This is called suppression of nucleate boiling. This phenomenon of cessation of nucleation is characteristic of flow
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher
Partial boiling Convective regime regime ~Liquid layer~ (( ~..-Vapor core.(_l~")")
e-
Z UJ W_ OIJ.l
o (3
U.I i1 CO
q5~ Increasing heat flux q 4
T
~
~_~ ~ nnbe
q 3 ~ :
Tw
t
Bubbles ~ /
2/ i
ILl
T
ATeff
ATsa t
~B fDew --~,./ line
hcec
l I
Z
< re"
ql /
2hce c
cee'
Tb a
U.I
I
Loss of available
QUALITY Figure 1 Nucleateboilingand convectiveevaporationcontributions in two-phaseheat transfer coefficient,h.~, nucleateboiling contribution; hc~,convectiveevaporationcontribution
I I
0
Figure 1 Parts de l~bullition nuclide et de l'kvaporation par convection sur le coefficient de traasfert de chaleur diphasique, h.bc ~bullition nuclide, hce~ ~vaporation par eonvection
boiling, which is markedly different from nucleate pool boiling with continual bubble generation on the wall. The quality at which nucleation ceases is called the transition quality beyond which convective evaporation occurs at the liquid-vapour interface (no bubbles are present). Heat transfer coefficients in the convective region increase with quality monotonically. Nucleate boiling was fully suppressed for all pure and mixed refrigerants tested. The quality at which nucleation ceases, however, is much lower for mixtures than for pure fluids. This is explained by examining local bubble dynamics. Bubble growth in a single component system is limited by the rate of heat diffusion to the bubble interface to provide latent heat of vaporization. In a mixture system, however, another limitation exists as a result of stripping of the more volatile component from the liquid region close to the bubble, which is caused by the volatility (or boiling point) difference of the pure components. Due to the concentration gradient near the interface, through which the more volatile component has to diffuse, mass transfer resistance to heat transfer is introduced, since the heat transfer is affected by the rate at which the more volatile component can be replenished at the interface. Figure 2 illustrates the bubble growth process of a binary mixture on a phase equilibrium diagram. The initial bulk liquid at Xb is superheated by an amount of ATsa, above the boiling point Tb, corresponding to the initial bulk liquid composition Xb. At the bubble interface the liquid phase composition is moved to X~ (point A) while the vapour composition within the bubble shifts to Y~ (point B). The consequence of this liquid phase composition shift toward the less volatile component at the interface is that the effective temperature driving potential becomes Tw-T~, less than the nominal value Tw-Tb, which is known as the loss of available wall superheat. Both mixture effects, the mass transfer resistance and the loss of wall superheat, have been shown to be responsible for slower-growing bubbles and smaller bubble departure size associated with mixtures, which in turn
Bubble i line
superheat X i
X b
Yi
Yb
COMPOSITION Figure 2 Bubble growth process of a binary mixture illustrated on a phase equilibrium diagram Figure 2 Dkveloppement de l'dbullition commenfante d'un mdlange binaire sur un diagramme d'dquilibre de phases
are related to the reduction in heat transfer coefficients with mixtures for nucleate pool boiling]U 2. The mixture effects become greater as the volatility difference between the pure components increases. This is the reason why nucleate boiling is suppressed at much lower qualities with mixtures than with pure fluids. Since experimental results supported Chen's u idea that the two-phase heat transfer coefficient is additive of nucleate boiling and convective evaporation contributions, the original form of Chen's correlation, Equation (1), was retained to correlate the measured heat transfer data '3. More detailed analysis is available elsewhere ~0. The final correlation for mixed refrigerants becomes htp=
hnbc + h~cc =
Nhun/Cun + CmeFph]o
(l)
where hnbc, h.... N, hon, Cu,, Cmc, F and h~o are a nucleate boiling contribution, a convective evaporation contribution, a nucleate boiling suppression factor, a pool boiling heat transfer coefficient obtained by UnaP ', a correction factor for mixtures in pool boiling obtained by UnaF ~, another correction factor for mixtures in evaporation, a two-phase enhancement factor, and a single-phase heat transfer coefficient for liquid-only flow obtained by Dittus and Boelter correlation,~, respectively. Table 1 lists various equations to calculate the parameters in Equation (1). The correlation shows good agreement with the experimental data yielding a mean deviation of 9.6% for R22/ R114 and R12/R152a mixtures at various concentrations under a wide range of conditions. The good agreement stems mainly from an excellent fit of Stephan and Abdelsalam's ~2nucleate pool boiling heat transfer correlation for pure refrigerants used in Unal's equation'L The heat flux dependence of N, the nucleate boiling suppression factor, is also well accounted for. Besides this good fit, the most distinctive feature of Equation (1) is that it works for both pure and mixed refrigerants. For pure
Rev. Int. Froid 1993 Vo116 No 3 203
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher 'Fable I S u m m a r y o f heat transfer coefficient correlation ['or refrigerant mixtures b y J u n g el al.' '" Tableau l Corr,;/alien entre h's coeffl'cients de transji, rt de chah, ur pour des m~;langes de [?igorig~nes, par Jung c t a l : ~' h, r = h .... + h.. = (,,ilh""
~ C.
f'~ h;.
N - 4048X,,~:Bo~"(l'orX,~ < I),N - 2.0 ().IX, ~ B o (',,, = [1.0 + (h: + bd(l.0 + b4)](1.0 ~ b.)
0.33(1`O~I ~
h: = (1.0 - A') ln[(l.01 - .V)/(1.OI - Y)] + X l n ( X / Y ) + ]Y h, = 0 . 0 ( f o r X >/ 0.01), h, = ( Y / X ) ('~ 1.0 (for X < 0.01) h4 -
152 (p/t~,,,,9 °''''.
X/Y
=
1.0 f o r ) ; -
} -
5)
A}'
0.0
where p,,,,, is the critical pressure of the more volatile component, h,,,, = h,/Cu,, =
h~ h~
/,,
x, ii X~ = h,x~ • ~,ix, h
h:
where h~ and h: are nucleate pool boiling heat transfer coefficients of c o m p o n e n t s I and 2 calculated by Stephan and Abdelsalam's correlation, h ....
hs. = 207k~ilqbd ].... (~) ..... where bd = 0.0146 fl[2o~g(p~ - p,)]~ with a contact angle/3 = 35 °. ~3",,,~ =
1.0
- 0.351Y -
Xp ~',
F = 2.37 (0.29 + XII).... =
h,,, :
(! ; ,)" 0 0~3 • --
"'
k, [di.l -
JL
..-)dl"" tC.,~"'
-~.----J
tW-)
fluids, the mixture correction factors Cun and Cm~ become 1.0, rendering Equation (!) to be identical to the correlation for pure fluids. Pressure drops were correlated following Martinelli and his co-workers (detailed information is available elsewhere~6). The final correlation for pressure drop for mixtures is
capacity basis, R l34a showed 33.3% increase in average heat transfer coefficient over RI2. The prediction by Equation (1) was 34.4% increase in average heat transfer coefficient of R 134a over R 12, showing good agreement. These comparisons with pure fluids indirectly substantiated the validity of correlations (1) and (2).
Heat trans/~r coefficients o f mixtures D ,o I
~
"i
(2)
where ~b,~ = 12.82 X, '47(1 - x ) ~-s and the friction factor Jio = 0.046 Re °.2. The mean deviation was 8.4% for all mixture data. Because no heat transfer coefficients and pressure drops for mixtures have been reported since correlations (1) and (2) were developed, further validity of correlations (1) and (2) could not be checked. As an indirect confirmation, however, more pure fluids' data were taken and compared. Jung obtained more than 300 local heat transfer coefficients and 80 pressure drops from R ll using the same experimental apparatus under similar condition. The mean deviations between the data and correlations (1) and (2) were 6.3% and 10.0% respectively 7. Recently, Eckels and Pate ~7 measured evaporation heat transfer coefficients for R134a in a horizontal tube of 8.0 m m id and 3.67 m long. Evaporation tests were performed at temperatures of 5-150C with inlet and exit qualities of 10% and 90%, respectively. The mass flux varied from 125 to 400 kg m-2s-~. Average heat transfer coefficients over the entire quality range were measured with an appropriate energy balance using a log mean temperature difference method. On the same cooling
204
Int. J. Refrig. 1993 Vol 16 No 3
Before comparing fluids for their heat transfer characteristics, a basis for fair comparison should be established. One way of comparing fluids would be on the basis of same mass flow rate at the same evaporating temperature. This approach would be unfair considering that what is actually required of each fluid is not the same flow rate, but the same capacity. Since the heat of evaporation differs greatly among many fluids at the same temperature, the required mass flow rate would vary accordingly among different fluids for desired capacity. Based on this argument, comparison will be made on the same cooling capacity basis. To accomplish this goal, a basic refrigeration cycle was modelled for various fluids of interest. In the model, no pressure drops were assumed in the evaporator and condenser. The compressor was specified by an isentropic compressor efficiency of 0.7. Since evaporation and condensation with mixtures occur over a range of temperatures (gliding temperature effect), evaporator and condenser temperatures at 50% quality are specified assuming a linear temperature glide over an entire twophase region. Under the fixed cooling capacity of 3 a n d 4 kW and condenser temperature of 40°C, evaporation temperature varied from - 10.0, 0.0 and to 10.0"C. The net refrigeration effect per unit mass, evaporator inlet
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher Table 2 Mass flow rate and evaporator inlet quality, rn (g s ') and x~,, (%) for pure and mixed refrigerants for the cooling capacity of 4 kW at different evaporation temperatures Tableau 2 Dbbit masse et concentration ?l I'entr~e de I'bvaporateur, m (g s ~) et .v,°. ( % ) pour des [?igorigbnes purs et en mblange et pour une puissance jHgor(fique de 4 kilowatts, 71 d(ff~'rentes temperatures d'~vaporation - IO°C Fluid
R 12 R22 R32 R 152a R 142b R 123 Rl41b 50% R22/50% R 142b 60% R22/40% R 123 20% R22/80% R 152a 70% R22/30% R 141 b 21)% R32/80% R 142b 10%R32/90%R152a 50% R32/50 % R 142b 40% R32/60% R 152a
J
1 I(KI0
9000
• d7000
O°C
nl
Xin
~l
Xin
hl
Xin
36.91 26.13 15.63 17.32 24.69 30.21 21.94 23.89 23.10 18.36 20.63 20.30 16.77 17.78 15.87
31.0 28.3 25.9 27.2 25.2 26.6 23.0 25.1 22.8 27.7 23.7 19.5 26.0 20.4 24.0
35.48 25.46 1~.41 16.76 23.69 28.89 21.14 23. I 1 22.43 17.79 20.08 19.66 16.27 17.35 15.44
25.8 23.8 22.0 22.7 20.9 21.8 18.9 20.9 18.9 23.0 19.7 16.1 21.6 16.7 20.0
34.20 24.87 15.25 16.25 22.78 27.67 20.41 22.41 21.82 17.27 19.58 19.07 15.81 16.95 15.09
20.3 18.9 17.6 17.7 16.2 16.8 14.5 16.4 14.7 18.1 15.3 12.6 16.9 13.6 l 6.1
i
l
l
RI2 * 50%R22150%RI42b [] 60%R22/40%R123 t~
~oo0o ~ ~
• 70%R22]30%R141b a 20%R32/80%RI42b
: ~;~R33~5~0%::5422~ b
J /
f
IO°C
/
8OOO
,~ ..~t~-~l~lJIP~ , ~ ' f J ~
o
no'c []l¢c]
~m~
o
E
°~°°
E
6OO0
0
4000
(9
[,- 5(xx) ,-r
o 1000
0.0
0.'2
014
,
,
016
Quality
,
018
1,0
RI2
50%R2 60%R22/ 20%R22/ 70%R22/ 20%R32] 10%R32/ 50%R32/ 40%R32/ R22 50%R142b40%RI23 g0%RI5~l30%R141680%RI42b9~%R152a50%RI42b60%R152a
Fluids
Figure 3 Predicted evaporation heat transfer coefficients of mixture as a function quality for the cooling capacity of 4 kW, T~ of 0°C, and q of 20kWm Figure 3 Coq~h'cients de tran.~[ert de chaleur de mblanges de .fi'igorigc;nes h l'~'vaporation en fonction de la concentration, pour une puissance Jr~eor(~que de 4 kilowatts, T~ = 0 deg C et q = 20 kilowatts m :
Figure 4 Predicted average heat transfer coefficients of refrigerant mixtures over the quality of 20 to 90% for the cooling capacity of 4 kW at various evaporation temperatures, q = 20 kW m 2 d = 8 mm Figure 4 Coefficients de transfert de chaleur moyen de mblanges de frigorig~nes pour des concentrations de 20 h 90%, h diff&entes temperatures d'~vaporation pour une puissanceJHgor!fique de 4 kilowatts, q = 20 kilowatts m 2, d = 8 mm
quality, and mass flow rate were determined for each fluid. Table 2 lists the inlet quality and mass flow rate of refrigerant mixtures for the cooling capacity of 4 kW. In this study, all physical properties required for thermodynamic cycle analysis and for evaluating heat transfer coefficients and pressure drops are obtained from references 18 and 19. Composition is a mass fraction of the more volatile component unless otherwise specified. Throughout the analysis, the inner diameter and length of the evaporator tube were assumed to be 8 mm and 7.96 m, respectively. Figure 3 illustrates local HTCs of various mixtures for 4 kW capacity. All mixtures exhibit a monotonic increase in heat transfer coefficient with quality indicating that nucleate boiling is suppressed. Nucleate boiling contribution in Figure 1 is not present for mixtures, even at very low qualities, due to the loss of wall superheat and the mass transfer resistance as discussed earlier. Since typical evaporator inlet
qualities exceed 20% as listed in Table 2, the main heat transfer mechanism inside the evaporator would not be nucleate boiling, but convective evaporation occurring at the liquid-vapour interface. Figure 4 and Table 3 show the average HTCs over a quality range of 20-90% for the capacity of 4 kW, which are obtained by numerical integration, and ratios of them with respect to the reference fluids of R 12 and R22. The general trend is that the two-phase heat transfer coefficients decrease as evaporation temperature increases. Jung et al. s l0 have shown that when nucleate boiling is suppressed, the heat transfer coefficient is directly proportional to k l ° 6 (Cpl/].ll)°'4frt(pr). As temperature increases, the property term k l °'6 ( C o l / i l l ) °.4 increases while the reduced pressure effect fn(pr) decreases. The effect of the latter, however, is stronger than that of the former, resulting in the decrease in heat transfer coefficients as temperature increases. Rev. Int. Froid 1993 Vo116 No 3
205
Evaporation heat transfer coefficient and pressure drop." D. Jung and R. Radermacher Table 3 Average heat transfer coefficients of pure and mixed refrigerants over 20 90% quality for the cooling capacity of 4 kW at various evaporation temperatures Tableau 3 Co£fficients de transfert de ehaleur moyen de /kigorig~nes purs et en m~lange pour une concentration de 20 h 90% et une puissance /rigor!fiquc de 4 kilowatts, h d(ff~rentes temperatures d'~vaporation -
Pure fluids
10°C
iO°C
0°C
Fluid
HTC
Ratio
t tTC
Ratio
}tTC
Ratio
RI 2 R22 R32 R152a R142b R123 R 141 b
8517 7886 7498 9532 8333 12565 13147
(I,926 0,88 1.119 0.978 1.475 1,543
7392 6916 6781 8306 7152 1(153 t 11287
0.936 0.9t 7 1.124 0.967 1.425 1.527
6490 6147 6230 7319 62 l i ~954 ~)748
0.947 096 1.127 0.957 i,379 !.502
RI2 substitutes
50%R22/50%RI42b 6 0 % R 2 2 / 4 0 % R 123 20% R22/80% R152a 70%R22/30%RI41b 20%R32/80%R142b 10% R32/90%R152a
7227 3915 9422 4268 6389 9191
0.848 0.46 I. 106 0.5 (/,75 1.079
6318 3496 8203 3833 5600 8041
0.855 0.473 1. I 1 0.519 0.758 1.088
%86 3 ! 54 7216 3465 4961 7110
0.861 0.486 I, l I I 0.533 0.764 1,095
R22 substitutes
50%R32/50%R142b 40%R32/60% R152a
5532 7818
0.701 0.991
4974 6928
0.72 1,00
4520 6217
0.735 1.01
(Note that the ratios are calculated based on R12 for pure fluids and RI2 substitutes, and on R22 for R22 substitutes, respectively. The heat transfer coefficient is in W m ~'°C ~)
There is a large variation in heat transfer coefficients among the possible substitutes for R12. For the same capacity, mixtures containing R32 and R152a show 8 10% increase in heat transfer coefficients, which is contrary to the widespread perception that heat transfer coefficients with mixtures are always less than those of R12 or R22. This increase is due to these pure fluids' excellent heat transfer characteristics, which will be further demonstrated later. While mixtures with large volatility difference, 60%R22/40%R123 and 70%R22/30%R141b, show large reduction in heat transfer coefficients as much as 55%, mixtures with moderate volatility difference exhibit 20-30% decrease for the same capacity. The reduction in heat transfer coefficient is caused by three factors: mass transfer resistance, degradation of transport properties, and decrease in mass flow rate. Since nucleate boiling is absent (no bubbles), the loss of wall superheat has negligible effect. Mass transfer resistance, which is still present during convective evaporation, causes the reduction in heat transfer coefficient, which is accounted for by the factor Cme in Equation (1). Note that Cme decreases with larger vapour and liquid composition difference ( Y - X) or larger volatility difference. As refrigerants are mixed, liquid thermal conductivity decreases while liquid viscosity increases. These changes in liquid properties are responsible for the reduction in heat transfer in the liquid layer. The heat transfer degradation due to liquid properties was shown to cause 6080% of the total reduction in HTCs with mixtures when nucleate boiling is suppressed for the same mass flow rateS-~0. Since mass flow rates vary with fluids studied, the effect of mass flow rate on heat transfer should be examined carefully. Local HTCs for all mixtures indicate that nucleation is suppressed. When nucleate boiling is absent, the nucleate boiling contribution in Equation (1), the first term, becomes negligible, and hence h,p becomes directly proportional to th°.8.8 As listed in Table 2, mass flow rates of 60%R22/40%R123 and 70%R22/
206
Int. J. Refrig. 1993 Vo116 No 3
'
I
'
i
' --
I
'
I
~ik
20000 -m RI2, (AHTC=7392 W/m2"C) ,o 50%R22150%R142b,(8875) 18OOO -D 60%R22/40%RI23, (5029) * 20%R22/80%R152a, (14198) 161)00 -m 70%R22/30%R141b, (6025) J j n 20%R32/80%R,42b, (8951) I / ~ M 14000 -A I0%R32/90%R 152a, (14947) Krco~t'-
(13414)
L) 32
~K~.m~ J/_mlm~"
/~llJ
" , ~
-
~
;
~
~
10000
~ ~ o ~°~600(}
41300
2000 0.0
0.2
0.4
0.6
Quality
0.8
1.0
Figure 5 Predicted local and average heat transfer coefficients of refrigerant mixtures over the quality of 20-90% for the mass flow rate of35.48gs ~andheatfluxof2OkWm "at0°C Figure 5 Coefficients de transfert de chaleur moyen et local de mdlanges de Jkigorig~nes pour une concentratiotl de 20 gt 90%, un d~bit masse de 35,48 g s : et un flux thermique de 20 kilowatts rn 2 ~ 0 des C
30%R141b are as much as 44% lower than that of RI2 because of the same capacity requirement. In order to separate the effect of the reduction in mass flow rate on heat transfer degradation from other effects, average heat transfer coefficients are calculated on the same mass flow rate basis. Figure 5 illustrates local and average heat transfer coefficients of various fluids for the mass flow rate of 35.48 g s-n at evaporation temperature of 0°C. Once again, nucleate boiling is suppressed under this condition. For the same mass flow rate, mixtures with large volatility difference (60%R22/40%R123 and 70%R22/ 30%R141b) show as much as 32% reduction in average heat transfer coefficients compared to Rt2, which is caused by the mass transfer resistance and physical
Evaporation heat transfer coefficient and pressure drop. D. Jung and R. Radermacher Table 4
Effect of the degradation of refrigerant side heat transfer coefficient on overall heat transfer coefficient (OHTC) for air and water systems Tableau 4 Effet de la diminution du coefficient de transfert de chaleur c6t~ frigorigene sur le coefficient de transJert de chaleur global, pour des syst~mes h air et 3 eau. Air system RI2
60%R22/40%R123
OHTC 98.7 % degradation
97.2 1.5%
Water system R12 60%R22/40%R123 1595
1275 20%
E
(Typical heat transfer coefficients for air and water sides are 100 and 2000 respectively and heat transfer coefficients for R12 and 60%R22/ 40 %R123 are 7392 and 3496 W m -~°C -~ obtained at 0°C. O H T C is in W m -~°C ')
property change. On the other hand, all other mixtures exhibit as much as 90% increase in average heat transfer coefficients, as shown in Figure 5. Thus, it may be concluded that HTCs of mixtures with large volatility difference are 55% lower than those of R12 for the same capacity. Out of the 55% reduction, mass transfer resistance and property degradation are responsible for 32%, and the remaining 23% is due to reduced mass flow rate to match the capacity. For mixtures with moderate volatility difference (50%R22/50%R142b, 20%R22/ 80%R152a, 20%R32/80%R142b, and 10%R32/ 90%R152a), however, the total reduction in heat transfer coefficients is 20-30% for the same capacity. Since volatility difference is small for these mixtures, the mass transfer resistance and property degradation are small. Therefore, most of the reduction in heat transfer for mixtures with moderate volatility difference stems from reduced mass flow rates to match the capacity.
Impact of the degradation of heat transfer coefficients with mixtures For mixtures, what is more important than the degradation of heat transfer coefficients alone is its impact on overall heat transfer coefficients. When either copper or aluminium tube is used, the thermal resistance for conduction can be neglected, and hence overall heat transfer coefficient can be expressed as U _
l
1
__
1
ho h~ + ho
r12
50%r22! (~)%R221 20%R2~ 70%R22/ 20%R32/ Io%r32f 50%R32/ 40~R32/ 50%R142b 40%R123 80%R152a 30~RI41bS0%RI42b90%RI52a 50~RI42b 60~R152a
R22
Fluids
Figure 6 Predicted pressure drop of refrigerant mixtures over the quality of 20-90% for the cooling capacity of 3 kW at various evaporation temperatures, q = 15 kW m ~, d = 8 m m Figure 6 Perte de charge estim~e pour des mklanges de frigorig~nes, avec une concentration de 20 h 90%, une puissance/kigor(fique de 3 kilowatts, ~ diff~rentes tempOratures d'Ovaporation, q = 15 kilowatts rn 2, d = 8 m m
While the maximum reduction in overall HTC for mixtures with large volatility difference is only 1.5% for an air system, it becomes as much as 20% for a water system. This clearly explains the conflict between the observation of Stoecker 5 (water system) and that of another report 6 (air system) mentioned earlier. It may be concluded that considerable reduction in HTC for mixtures with a large gliding temperature difference (or volatility difference) may or may not be of significance, depending upon the heat transfer fluid. If the heat transfer fluid side resistance is an order of magnitude larger than that of the refrigerant side (air system), its overall effect is small. However, when the heat transfer fluid side resistance is the same order of magnitude (water system), the reduction in heat transfer coefficient associated with mixtures would significantly affect the overall heat transfer coefficient.
Pressure drop of mixtures
hi
(3)
+ ho
where h~ and ho are the refrigerant and heat transfer fluid side heat transfer coefficients, respectively. There are two popular heat transfer fluids found in normal HVAC systems: air and water. When air is used as heat source or sink fluid, the typical heat transfer coefficient on the air side is less than 100 Wm-2K '. On the other hand, when water or its mixture with antifreeze is used as a heat transfer fluid, heat transfer coefficients fall in the range of 1000-2000 Wm-2K-L Using these typical values, a comparison will be made with a mixture of 60%R22/40%R123 with large volatility difference against RI2. For the capacity of 4 kW, the heat transfer coefficients for R12 and 60%R22/40%R123 mixture at evaporation temperature of 0°C are 7392.0 and 3496.0 W m-2K 2, respectively. Table 4 lists overall HTCs obtained with these values for two heat transfer fluids.
Figure 6 illustrates the predicted pressure drop of various refrigerants for the cooling capacity of 3 kW. As with the heat transfer coefficient, the pressure drop decreases as temperature increases. One point clearly seen in Figure 6 is that the fluids with larger HTCs also have larger pressure drops, which is similar to a single-phase turbulent flow situation. All mixture substitutes for R12 and R22 show smaller pressure drops than those of Ri2 and R22 due mainly to the reduced mass flow rates to maintain the same capacity. Effect of uncertainties in transport properties on heat transfer prediction Since any transport correlations require physical properties, their accuracy greatly depends upon the accuracies of physical properties. There are well-established thermodynamic properties of both pure and mixed refrigerants and their uncertainties are very small. There, howRev. Int. Froid 1993 Vo116 No 3
207
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher Table 5 Effect of uncertainties in transport properties on average heat transfer coelticient prediction Tableau 5 Eff('ct des incertitudes dans les propri~tes de transport sur la pr~vision du co£~cient moyen de tran.~/~,rt de chah'ur
R 12 HTC
60% R22/40% R [ 23 dO
HTC
5(t'% R22/50%R142b •
3496
1{T("
(13
Unchanged
7392
55~.6
10% increase in VL KL VV ST All
7069 7820 7448 7392 7534
- 4.37 5.8 0.76 0.0 1.92
3422 3342 3700 3496 3564
4.43 5.82 0,77 0,0 1,92
5~ 50 5(h~4 5~Q5 55~5 %90
~4,22 5.69 0,07 0,0 1.86
10% decrease in VL KL VV ST All
7767 6947 7330 7393 7238
5.1 - 6.0 - 0.83 0.0 2.07
3676 3284 3467 3497 3423
5.14 ~. 1 0.85 0.0 - 2.1
5*65 5250 5~3(1 5587 5474
50 6.(! 1.(~ !).0 • 2,0
(VL, VV, KL, ST, and • are viscosities of liquid and vapour, thermal conductivity of liquid, surface tension, and percent increase or decrease in HTC as compared to the HTC obtained without changing properties respectively. HTCs are predicted at 0°C for 4 kW capacity and are in W m : °C ~)
ever, is a lack of transport property data of pure fluids, not to mention mixtures. To make matters even worse, a large scatter is often found with mixture data, as revealed elsewhere '9. Therefore, it would be useful to carry out a sensitivity analysis for studying the effect of transport properties on heat transfer prediction. For this purpose, viscosities, thermal conductivities of liquid and vapour, and surface tension are varied from 90 to 110% of the measured data used in the above analysis. Table 5 lists heat transfer coefficients and percent changes in heat transfer coefficients qb of pure and mixed refrigerants as properties change. For both pure and mixed refrigerants, HTCs are most sensitive to the viscosity and thermal conductivity of liquid. With 10% uncertainty in both viscosity and thermal conductivity, 6% change in HTCs is observed for both pure and mixed refrigerants. Heat transfer coefficients are not sensitive to other properties.
Conclusions Jung et al.'s correlations for heat transfer coefficient and pressure drop are used to predict HTCs and pressure drops of prospective mixtures to replace RI2 and R22. To compare working fluids on a fair basis, comparisons are made based on the same cooling capacity basis. For this purpose, a basic vapour compression cycle was modelled to calculate required mass flow rates and evaporator inlet qualities for all fluids studied. The following conclusions can be drawn from this study: 1.
2.
208
In the evaporating temperature range of - 1 0 to 10°C, the evaporator inlet quality is usually greater than 20% for all fluids considered. In the quality range greater than 20%, nucleate boiling is fully suppressed for all pure and mixed refrigerants. The transition qualities for mixtures are lower than those of pure fluids due to the loss of wall superheat and mass transfer resistance. In this convective evaporation region, the heat transfer coefficient increases monotonically with quality. For the same mass flow rate, certain mixtures containing R32, R152a and R142b exhibit as much as Int. J. Refrig. 1993 Vo116 No 3
90% higher HTCs than RI2 and R22. For the same capacity, however, mixtures containing R32 and R152a show only 8 10% increase in HTCs because of reduced mass flow rates. For the same capacity, mixtures with large volatility difference show up to 55% reduction compared to Rl2 and R22, which is caused by mass transfer resistance and property degradation due to mixing (32%) and reduced mass flow rates (23%). On the other hand, mixtures with moderate volatility difference exhibit 20-30% reduction, which is mainly due to reduced mass flow rates to match the capacity. . The significance of the reduction in HTCs for mixtures with large volatility difference depends upon the heat transfer fluids. If the major resistance to heat flow exists in the heat transfer fluid side (air system), the reduction in overall HTC is less than t.5%. However, when the resistance in the heat transfer fluid side is the same order of magnitude (water system), as much as 20.0% reduction in overall HTC is expected. . For a given fluid, pressure drop becomes smaller as evaporation temperature decreases. For the same capacity, pressure drops of all mixtures considered are smaller than those of RI2 and R22 because of reduced mass flow rates. . Heat transfer prediction is more sensitive to the properties of liquid than those of vapour. With typical uncertainties of 10% in transport properties, HTCs changed less than 6%.
Acknowledgement The funding for this work was provided by the US Environmental Protection Agency and Whirlpool Corporation. The authors acknowledge D. Didion, M. McLinden and G. Morrison at the National Institute of Standards and Technology for their helpful comments and discussions throughout the study. References I
EPA report CFCs and Stratospheric Ozone United States Environmental Protection Agency (December 19873
Evaporation heat transfer coefficient and pressure drop. D. Jung and R. Radermacher 2
3
4 5 6 7
8 9 l0
Montreal Protocol on Substances That Deplete the 'Ozone Layer Final Act, United Nations Environment Programme (1987) Malroy, W., Kauffeld, M., McLinden, M., Didion, D. Experimental evaluation of two refrigerant mixtures in a breadboard air conditioner Proc lnt lnst Refrigeration, Purdue Conference on CFCs Commissions BI, B2, E1 & E2 (1988) 27-34 Kruse, H. The advantages of non-azeotropic refrigerant mixtures for heat pump application Int J Refrig (1981) 4 119-125 Stoecker, W.F. Improving the Energy Effectiveness of Domestic R£/r~gerators by the Application ~tf Refrigerant Mixtures ORNL report, ORNL/Sub-78/55463/1 (1978) ORNL report An Evaluation o/'a Two-Evaporator Re/'rigeratorFreezer Using Nonazeotropic Refrigerant Mixtures ORNL/Sub/ 82-47952/1 (1982) Jung, D.S., Radermacher, R. Prediction of heat transfer coefficient and pressure drop of various refrigerants (submitted to ASHRAE Trans May 1990) Jung,D.S., McLinden, M., Radermacher, R., Didion, D. Horizontal flow boiling heat transfer experiments with a mixture of R22/R114 lnt J Heat Mass Transfer (1989) 32 (1) 131 145 Jung, D.S., MeLinden, M., Radermaeher,R., Didion, D. A study of flow boiling heat transfer with refrigerant mixtures lnt J Heat Mass Tran.~/br (1989) 32 (9) 1751 1764 Jung, D.S. Mixture effects on horizontal convective boiling heat transfer PhD thesis University of Maryland (May 1988) (Electric Power Research hzstitute Report EPRI ER-6364 Project 8006-2)
1[
12 13
14
15 16
17
18
19
Unal, H.C. Prediction of nucleate boiling heat transfer coefficients for binary mixtures lnt J Heat and Mass Transjer (1987) 29 (4) 637-640 Stephan, K. Heat transfer in boiling mixtures Proc 7th lnt Heat Transfer Conf Munich (1982) Chen, J.C. Correlation for boiling heat transfer to saturated fluids in convective flow lnd Eng Chem Process Design Develop (1966) 5 (3) Stephan, K., Abdelsalam, M. Heat transfer correlations for natural convection boiling lnt J Heat and Mass Transfer (1980) 23 73 87 i'~ Dittus, F.W., Boelter, L.M.K. Public Engineering Vol 2 University of California at Berkeley (1930) 443 Jung, D.S. and Radermaeher, R. Prediction of pressure drop during horizontal annular flow boiling of pure and mixed refrigerants Int J Heat Mass TransJer (1989) 32 (12) 2435 2446 Eekels, S.J., Pate, M.B. An experimental comparison of evaporation and condensation heat transfer coefficients for HFC-134a and CFC-12 (submitted to lnt J ReJ~ig 1990) Morrison, G., McLinden, M. Application of a Hard Sphere Equation of State to Refrigerants and Re[rigerant Mixtures NBS Technical Note 1226, NBS, Gaithersburg, MD (1986) Jung, D.S., Radermacher, R. Transport properties and surface tension of pure and mixed refrigerants (submitted to A S H R A E Trans April 1990)
Rev. Int. Froid 1993 Vo116 No 3
209
A study on the prediction of heat transfer coefficient and pressure drop of refrigerant mixtures is reported. Heat transfer coefficients and pressure drops of prospective mixtures to replace R12 and R22 are predicted on the same cooling capacity basis assuming evaporation in horizontal tubes. Results indicate that nucleate boiling is suppressed at qualities greater than 20% for all mixtures, and evaporation becomes the main heat transfer mechanism. For the same capacity, some mixtures containing R32 and R 152a show 8 10% increase in heat transfer coefficients. Some mixtures with large volatility difference exhibit as much as 55% reduction compared to RI2 and R22, caused by mass transfer resistance and property degradation due to mixing (32%) and reduced mass flow rates (23%). Other mixtures with moderate volatility difference exhibit 2030% degradation due mainly to reduced mass flow rates. The overall impact of heat transfer degradation, however, is insignificant if major heat transfer resistance exists in the heat transfer fluid side (air system). If the resistance in the heat transfer fluid side is of the same order of magnitude as that on the refrigerant side (water system), considerable reduction in overall heat transfer coefficient of up to 20% is expected. A study of the effect of uncertainties in transport properties on heat transfer shows that transport properties of liquid affect heat transfer more than other properties. Uncertainty of 10% in transport properties causes a change of less than 6% in heat transfer prediction. (Keywords:heat transfer coefficientprediction;alternativerefrigerants;refrigerantmixtures)
Prrvision du coefficient de transfert de chaleur et de la perte de charge lors de l'rvaporation de mrlanges de frigorigrnes dans des tubes horizontaux Cette ktude porte sur l'estimation du coefficient de transfert de chaleur et de la perte de charge pour des mklanges de Jkigorigbnes, substituts potentiels du RI2 et du R22 pour la mdme capacitb fr~gor(fique et bvaporation en tubes horizontaux. Les rbsultats indiquent que l'bbullition nuclkke est supprimbe pour tousles mblanges prbsentant des concentrations supbrieures h 20% et que l'bvaporation devient alors le mbcanisme principal du transfert de chaleur. A puissance bgale, quelques mklanges contenant du R32 et du R152a montrent u n e augmentation des coefficients de transfert de chaleur de 8~10%. Pour quelques mdlanges de produits gt volatilitbs trbs d,iffkrentes, on observe une rbduction de 55% par rapport au R12 et au R22, causde par la rbsistance au transfert de masse et la dbgradation des propribtbs dues au mHange (32%) et gl la rbduction des dbbits masse (23%). D 'autres mblanges, ayant une moindre difference de volatilitk, montrent une dbgradation de 20-30% due surtout h la rbduction des dbbits masse. Nbanmoins, l'effet global de la rbduction du tramfert de chaleur est insignifiant si la rbsistance au transfert de chaleur est plus forte du c6tb du fluide qui transJOre la chaleur (systbme h air). Si la rdsistance du cdtb du fluide de transfert de chaleur est du m~me ordre de grandeur que celle du crtb du frigorig~ne (systkme h eau), on pr~voit une rbduction considbrable du coefficient de tramfert de chaleur global s'Nevant h 2(t%. Une ktude sur l'effet des incertitudes clans les propribtks de transfert sur le transfert de chaleur montre que les propribtbs de transfert du liquide influencent plus Jbrtement le tran~fert de chaleur que d'autres propribtbs. Une incertitude de 10% dans les propribtbs de transfert modifie de moins de 6% les prbvisions du transfert de chaleur. (Mots clrs: prbvision du coefficient de transfert de chaleur; frigorigrne de remplacement; mrlange de frigorigrnes)
Fully halogenated chlorofluorocarbons (CFC) are responsible for destroying the stratospheric ozone layer and increasing global warming 1. At present, research is being undertaken at many institutions to replace fully halogenated CFCs such as R l l , R12, R I 3 B I , Rl13, R l I 4 and R115, of which trade and production are regulated by the Montreal Protocol z. Thanks to extensive research during the past years, new ozone-safe fluids such as 0140-7007/93/030201~)9 ~ 1993 Butterworth Heinemann Ltd and IIR
R134a, R141b, R142b, R143a, R123 and R124 have been developed. These new fluids, however, are to be tested for certain requirements such as toxicity and material compatibility. In replacing ozone-depleting substances, one of the basic properties to be satisfied by any substitute fluids is vapour pressure. As vapour pressure changes, so does volumetric capacity. A change in volumetric capacity
Rev. Int. Froid 1993 Vol 16 No 3
201
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher
Nomenclature
G
D F .!'
G h k L rn N P q T X X
Y U
specific heat (J kg ~K- ~) tube diameter (m) heat transfer enhancement factor friction factor mass flux (kg m 2s-t) heat transfer coefficient (W m 2 K ~) latent heat of vaporization (kJ kg-t) thermal conductivity (W m-~ K 1) tube length (m) mass flow rate (g s-~) factor due to nucleate boiling pressure (kPa) heat flux (Win 2) temperature (K or °C) liquid phase concentration quality vapour phase concentration overall heat transfer coefficient (W m-2 oC 1)
implies a change in compressor, a key component in HVAC equipment. Newly developed ozone-safe fluids may not provide the desired vapour pressure of the fluids to be replaced. An easy way of manipulating vapour pressure to obtain desired pressure is using mixtures. For instance, vapour pressure of R12 can be obtained with R22/R142b mixture, of which R22 and R142b are more and less volatile than the target fluid, R12, respectively. Non-azeotropic refrigerant mixtures (NARMs) are not only able to meet the vapour pressure requirement, but also increase the system performance by utilizing a gliding temperature effect, as evidenced by some studies 3,4. Since the introduction of N A R M S in the HVAC area, there has been an unresolved issue of degradation of heat transfer coefficients (HTC) associated with NARMs. For instance, Stoecker 5 tested a prototype refrigerator with R22/R114 mixtures. In his experiments, the heat transfer fluid employed was not air but a mixture of ethylene glycol and water. Stoecker has observed a noticeable decrease in HTCs and degradation of system performance. Another report, however, revealed that the degradation of heat transfer coefficients associated with R22/R114 mixtures is not significant enough to measure its effect on refrigerant-to-air heat exchange in an evaporator, which was based on actual evaporator test dataL Since N A R M s are prospective substitutes for ozonedepleting fluids, their heat transfer characteristics should be clarified to help engineers design proper heat exchangers. Experimental investigation would be the most reliable means to fulfil this goal, even if it would be cost- and time-consuming. As an interim solution toward better understanding of heat transfer with mixtures, a study on the prediction of heat transfer coefficients and pressure drops, based on reliable experimental data, might be useful. This report succeeds our previous study on the prediction of heat transfer coefficients of newly marketed pure refrigerants v. The goal of this paper is to predict and compare heat transfer coefficients of refrigerant mixtures 202
Int. J. Refrig. 1993 Vo116 No 3
p cr /z q5
density (kg m ~) surface tension (N m ~1 viscosity (Pa s) pressure drop multiplier
Dimensionless numbers Bo Pr Re ~u
Boiling number, q/Gh/~. Prandtl number of liquid, ~)~t/k Reynolds number, GD/p Martinelli's parameter in Table 1
Subscripts cec fo I Io nbc tp v
convective evaporation contribution total flow assumed as liquid liquid liquid only nucleate boiling contribution two-phase vapour
during evaporation, and to determine the impact of heat transfer degradation with mixtures on overall heat transfer coefficients. Pressure drop will also be predicted. Finally, the effect of uncertainties of transport properties on heat transfer prediction will be examined.
Analysis Selection of correlations To accomplish the objectives, correlations developed by Jung et al) -1° are used. Jung et al. measured more than 3000 local heat transfer coefficients and 400 pressure drops of R22, RI14, R12, R152a and their mixtures at various concentrations during evaporation. Their test section was made of a stainless steel tube of 9. I mm id and 8.0 m long. The test section was heated by passing a direct current to obtain data within a desired quality range. Data were taken at reduced pressure of 0.08 to 0.16 at the exit of the test section. Applied heat flux was from 10.0 to 45.0 kW m -z, while mass flow rate of refrigerants varied from 16.0 to 46.0 g s -1, equivalent to 250720 kg m 2 s-~. The exit quality varied from 40 to 95%. The overall compositions were 0.00, 0.21, 0.60, 0.89 and 1.00 mole fraction R12 for RI2/RI52a mixtures, and 0.00, 0.23, 0.47, 0.77 and 1.00 mole fraction R22 for R22/ R114 mixtures. These ranges and physical dimension of the test section are typical of air-source residential heat pumps of 3-5 kW capacity. Figure 1 illustrates local HTCs obtained at various heat fluxes for a given mass flow rate. Heat transfer coefficients are strongly dependent upon heat flux at low qualities. This heat transfer regime is termed a partial boiling regime, which is characterized by the formation of bubbles on the wall. As quality increases further, the annular liquid film becomes thinner so that heat transfer becomes more effective. At last, the wall superheat (Twall--Tsar) becomes smaller than the threshold wall superheat to sustain nucleation on the heating surface. This is called suppression of nucleate boiling. This phenomenon of cessation of nucleation is characteristic of flow
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher
Partial boiling Convective regime regime ~Liquid layer~ (( ~..-Vapor core.(_l~")")
e-
Z UJ W_ OIJ.l
o (3
U.I i1 CO
q5~ Increasing heat flux q 4
T
~
~_~ ~ nnbe
q 3 ~ :
Tw
t
Bubbles ~ /
2/ i
ILl
T
ATeff
ATsa t
~B fDew --~,./ line
hcec
l I
Z
< re"
ql /
2hce c
cee'
Tb a
U.I
I
Loss of available
QUALITY Figure 1 Nucleateboilingand convectiveevaporationcontributions in two-phaseheat transfer coefficient,h.~, nucleateboiling contribution; hc~,convectiveevaporationcontribution
I I
0
Figure 1 Parts de l~bullition nuclide et de l'kvaporation par convection sur le coefficient de traasfert de chaleur diphasique, h.bc ~bullition nuclide, hce~ ~vaporation par eonvection
boiling, which is markedly different from nucleate pool boiling with continual bubble generation on the wall. The quality at which nucleation ceases is called the transition quality beyond which convective evaporation occurs at the liquid-vapour interface (no bubbles are present). Heat transfer coefficients in the convective region increase with quality monotonically. Nucleate boiling was fully suppressed for all pure and mixed refrigerants tested. The quality at which nucleation ceases, however, is much lower for mixtures than for pure fluids. This is explained by examining local bubble dynamics. Bubble growth in a single component system is limited by the rate of heat diffusion to the bubble interface to provide latent heat of vaporization. In a mixture system, however, another limitation exists as a result of stripping of the more volatile component from the liquid region close to the bubble, which is caused by the volatility (or boiling point) difference of the pure components. Due to the concentration gradient near the interface, through which the more volatile component has to diffuse, mass transfer resistance to heat transfer is introduced, since the heat transfer is affected by the rate at which the more volatile component can be replenished at the interface. Figure 2 illustrates the bubble growth process of a binary mixture on a phase equilibrium diagram. The initial bulk liquid at Xb is superheated by an amount of ATsa, above the boiling point Tb, corresponding to the initial bulk liquid composition Xb. At the bubble interface the liquid phase composition is moved to X~ (point A) while the vapour composition within the bubble shifts to Y~ (point B). The consequence of this liquid phase composition shift toward the less volatile component at the interface is that the effective temperature driving potential becomes Tw-T~, less than the nominal value Tw-Tb, which is known as the loss of available wall superheat. Both mixture effects, the mass transfer resistance and the loss of wall superheat, have been shown to be responsible for slower-growing bubbles and smaller bubble departure size associated with mixtures, which in turn
Bubble i line
superheat X i
X b
Yi
Yb
COMPOSITION Figure 2 Bubble growth process of a binary mixture illustrated on a phase equilibrium diagram Figure 2 Dkveloppement de l'dbullition commenfante d'un mdlange binaire sur un diagramme d'dquilibre de phases
are related to the reduction in heat transfer coefficients with mixtures for nucleate pool boiling]U 2. The mixture effects become greater as the volatility difference between the pure components increases. This is the reason why nucleate boiling is suppressed at much lower qualities with mixtures than with pure fluids. Since experimental results supported Chen's u idea that the two-phase heat transfer coefficient is additive of nucleate boiling and convective evaporation contributions, the original form of Chen's correlation, Equation (1), was retained to correlate the measured heat transfer data '3. More detailed analysis is available elsewhere ~0. The final correlation for mixed refrigerants becomes htp=
hnbc + h~cc =
Nhun/Cun + CmeFph]o
(l)
where hnbc, h.... N, hon, Cu,, Cmc, F and h~o are a nucleate boiling contribution, a convective evaporation contribution, a nucleate boiling suppression factor, a pool boiling heat transfer coefficient obtained by UnaP ', a correction factor for mixtures in pool boiling obtained by UnaF ~, another correction factor for mixtures in evaporation, a two-phase enhancement factor, and a single-phase heat transfer coefficient for liquid-only flow obtained by Dittus and Boelter correlation,~, respectively. Table 1 lists various equations to calculate the parameters in Equation (1). The correlation shows good agreement with the experimental data yielding a mean deviation of 9.6% for R22/ R114 and R12/R152a mixtures at various concentrations under a wide range of conditions. The good agreement stems mainly from an excellent fit of Stephan and Abdelsalam's ~2nucleate pool boiling heat transfer correlation for pure refrigerants used in Unal's equation'L The heat flux dependence of N, the nucleate boiling suppression factor, is also well accounted for. Besides this good fit, the most distinctive feature of Equation (1) is that it works for both pure and mixed refrigerants. For pure
Rev. Int. Froid 1993 Vo116 No 3 203
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher 'Fable I S u m m a r y o f heat transfer coefficient correlation ['or refrigerant mixtures b y J u n g el al.' '" Tableau l Corr,;/alien entre h's coeffl'cients de transji, rt de chah, ur pour des m~;langes de [?igorig~nes, par Jung c t a l : ~' h, r = h .... + h.. = (,,ilh""
~ C.
f'~ h;.
N - 4048X,,~:Bo~"(l'orX,~ < I),N - 2.0 ().IX, ~ B o (',,, = [1.0 + (h: + bd(l.0 + b4)](1.0 ~ b.)
0.33(1`O~I ~
h: = (1.0 - A') ln[(l.01 - .V)/(1.OI - Y)] + X l n ( X / Y ) + ]Y h, = 0 . 0 ( f o r X >/ 0.01), h, = ( Y / X ) ('~ 1.0 (for X < 0.01) h4 -
152 (p/t~,,,,9 °''''.
X/Y
=
1.0 f o r ) ; -
} -
5)
A}'
0.0
where p,,,,, is the critical pressure of the more volatile component, h,,,, = h,/Cu,, =
h~ h~
/,,
x, ii X~ = h,x~ • ~,ix, h
h:
where h~ and h: are nucleate pool boiling heat transfer coefficients of c o m p o n e n t s I and 2 calculated by Stephan and Abdelsalam's correlation, h ....
hs. = 207k~ilqbd ].... (~) ..... where bd = 0.0146 fl[2o~g(p~ - p,)]~ with a contact angle/3 = 35 °. ~3",,,~ =
1.0
- 0.351Y -
Xp ~',
F = 2.37 (0.29 + XII).... =
h,,, :
(! ; ,)" 0 0~3 • --
"'
k, [di.l -
JL
..-)dl"" tC.,~"'
-~.----J
tW-)
fluids, the mixture correction factors Cun and Cm~ become 1.0, rendering Equation (!) to be identical to the correlation for pure fluids. Pressure drops were correlated following Martinelli and his co-workers (detailed information is available elsewhere~6). The final correlation for pressure drop for mixtures is
capacity basis, R l34a showed 33.3% increase in average heat transfer coefficient over RI2. The prediction by Equation (1) was 34.4% increase in average heat transfer coefficient of R 134a over R 12, showing good agreement. These comparisons with pure fluids indirectly substantiated the validity of correlations (1) and (2).
Heat trans/~r coefficients o f mixtures D ,o I
~
"i
(2)
where ~b,~ = 12.82 X, '47(1 - x ) ~-s and the friction factor Jio = 0.046 Re °.2. The mean deviation was 8.4% for all mixture data. Because no heat transfer coefficients and pressure drops for mixtures have been reported since correlations (1) and (2) were developed, further validity of correlations (1) and (2) could not be checked. As an indirect confirmation, however, more pure fluids' data were taken and compared. Jung obtained more than 300 local heat transfer coefficients and 80 pressure drops from R ll using the same experimental apparatus under similar condition. The mean deviations between the data and correlations (1) and (2) were 6.3% and 10.0% respectively 7. Recently, Eckels and Pate ~7 measured evaporation heat transfer coefficients for R134a in a horizontal tube of 8.0 m m id and 3.67 m long. Evaporation tests were performed at temperatures of 5-150C with inlet and exit qualities of 10% and 90%, respectively. The mass flux varied from 125 to 400 kg m-2s-~. Average heat transfer coefficients over the entire quality range were measured with an appropriate energy balance using a log mean temperature difference method. On the same cooling
204
Int. J. Refrig. 1993 Vol 16 No 3
Before comparing fluids for their heat transfer characteristics, a basis for fair comparison should be established. One way of comparing fluids would be on the basis of same mass flow rate at the same evaporating temperature. This approach would be unfair considering that what is actually required of each fluid is not the same flow rate, but the same capacity. Since the heat of evaporation differs greatly among many fluids at the same temperature, the required mass flow rate would vary accordingly among different fluids for desired capacity. Based on this argument, comparison will be made on the same cooling capacity basis. To accomplish this goal, a basic refrigeration cycle was modelled for various fluids of interest. In the model, no pressure drops were assumed in the evaporator and condenser. The compressor was specified by an isentropic compressor efficiency of 0.7. Since evaporation and condensation with mixtures occur over a range of temperatures (gliding temperature effect), evaporator and condenser temperatures at 50% quality are specified assuming a linear temperature glide over an entire twophase region. Under the fixed cooling capacity of 3 a n d 4 kW and condenser temperature of 40°C, evaporation temperature varied from - 10.0, 0.0 and to 10.0"C. The net refrigeration effect per unit mass, evaporator inlet
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher Table 2 Mass flow rate and evaporator inlet quality, rn (g s ') and x~,, (%) for pure and mixed refrigerants for the cooling capacity of 4 kW at different evaporation temperatures Tableau 2 Dbbit masse et concentration ?l I'entr~e de I'bvaporateur, m (g s ~) et .v,°. ( % ) pour des [?igorigbnes purs et en mblange et pour une puissance jHgor(fique de 4 kilowatts, 71 d(ff~'rentes temperatures d'~vaporation - IO°C Fluid
R 12 R22 R32 R 152a R 142b R 123 Rl41b 50% R22/50% R 142b 60% R22/40% R 123 20% R22/80% R 152a 70% R22/30% R 141 b 21)% R32/80% R 142b 10%R32/90%R152a 50% R32/50 % R 142b 40% R32/60% R 152a
J
1 I(KI0
9000
• d7000
O°C
nl
Xin
~l
Xin
hl
Xin
36.91 26.13 15.63 17.32 24.69 30.21 21.94 23.89 23.10 18.36 20.63 20.30 16.77 17.78 15.87
31.0 28.3 25.9 27.2 25.2 26.6 23.0 25.1 22.8 27.7 23.7 19.5 26.0 20.4 24.0
35.48 25.46 1~.41 16.76 23.69 28.89 21.14 23. I 1 22.43 17.79 20.08 19.66 16.27 17.35 15.44
25.8 23.8 22.0 22.7 20.9 21.8 18.9 20.9 18.9 23.0 19.7 16.1 21.6 16.7 20.0
34.20 24.87 15.25 16.25 22.78 27.67 20.41 22.41 21.82 17.27 19.58 19.07 15.81 16.95 15.09
20.3 18.9 17.6 17.7 16.2 16.8 14.5 16.4 14.7 18.1 15.3 12.6 16.9 13.6 l 6.1
i
l
l
RI2 * 50%R22150%RI42b [] 60%R22/40%R123 t~
~oo0o ~ ~
• 70%R22]30%R141b a 20%R32/80%RI42b
: ~;~R33~5~0%::5422~ b
J /
f
IO°C
/
8OOO
,~ ..~t~-~l~lJIP~ , ~ ' f J ~
o
no'c []l¢c]
~m~
o
E
°~°°
E
6OO0
0
4000
(9
[,- 5(xx) ,-r
o 1000
0.0
0.'2
014
,
,
016
Quality
,
018
1,0
RI2
50%R2 60%R22/ 20%R22/ 70%R22/ 20%R32] 10%R32/ 50%R32/ 40%R32/ R22 50%R142b40%RI23 g0%RI5~l30%R141680%RI42b9~%R152a50%RI42b60%R152a
Fluids
Figure 3 Predicted evaporation heat transfer coefficients of mixture as a function quality for the cooling capacity of 4 kW, T~ of 0°C, and q of 20kWm Figure 3 Coq~h'cients de tran.~[ert de chaleur de mblanges de .fi'igorigc;nes h l'~'vaporation en fonction de la concentration, pour une puissance Jr~eor(~que de 4 kilowatts, T~ = 0 deg C et q = 20 kilowatts m :
Figure 4 Predicted average heat transfer coefficients of refrigerant mixtures over the quality of 20 to 90% for the cooling capacity of 4 kW at various evaporation temperatures, q = 20 kW m 2 d = 8 mm Figure 4 Coefficients de transfert de chaleur moyen de mblanges de frigorig~nes pour des concentrations de 20 h 90%, h diff&entes temperatures d'~vaporation pour une puissanceJHgor!fique de 4 kilowatts, q = 20 kilowatts m 2, d = 8 mm
quality, and mass flow rate were determined for each fluid. Table 2 lists the inlet quality and mass flow rate of refrigerant mixtures for the cooling capacity of 4 kW. In this study, all physical properties required for thermodynamic cycle analysis and for evaluating heat transfer coefficients and pressure drops are obtained from references 18 and 19. Composition is a mass fraction of the more volatile component unless otherwise specified. Throughout the analysis, the inner diameter and length of the evaporator tube were assumed to be 8 mm and 7.96 m, respectively. Figure 3 illustrates local HTCs of various mixtures for 4 kW capacity. All mixtures exhibit a monotonic increase in heat transfer coefficient with quality indicating that nucleate boiling is suppressed. Nucleate boiling contribution in Figure 1 is not present for mixtures, even at very low qualities, due to the loss of wall superheat and the mass transfer resistance as discussed earlier. Since typical evaporator inlet
qualities exceed 20% as listed in Table 2, the main heat transfer mechanism inside the evaporator would not be nucleate boiling, but convective evaporation occurring at the liquid-vapour interface. Figure 4 and Table 3 show the average HTCs over a quality range of 20-90% for the capacity of 4 kW, which are obtained by numerical integration, and ratios of them with respect to the reference fluids of R 12 and R22. The general trend is that the two-phase heat transfer coefficients decrease as evaporation temperature increases. Jung et al. s l0 have shown that when nucleate boiling is suppressed, the heat transfer coefficient is directly proportional to k l ° 6 (Cpl/].ll)°'4frt(pr). As temperature increases, the property term k l °'6 ( C o l / i l l ) °.4 increases while the reduced pressure effect fn(pr) decreases. The effect of the latter, however, is stronger than that of the former, resulting in the decrease in heat transfer coefficients as temperature increases. Rev. Int. Froid 1993 Vo116 No 3
205
Evaporation heat transfer coefficient and pressure drop." D. Jung and R. Radermacher Table 3 Average heat transfer coefficients of pure and mixed refrigerants over 20 90% quality for the cooling capacity of 4 kW at various evaporation temperatures Tableau 3 Co£fficients de transfert de ehaleur moyen de /kigorig~nes purs et en m~lange pour une concentration de 20 h 90% et une puissance /rigor!fiquc de 4 kilowatts, h d(ff~rentes temperatures d'~vaporation -
Pure fluids
10°C
iO°C
0°C
Fluid
HTC
Ratio
t tTC
Ratio
}tTC
Ratio
RI 2 R22 R32 R152a R142b R123 R 141 b
8517 7886 7498 9532 8333 12565 13147
(I,926 0,88 1.119 0.978 1.475 1,543
7392 6916 6781 8306 7152 1(153 t 11287
0.936 0.9t 7 1.124 0.967 1.425 1.527
6490 6147 6230 7319 62 l i ~954 ~)748
0.947 096 1.127 0.957 i,379 !.502
RI2 substitutes
50%R22/50%RI42b 6 0 % R 2 2 / 4 0 % R 123 20% R22/80% R152a 70%R22/30%RI41b 20%R32/80%R142b 10% R32/90%R152a
7227 3915 9422 4268 6389 9191
0.848 0.46 I. 106 0.5 (/,75 1.079
6318 3496 8203 3833 5600 8041
0.855 0.473 1. I 1 0.519 0.758 1.088
%86 3 ! 54 7216 3465 4961 7110
0.861 0.486 I, l I I 0.533 0.764 1,095
R22 substitutes
50%R32/50%R142b 40%R32/60% R152a
5532 7818
0.701 0.991
4974 6928
0.72 1,00
4520 6217
0.735 1.01
(Note that the ratios are calculated based on R12 for pure fluids and RI2 substitutes, and on R22 for R22 substitutes, respectively. The heat transfer coefficient is in W m ~'°C ~)
There is a large variation in heat transfer coefficients among the possible substitutes for R12. For the same capacity, mixtures containing R32 and R152a show 8 10% increase in heat transfer coefficients, which is contrary to the widespread perception that heat transfer coefficients with mixtures are always less than those of R12 or R22. This increase is due to these pure fluids' excellent heat transfer characteristics, which will be further demonstrated later. While mixtures with large volatility difference, 60%R22/40%R123 and 70%R22/30%R141b, show large reduction in heat transfer coefficients as much as 55%, mixtures with moderate volatility difference exhibit 20-30% decrease for the same capacity. The reduction in heat transfer coefficient is caused by three factors: mass transfer resistance, degradation of transport properties, and decrease in mass flow rate. Since nucleate boiling is absent (no bubbles), the loss of wall superheat has negligible effect. Mass transfer resistance, which is still present during convective evaporation, causes the reduction in heat transfer coefficient, which is accounted for by the factor Cme in Equation (1). Note that Cme decreases with larger vapour and liquid composition difference ( Y - X) or larger volatility difference. As refrigerants are mixed, liquid thermal conductivity decreases while liquid viscosity increases. These changes in liquid properties are responsible for the reduction in heat transfer in the liquid layer. The heat transfer degradation due to liquid properties was shown to cause 6080% of the total reduction in HTCs with mixtures when nucleate boiling is suppressed for the same mass flow rateS-~0. Since mass flow rates vary with fluids studied, the effect of mass flow rate on heat transfer should be examined carefully. Local HTCs for all mixtures indicate that nucleation is suppressed. When nucleate boiling is absent, the nucleate boiling contribution in Equation (1), the first term, becomes negligible, and hence h,p becomes directly proportional to th°.8.8 As listed in Table 2, mass flow rates of 60%R22/40%R123 and 70%R22/
206
Int. J. Refrig. 1993 Vo116 No 3
'
I
'
i
' --
I
'
I
~ik
20000 -m RI2, (AHTC=7392 W/m2"C) ,o 50%R22150%R142b,(8875) 18OOO -D 60%R22/40%RI23, (5029) * 20%R22/80%R152a, (14198) 161)00 -m 70%R22/30%R141b, (6025) J j n 20%R32/80%R,42b, (8951) I / ~ M 14000 -A I0%R32/90%R 152a, (14947) Krco~t'-
(13414)
L) 32
~K~.m~ J/_mlm~"
/~llJ
" , ~
-
~
;
~
~
10000
~ ~ o ~°~600(}
41300
2000 0.0
0.2
0.4
0.6
Quality
0.8
1.0
Figure 5 Predicted local and average heat transfer coefficients of refrigerant mixtures over the quality of 20-90% for the mass flow rate of35.48gs ~andheatfluxof2OkWm "at0°C Figure 5 Coefficients de transfert de chaleur moyen et local de mdlanges de Jkigorig~nes pour une concentratiotl de 20 gt 90%, un d~bit masse de 35,48 g s : et un flux thermique de 20 kilowatts rn 2 ~ 0 des C
30%R141b are as much as 44% lower than that of RI2 because of the same capacity requirement. In order to separate the effect of the reduction in mass flow rate on heat transfer degradation from other effects, average heat transfer coefficients are calculated on the same mass flow rate basis. Figure 5 illustrates local and average heat transfer coefficients of various fluids for the mass flow rate of 35.48 g s-n at evaporation temperature of 0°C. Once again, nucleate boiling is suppressed under this condition. For the same mass flow rate, mixtures with large volatility difference (60%R22/40%R123 and 70%R22/ 30%R141b) show as much as 32% reduction in average heat transfer coefficients compared to Rt2, which is caused by the mass transfer resistance and physical
Evaporation heat transfer coefficient and pressure drop. D. Jung and R. Radermacher Table 4
Effect of the degradation of refrigerant side heat transfer coefficient on overall heat transfer coefficient (OHTC) for air and water systems Tableau 4 Effet de la diminution du coefficient de transfert de chaleur c6t~ frigorigene sur le coefficient de transJert de chaleur global, pour des syst~mes h air et 3 eau. Air system RI2
60%R22/40%R123
OHTC 98.7 % degradation
97.2 1.5%
Water system R12 60%R22/40%R123 1595
1275 20%
E
(Typical heat transfer coefficients for air and water sides are 100 and 2000 respectively and heat transfer coefficients for R12 and 60%R22/ 40 %R123 are 7392 and 3496 W m -~°C -~ obtained at 0°C. O H T C is in W m -~°C ')
property change. On the other hand, all other mixtures exhibit as much as 90% increase in average heat transfer coefficients, as shown in Figure 5. Thus, it may be concluded that HTCs of mixtures with large volatility difference are 55% lower than those of R12 for the same capacity. Out of the 55% reduction, mass transfer resistance and property degradation are responsible for 32%, and the remaining 23% is due to reduced mass flow rate to match the capacity. For mixtures with moderate volatility difference (50%R22/50%R142b, 20%R22/ 80%R152a, 20%R32/80%R142b, and 10%R32/ 90%R152a), however, the total reduction in heat transfer coefficients is 20-30% for the same capacity. Since volatility difference is small for these mixtures, the mass transfer resistance and property degradation are small. Therefore, most of the reduction in heat transfer for mixtures with moderate volatility difference stems from reduced mass flow rates to match the capacity.
Impact of the degradation of heat transfer coefficients with mixtures For mixtures, what is more important than the degradation of heat transfer coefficients alone is its impact on overall heat transfer coefficients. When either copper or aluminium tube is used, the thermal resistance for conduction can be neglected, and hence overall heat transfer coefficient can be expressed as U _
l
1
__
1
ho h~ + ho
r12
50%r22! (~)%R221 20%R2~ 70%R22/ 20%R32/ Io%r32f 50%R32/ 40~R32/ 50%R142b 40%R123 80%R152a 30~RI41bS0%RI42b90%RI52a 50~RI42b 60~R152a
R22
Fluids
Figure 6 Predicted pressure drop of refrigerant mixtures over the quality of 20-90% for the cooling capacity of 3 kW at various evaporation temperatures, q = 15 kW m ~, d = 8 m m Figure 6 Perte de charge estim~e pour des mklanges de frigorig~nes, avec une concentration de 20 h 90%, une puissance/kigor(fique de 3 kilowatts, ~ diff~rentes tempOratures d'Ovaporation, q = 15 kilowatts rn 2, d = 8 m m
While the maximum reduction in overall HTC for mixtures with large volatility difference is only 1.5% for an air system, it becomes as much as 20% for a water system. This clearly explains the conflict between the observation of Stoecker 5 (water system) and that of another report 6 (air system) mentioned earlier. It may be concluded that considerable reduction in HTC for mixtures with a large gliding temperature difference (or volatility difference) may or may not be of significance, depending upon the heat transfer fluid. If the heat transfer fluid side resistance is an order of magnitude larger than that of the refrigerant side (air system), its overall effect is small. However, when the heat transfer fluid side resistance is the same order of magnitude (water system), the reduction in heat transfer coefficient associated with mixtures would significantly affect the overall heat transfer coefficient.
Pressure drop of mixtures
hi
(3)
+ ho
where h~ and ho are the refrigerant and heat transfer fluid side heat transfer coefficients, respectively. There are two popular heat transfer fluids found in normal HVAC systems: air and water. When air is used as heat source or sink fluid, the typical heat transfer coefficient on the air side is less than 100 Wm-2K '. On the other hand, when water or its mixture with antifreeze is used as a heat transfer fluid, heat transfer coefficients fall in the range of 1000-2000 Wm-2K-L Using these typical values, a comparison will be made with a mixture of 60%R22/40%R123 with large volatility difference against RI2. For the capacity of 4 kW, the heat transfer coefficients for R12 and 60%R22/40%R123 mixture at evaporation temperature of 0°C are 7392.0 and 3496.0 W m-2K 2, respectively. Table 4 lists overall HTCs obtained with these values for two heat transfer fluids.
Figure 6 illustrates the predicted pressure drop of various refrigerants for the cooling capacity of 3 kW. As with the heat transfer coefficient, the pressure drop decreases as temperature increases. One point clearly seen in Figure 6 is that the fluids with larger HTCs also have larger pressure drops, which is similar to a single-phase turbulent flow situation. All mixture substitutes for R12 and R22 show smaller pressure drops than those of Ri2 and R22 due mainly to the reduced mass flow rates to maintain the same capacity. Effect of uncertainties in transport properties on heat transfer prediction Since any transport correlations require physical properties, their accuracy greatly depends upon the accuracies of physical properties. There are well-established thermodynamic properties of both pure and mixed refrigerants and their uncertainties are very small. There, howRev. Int. Froid 1993 Vo116 No 3
207
Evaporation heat transfer coefficient and pressure drop: D. Jung and R. Radermacher Table 5 Effect of uncertainties in transport properties on average heat transfer coelticient prediction Tableau 5 Eff('ct des incertitudes dans les propri~tes de transport sur la pr~vision du co£~cient moyen de tran.~/~,rt de chah'ur
R 12 HTC
60% R22/40% R [ 23 dO
HTC
5(t'% R22/50%R142b •
3496
1{T("
(13
Unchanged
7392
55~.6
10% increase in VL KL VV ST All
7069 7820 7448 7392 7534
- 4.37 5.8 0.76 0.0 1.92
3422 3342 3700 3496 3564
4.43 5.82 0,77 0,0 1,92
5~ 50 5(h~4 5~Q5 55~5 %90
~4,22 5.69 0,07 0,0 1.86
10% decrease in VL KL VV ST All
7767 6947 7330 7393 7238
5.1 - 6.0 - 0.83 0.0 2.07
3676 3284 3467 3497 3423
5.14 ~. 1 0.85 0.0 - 2.1
5*65 5250 5~3(1 5587 5474
50 6.(! 1.(~ !).0 • 2,0
(VL, VV, KL, ST, and • are viscosities of liquid and vapour, thermal conductivity of liquid, surface tension, and percent increase or decrease in HTC as compared to the HTC obtained without changing properties respectively. HTCs are predicted at 0°C for 4 kW capacity and are in W m : °C ~)
ever, is a lack of transport property data of pure fluids, not to mention mixtures. To make matters even worse, a large scatter is often found with mixture data, as revealed elsewhere '9. Therefore, it would be useful to carry out a sensitivity analysis for studying the effect of transport properties on heat transfer prediction. For this purpose, viscosities, thermal conductivities of liquid and vapour, and surface tension are varied from 90 to 110% of the measured data used in the above analysis. Table 5 lists heat transfer coefficients and percent changes in heat transfer coefficients qb of pure and mixed refrigerants as properties change. For both pure and mixed refrigerants, HTCs are most sensitive to the viscosity and thermal conductivity of liquid. With 10% uncertainty in both viscosity and thermal conductivity, 6% change in HTCs is observed for both pure and mixed refrigerants. Heat transfer coefficients are not sensitive to other properties.
Conclusions Jung et al.'s correlations for heat transfer coefficient and pressure drop are used to predict HTCs and pressure drops of prospective mixtures to replace RI2 and R22. To compare working fluids on a fair basis, comparisons are made based on the same cooling capacity basis. For this purpose, a basic vapour compression cycle was modelled to calculate required mass flow rates and evaporator inlet qualities for all fluids studied. The following conclusions can be drawn from this study: 1.
2.
208
In the evaporating temperature range of - 1 0 to 10°C, the evaporator inlet quality is usually greater than 20% for all fluids considered. In the quality range greater than 20%, nucleate boiling is fully suppressed for all pure and mixed refrigerants. The transition qualities for mixtures are lower than those of pure fluids due to the loss of wall superheat and mass transfer resistance. In this convective evaporation region, the heat transfer coefficient increases monotonically with quality. For the same mass flow rate, certain mixtures containing R32, R152a and R142b exhibit as much as Int. J. Refrig. 1993 Vo116 No 3
90% higher HTCs than RI2 and R22. For the same capacity, however, mixtures containing R32 and R152a show only 8 10% increase in HTCs because of reduced mass flow rates. For the same capacity, mixtures with large volatility difference show up to 55% reduction compared to Rl2 and R22, which is caused by mass transfer resistance and property degradation due to mixing (32%) and reduced mass flow rates (23%). On the other hand, mixtures with moderate volatility difference exhibit 20-30% reduction, which is mainly due to reduced mass flow rates to match the capacity. . The significance of the reduction in HTCs for mixtures with large volatility difference depends upon the heat transfer fluids. If the major resistance to heat flow exists in the heat transfer fluid side (air system), the reduction in overall HTC is less than t.5%. However, when the resistance in the heat transfer fluid side is the same order of magnitude (water system), as much as 20.0% reduction in overall HTC is expected. . For a given fluid, pressure drop becomes smaller as evaporation temperature decreases. For the same capacity, pressure drops of all mixtures considered are smaller than those of RI2 and R22 because of reduced mass flow rates. . Heat transfer prediction is more sensitive to the properties of liquid than those of vapour. With typical uncertainties of 10% in transport properties, HTCs changed less than 6%.
Acknowledgement The funding for this work was provided by the US Environmental Protection Agency and Whirlpool Corporation. The authors acknowledge D. Didion, M. McLinden and G. Morrison at the National Institute of Standards and Technology for their helpful comments and discussions throughout the study. References I
EPA report CFCs and Stratospheric Ozone United States Environmental Protection Agency (December 19873
Evaporation heat transfer coefficient and pressure drop. D. Jung and R. Radermacher 2
3
4 5 6 7
8 9 l0
Montreal Protocol on Substances That Deplete the 'Ozone Layer Final Act, United Nations Environment Programme (1987) Malroy, W., Kauffeld, M., McLinden, M., Didion, D. Experimental evaluation of two refrigerant mixtures in a breadboard air conditioner Proc lnt lnst Refrigeration, Purdue Conference on CFCs Commissions BI, B2, E1 & E2 (1988) 27-34 Kruse, H. The advantages of non-azeotropic refrigerant mixtures for heat pump application Int J Refrig (1981) 4 119-125 Stoecker, W.F. Improving the Energy Effectiveness of Domestic R£/r~gerators by the Application ~tf Refrigerant Mixtures ORNL report, ORNL/Sub-78/55463/1 (1978) ORNL report An Evaluation o/'a Two-Evaporator Re/'rigeratorFreezer Using Nonazeotropic Refrigerant Mixtures ORNL/Sub/ 82-47952/1 (1982) Jung, D.S., Radermacher, R. Prediction of heat transfer coefficient and pressure drop of various refrigerants (submitted to ASHRAE Trans May 1990) Jung,D.S., McLinden, M., Radermacher, R., Didion, D. Horizontal flow boiling heat transfer experiments with a mixture of R22/R114 lnt J Heat Mass Transfer (1989) 32 (1) 131 145 Jung, D.S., MeLinden, M., Radermaeher,R., Didion, D. A study of flow boiling heat transfer with refrigerant mixtures lnt J Heat Mass Tran.~/br (1989) 32 (9) 1751 1764 Jung, D.S. Mixture effects on horizontal convective boiling heat transfer PhD thesis University of Maryland (May 1988) (Electric Power Research hzstitute Report EPRI ER-6364 Project 8006-2)
1[
12 13
14
15 16
17
18
19
Unal, H.C. Prediction of nucleate boiling heat transfer coefficients for binary mixtures lnt J Heat and Mass Transjer (1987) 29 (4) 637-640 Stephan, K. Heat transfer in boiling mixtures Proc 7th lnt Heat Transfer Conf Munich (1982) Chen, J.C. Correlation for boiling heat transfer to saturated fluids in convective flow lnd Eng Chem Process Design Develop (1966) 5 (3) Stephan, K., Abdelsalam, M. Heat transfer correlations for natural convection boiling lnt J Heat and Mass Transfer (1980) 23 73 87 i'~ Dittus, F.W., Boelter, L.M.K. Public Engineering Vol 2 University of California at Berkeley (1930) 443 Jung, D.S. and Radermaeher, R. Prediction of pressure drop during horizontal annular flow boiling of pure and mixed refrigerants Int J Heat Mass TransJer (1989) 32 (12) 2435 2446 Eekels, S.J., Pate, M.B. An experimental comparison of evaporation and condensation heat transfer coefficients for HFC-134a and CFC-12 (submitted to lnt J ReJ~ig 1990) Morrison, G., McLinden, M. Application of a Hard Sphere Equation of State to Refrigerants and Re[rigerant Mixtures NBS Technical Note 1226, NBS, Gaithersburg, MD (1986) Jung, D.S., Radermacher, R. Transport properties and surface tension of pure and mixed refrigerants (submitted to A S H R A E Trans April 1990)
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