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Condensation heat transfer of binary refrigerant mixtures of R22 and R114 inside a horizontal tube with internal spiral grooves
Condensation heat transfer of binary refrigerant mixtures of R22 and Rl14 inside a horizontal tube with internal spiral grooves* Shigeru Koyama, Akio Miyarat, Hiroshi Takamatsu and Tetsu Fujii Institute of Advanced Material Study, Kyushu University, 6-1 Kasuga-shi, F u k u o k a 816, Japan t Faculty of Science and Engineering, Saga University, 1 Honjo-machi, Saga-shi, Saga 840, Japan Received 3 July 1989; revised 20 February 1990
An experimental study of the condensation of pure and mixed refrigerants of R22 and R 114 inside a spirally grooved horizontal copper tube has been carried out. A double-tube counterflow condenser in the pressure range 3-21 bar and at a mass flow-rate 2670 kg h- 1 was used. The axial distributions of refrigerant, tube wall and cooling water temperatures, wall heat flux density and vapour quality are shown graphically. The variation of tube wall temperature around the circumference of the tube is also shown. The local Nusselt number depends on the molar fraction, whereas the average Nusselt number can be correlated by an equation which is modified from a previously established equation for pure refrigerants inside a horizontal smooth tube. The frictional pressure drop evaluated is correlated well by the Lockhart-Martinelli parameters and is independent of the concentration of the mixture.
(Keywords:condensation;heat transfer; pressuredrop; horizontaltube; refrigerantmixtures)
Transfert de chaleur par condensation des mrlanges binaires des frigorigrnes R22 et R114, dans un tube horizontal/t rainures internes en spirale On a effectuk une ktude expkrimentale sur la condensation des frigorigbnes purs et en mklanyes de R22 et de Rl14, clans un tube horizontal en cuivre h rainures en spirale. On a utilisb un condenseur h contre-courant h double tube, ~ des pressions allant de 3 h 21 bars et hun dkbit masse de 26 b 70 kg h i On montre, sousJorme graphique, les distributions axiales du frigorigkne, les tempbratures de la paroi du tube et de l'eau de refroidissement, la densit~ de flux thermique de la paroi et la qualit~ de la vapeur. On montre kgalement la variation de la temperature de la paroi du tube le Ion9 de sa circonfkrence. Le nombre de Nusselt local dkpend de la fraction molaire, alors que le nombre de Nusselt moyen peut ktre corrklk par une kquation dbrivbe d'une kquation prkalablement ktablie pour des frigoriyknes purs clans un tube lisse horizontal. La chute de pression par frottement qui est bvaluke, est en bonne correlation avec les paramktres de Lockhart-Martinelli et est indbpendante de la concentration du mblange.
(Mots cl+s: condensation; transfert de chaleur; chute de pression; tube horizontal; mrlange de frigorigrne)
In recent years increasing attention has been paid to improving the thermodynamic performance of vapour compression heat p u m p systems by using non-azeotropic refrigerant mixtures as the working fluid. In the development of an improved system, the accurate prediction of the heat transfer characteristics in condensers and evaporators is required, in addition to the thermodynamic properties of refrigerant mixtures. There have been several studiesX-5 on the condensation of refrigerant mixtures inside a horizontal tube. Bokhanovskiy x carried out condensation inside a double-tube coiled heat exchanger with R22-R12 mixtures and indicated that in the region of low heat flux density the average heat transfer coefficients of R22-R 12 mixtures are lower than those of the pure refrigerants. Stoecker and K o r n o t a / experimentally investigated the heat transfer and flow characteristics of the condensation of R 12-R 114 mixtures inside a horizontal glass tube. No remarkable effect of composition on the flow pattern is observed in their photographs. Tandon et al. 3 proposed a flow
regimes m a p for the condensation of R22 RI2 mixtures inside a horizontal glass tube. In their m a p the flow pattern is not affected by composition. They also proposed 4 a correlation of the local heat transfer coefficient. Mochizuki et al. 5 presented a correlation of the average heat transfer coefficient of condensation of R1 I-R113 mixtures. These correlations are independent of composition; this is caused by a small diffusion resistance between the vapour and the vapour-liquid interface because of small differences between the compositions in vapour and condensate phases. As there are many possible combinations between the components and the mass fraction and heat transfer surface configuration, more work on the mechanism of heat transfer is required. This paper reports an investigation of the condensation of mixtures of R22 and R 114 inside a horizontal tube with internal spiral grooves. The results for heat transfer and pressure drop are compared with those reported 6-9 for pure refrigerants inside a horizontal smooth tube.
Experimental apparatus and measurement * Based on a paper presented at the Purdue/IIR Conference, Purdue University, West Lafayette, IN, USA, 18-21 July 1988 0140-7007/90/040256-08 © 1990 Butterworth-Heinemann Limited and IIR
256
Int. J. Refrig. 1990 Vol 13 July
Figure 1 is a schematic diagram of the experimental
Condensation heat transfer of binary refrigerant mixtures." S. Koyama et aL
Nomenclature
do
f,
G h 1 Nu P Ph Pr
g R
Re t T U W x
Y Y z
Isobaric specific heat capacity (J kg-a K-x) Mean inner diameter of inner tube (m) Outer diameter of inner tube (m) Function of molar fraction Mass velocity (kg m- 2 s- 1) Enthalpy (kJ kg- 1) Total condensing length (m) Nusselt number Pressure (bar) Phase change number Prandtl number Heat transfer rate (W) p-# ratio Two phase Reynolds number Temperature (K or °C) Velocity (m s- ~) Mass flow-rate (kg s- x or kg h- 1) Vapour quality Molar fraction Mass fraction Tube length
2 /~
Thermal conductivity (W m- 1 K - 1) Dynamic viscosity (Pa s) Void fraction Density (kgm -3) Lockhart-Martinelli parameter Lockhart-Martinelli parameter
p @v
Xtt
Subscripts b Bulk cal Calculated value exp Experimental value F Friction L Liquid M Momentum r Refrigerant s Heat sink water sat Saturation state v Vapour wi Inside of tube wall wo Outside of tube wall 1 Inlet of subsection 2 Outlet of subsection
Greek letters A
Heat transfer coefficient (W m- 2 K - 1) Change in physical quantity
Superscript Average
t"---'l I
:_ . . . .
r . . . . .
< . . . . .
r" - D . O - , ~ - -
i I~
400x12~4-800
1
"
I •,
" 4o0
II II ,~
Figure 1 Schematic diagram of experimental apparatus, (1) Compressor; (2) oil separator; (3) condenser; (4) sight glass; (5) subcooler; (6) turbine flow meter; (7) expansion valve; (8) evaporator; (9) accumulator; (10) reservoir; (11) inverter; (12) heat source unit; (13) heat sink unit; (14) float flow meter Figure 1 Schbma de l'appareil expbrimental. (1) Compresseur; (2) sbparateur d'huile; (3) condenseur; (4) voyant; (5) sous-refroidisseur; (6) dbbitmbtre it turbine; (7) dbtendeur; (8) bvaporateur; (9) accumulateur; (10) rbservoir; (11) convertisseur de frbquence de courant; (12) dispositif de source de chaleur; (13) dispositif de puits de chaleur; (14) dbbitmbtre it flotteur
Figure 2 Test condenser (length unit: mm). (1) Mixing chamber (refrigerant); (2) mixing chamber (heat sink water); (3) thermocouple (refrigerant); (4) pressure tap; (5) thermocouple (heat sink water); (6) thermocouple (outside surface of inner tube); (7) partition plate; (8) O-ring Figure 2 Condenseurexpdrimental (longueur en ram). (I) Chambre de mdlange (frifforiff~ne); (2) chambre de m~lange (eau comme puits de chaleur); (3) thermocouple (frigorig~ne); (4) robinet de pression; (5) thermocouple (eau comme puits de chaleur); (6) thermocouple (surface externe du tube int~rieur); (7) plaque de s~paration; (8) joint torique
apparatus. It is a vapour compression heat pump loop composed of a 0.75 kW hermetic compressor (1), an oil separator (2), a test condenser (3), a subcooler (5), an expansion valve (7) and an evaporator (8). The heat source and heat sink units (12 and 13) are used to supply heat source/sink water at constant temperature to the evaporator and condenser, respectively. The refrigerant flow-rate and pressure level in the loop are regulated by varying the rotating speed of the compressor and opening the expansion valve. The pressure level can also be modified by varying the flow-rate and/or temperature of
the water supplied to the condenser and evaporator. The concentration of refrigeration oil (SUNISO 3GS) is estimated to be about 0.1% at the oil separator outlet. Schematic views of the test condenser and its subdivided section are shown in Figure 2. The test condenser is a double-tube counterflow condenser, in which the refrigerant flows inside the inner tube and the heat sink water flows countercurrently in the outer annulus. The inner tube shown in Figure 3 is made of copper with dimensions as follows: outer diameter, 9.52 mm; mean wall thickness, 0.60 mm; mean inner diameter, 8.32 mm;
Rev. Int. Froid 1990 Vol 13 Juillet
257
Condensation heat transfer of binary refrigerant mixtures: S. Koyama et al. groove depth, 0.15 mm; number of grooves, 60; lead angle of grooves, 30 °. The outer tube is made of polycarbonate resin with an inner diameter of 16 mm. The total length of the condenser is 4.8 m. To measure the local heat transfer characteristics, the outer annulus is subdivided into 12 subsections 0.4 m in length and the effective heat transfer length in each subsection is estimated to be 0.37 m. The flow-rates of the refrigerant and heat sink water are measured with a turbine flow meter and a float flow meter, respectively, as shown in Figure 1. At the inlet and outlet of each subsection the refrigerant temperature and the heat sink water temperature are measured with sheathed chromel-alumel thermocouples (0.5mm o.d.) and the refrigerant pressure is measured with a pressure transducer and a differential pressure transducer. At the central positions of each subsection the circumferential temperature distribution of the outer surface of the inner tube is measured with four copper--constantan thermocouples, where four constantan wires (0.1 mm o.d.) are welded on the top, bottom, right and left sides, and a copper wire (0.1 mm o.d.) is welded 50mm away from the central position. The refrigerant mixture is sampled at the discharge port of the compressor and the bulk mass fraction is measured with a gas chromatograph. The range of test conditions is given in Table 1. The test fluids are R22, R114 and three mixtures containing about 25, 50 and 75% bulk molar fractions of Rl14.
expressed as:
W~[(1 --x,)h,., + x,hv, ] - ( 2 = Wr[(1 -x2)hL2+X2hv2 ]
(1)
where W, is the mass flow-rate of refrigerant, Q is the heat transfer rate calculated from the flow-rate and temperature rise of the heat sink water, h v and h E are the specific enthalpies of the vapour and liquid, respectively, and the subscripts I and 2 denote the inlet and outlet of the subsection, respectively. The vapour quality x is given by: x-
(2)
~L--L
where .vb is the bulk mass fraction of R114. The local mass fractions Yvand YLare determined from the vapour-liquid equilibria functionally expressed as: Yv = Yv(P, 'Tsar)
(3)
);L
(4)
=
YL(P, T~a,)
where P is the pressure. The specific enthalpy of the vapour hv is functionally expressed as: hv=Hv(P, T, Yv)
(5)
where the measured refrigerant temperature T~ is employed as T when T~> T~at,and T~a,is employed as T when T~< T~at. The specific enthalpy of liquid hL is also expressed functionally as: h E = HL(P , T, YL)
D a t a reduction
Fi#ure 4 shows the physical model employed in the data reduction. The axial distributions of the vapour quality x, the saturation temperature Ts~t and the vapour and liquid mass fractions of R114 Yv and YL are calculated assuming that the radial distributions of temperature and concentration in the vapour and liquid phases can be ignored. The energy balance in each subsection of length Az is
(6)
where T~.t is employed as T under the assumption that the effect of subcooling of the liquid is negligible. The axial distributions of x, T~.1, Yv, and YL can be obtained by solving the simultaneous equations composed of Equations (1)-(6) with measured values of Wr, P, Q and Yb under the condition that x at the refrigerant inlet is equal to 1. The phase equilibria and the enthalpies are cal-
•
AZ
,//// P1 ,Trl hvl, yvl Xl, Tsar1 J
hLI,~L
Vapor ~
P2,Tr2 ' hv2, .Yv2
~ ~ - - - ,
x2, Tsar2
Liquid
hL2,YL2
0 Figure 3 Schematic diagram of inner test tube (length unit: mm) Figure 3 Tube experimental (tube intbrieur, longueur en mm)
Figure 4 Physical model used in calculations Figure 4 ModOle physique
Table 1 Experimental conditions Tableau 1 Conditions exp~rimentales
Refrigerant
R22
Molar fraction of R114 Mass flow-rate (kg h- 1) Pressure (bar) Inlet saturation temperature (°C) Inlet superheat (°C) Heat transfer rate (kW)
0 36-70 15-21 40-54 20-41 2.0-3.7
258
Int. J. Refrig. 1990 Vol 13 July
R22-R 114 0.23-0.24 30-66 11-17 44-63 19-25 1,4-2.8
0.464).49 38-58 8-12 52~5 11-18 1.7-2.4
R114 0.73-0.75 40-52 6-7 51-58 11-23 1.6-2.0
1 26-51 3-5 35-59 2-26 1.0-1.9
Condensation heat transfer of binary refrigerant mixtures." S. Koyama et aL
culated by using the modified Benedict-Webb-Rubin equation of state ~°'~t, in which the value of a binary interaction parameter for the pair R22 and R114 is 0.976, referring to measured values of vapour-liquid equilibria of R22-R114 x2. The local heat transfer coefficient a is defined as: q
(7)
~a,-- Li where
(8)
Q
q = 7zdiAz T.
Twi =
' ^ ln(d°/di)
wo -t- ~ ~
(9)
where q is the heat flux density in the subsection Az(= 0.37 m), d~ and d o are the inner and outer diameters of the test tube, respectively, Twois the arithmetic mean of four outer wall temperatures measured in each subsection and 2,, is the thermal conductivity of the tube. The average heat transfer coefficient ~ is defined as:
where @=
{ (hv)z=o-(hL)==,} VVr udil
(11)
fraction does not affect the calculation of frictional pressure drop. In the dimensionless expressions of the local heat transfer coefficient and the frictional pressure drop the thermophysical properties of saturated vapour and liquid at each local position are used as reference properties. In the dimensionless expression of average heat transfer coefficient the thermophysical properties of superheated vapour at the inlet and those of saturated liquid at the end-point of condensation are used. The estimation of the thermophysical properties of mixtures are made as follows: (1) the dynamic viscosity of the liquid is calculated from the equation t4 expressed as In ]/L ~- Y~YLi(In #Li), and that of vapour is calculated from the Wilke's equationX4; (2) the thermal conductivity of liquid is calculated from the equation proposed by Chen et al.t 5 and that ofvapour is calculated from Wassilijewa's equation 14 and Lindsay-Bromley's equationt4; and (3) the isobaric specific heat capacities of liquid and vapour are calculated from the equation Cp= ZYicm.
Results and discussion
Local heat transfer Figure 5(a) and (b) shows the examples of the experimental results for pure refrigerant R22 and the refrigerant mixture 52 mol% R22-48 mol% R114, respectively. The section number on the abscissa is counted from the 100
(12)
~ . , _ (T~.t)z = o + (T,.,)= =, 2 "Fwi --
Z
'
'
/ 9O 80
TwiAz l
' ' o. R22 Wr = Pin = Ws=
I
'
AP F = A P - AP u
(14)
AP=PE-Pz
I
'
'
o : Ts o
6 X 104
: Twi
a :Tr
5¢,1
E
4~
0~= A
t 2r
4O
1.0
t~
30
0.8 0.6
20
0.4
v ~
1 2
3
4- 5
6
7
8
0.2 0
9 10 11 12
Section No.
(15)
rG,ex2 G,2(1-x)~-I l-G~x2 G,~(1-x)~] APM = I - - ~ + L ¢Pv O:~-~L
I,-LT : H
(1--~)PL
_]2
100
(16)
~={1 + ( p ~ ( 1 - - x ~ X
I
go 80
AP is the measured static pressure drop, APM is the pressure recovery due to the momentum change, Gr[ = 4 Wr/(ltd 2)] is the mass velocity of refrigerant, Pv and PL are the densities of vapour and liquid, respectively, and is the void fraction. As there is no available equation of for an internally grooved tube, it is estimated by the following equation x3 for two-phase flow in a smooth tube: \PL.]\"
eA
50
10
where
'
v : Ts~t o : 1-X • :q
7O
p 6o
'
70.0 kg/h 2 . 1 5 6 MPo 216.1 k g / h
(13)
I is the total condensing length, the superscript - denotes the mean value, and the subscripts z = 0 and z = l denote the refrigerant inlet and the position where the condensation has just completed, respectively. The frictional pressure drop of each subsetion AP F is calculated by the following equation:
'
a
. . . . I . . . . I ' b. 5 2 r n o l ~ R 2 2 + 4 8 m o l g R 1 1 4 Wr = 4 5 . 8 k g / h Pin = 1.173 MPo W= = 8 6 . 5 k g / h v : Tsot O:l-X
70 60
o:Tg o:Twi ~ : Tr
~
3
(17)
It is confirmed that the use of the other equations for void
~
t ~
J 1 2
x(O.4+O.6FX(OL/PO+O'g(1--X)]I/E~ -1 \ L x+O.4(1-x) J /J
5o4 4~
40:
10
6 X 10 4
E
~50
20
,/
'
~ 3
4
5
6
7
8
9 10 11 12
1,0 0.8
0.6 0.2 0
Secti on No. Figure 5 Examples of experimental results. (u) 1122; (b) 52mol%R22-48mol%R114 Figure 5 Exemples de r~sultats exp~rimentaux
Rev. Int. Froid 1 9 9 0 Vol 13 Juillet
259
Condensation heat transfer of binary refrigerant mixtures. S. Koyama et al. 55
'
'
i
'
I
'
'
'
R22
o.
'
I
'
o :Top
Wr=70.Okg/h
/, :R-side u :Bottom v :L-side
0
45
e~ 0e
~- 40 35
30
60
,
,
~
I
~
,
T
,
I
,
2 3 4 5 6 7 8 9101112 Section No. '
'
'
I
'
'
'
'
I
'
b. 52mol%R22+48mol%R114 55
5O
$
Wr=45.8kg/h
0 45
N u = ~di o :Top A :R-side u :Bottom v :L-side
0
e
4O
35
0
2 3 4~ 5~ 6I 7 ~ 8 ~ Section No.
~9=10111~1
2
Figure 6 Examples of axial distribution of tube wall temperature. (a) R22; (b) 52mo1% R22-48mol% R 114 Figure 6 Exemples de la distribution axiale de la temp@rature de la paroi du tube
refrigerant inlet. In Figure 5(a), for pure refrigerant, superheated vapour entering the test section is cooled down to the saturated vapour state with accompanying condensation, and the condensation proceeds at almost constant temperature until complete. The condensate is then subcooled. Accordingly, the temperature difference (T~,,-T~) in the condensing region increases in the direction of flow of refrigerant and reaches a maximum at x - 0 . In the condensation region of superheated vapour the heat flux density q decreases in the flow direction of the refrigerant, whereas in the condensation region of saturated vapour it remains almost constant. In Figure 5(b), for refrigerant mixtures, the saturation temperature T~= gradually decreases in the direction of flow of refrigerant, and the temperature difference (T,,~-Ts) is kept nearly constant in the condensing region by regulating the flow-rates of refrigerant and heat sink water. The heat flux density q gradually decreases in the direction of flow of refrigerant. Figure 6(a) and (b) shows the axial distributions of the outer wall surface temperatures for pure R22 and the mixture 52 mol% R22-48mol% R114, respectively. In Figure 6(a), which corresponds to the test run shown in Figure 5(a), the circumferential temperature variation is small near the refrigerant inlet and gradually increases in the direction of flow of refrigerant. This suggests that the flow pattern is annular near the inlet and changes into a
260
semi-annular flow as condensation proceeds. In the semi-annular flow region (Sections 6-9), the wall temperature decreases circumferentially in the order right side, top, bottom and left side. This fact allows the inference that the liquid film is deformed circumferentially by the grooves cut like a right hand screw. In Fi,qure 6(b), which corresponds to the test run shown in Fiyure 5(b), the circumferential temperature difference in the condensing region is small over the whole condensation region. Stoecker et al. 2 and Tandon et al? reported that the flow pattern is not affected by the composition of the refrigerant mixture. In the semi-annular flow region of refrigerant mixtures the temperature at the vapour liquid interface probably varies circumferentially. Figure 7(a) (e) shows a comparison between the local Nusselt number (NU)exp obtained experimentally and (Nu)ca I estimated from an empirical equation for pure refrigerants inside a horizontal smooth tube 6. The local Nusselt number is defined as:
Int. J. Refrig. 1990 Vol 13 July
(18)
where )-L is the thermal conductivity of liquid. In Fiyure 7(a) and (b) the (NU)exp values for R22 and Rl14 are higher than the (NU)cal values by 40-70% and 50-80%, respectively. These facts reveal the effectiveness of condensation enhancement due to the internal spiral grooves. The ratios (Nu)exp/(Nu)ca~ for the refrigerant mixtures are smaller than those for the pure refrigerants and the axial variations of the ratio are affected by both the bulk concentration and the flow-rate of refrigerant as shown in Fiqure 7(c)-(e).
Average heat transfer Figure 8 shows a comparison between the measured average Nusselt number of the pure refrigerants and the following empirical equation, which was obtained for the condensation of pure organic substances inside a horizontal smooth tube by Fujii and Nagata 7.
where
Nu = .
(20)
AL
Ph = Re z-
CpL(Tsat- "Fwi) (hvsat)z = 0 - (hLsa,)z= I
PL Uvl /~L
(21)
(22)
R --[ DLIIL] 1/2 - LP~_I
(23)
p r L = #LCpL
(24)
2L
Nu is the average Nusselt number, Ph is the phase change number, Re~ is the two-phase Reynolds number, R is the p-/~ ratio, Pr L is the Prandtl number of liquid, Uv is the mean vapour velocity at the refrigerant inlet, and the subscript sat denotes the saturation state. The measured values are about 100% higher than the values predicted by Equation (19). This shows that the internal spiral
Condensation heat transfeY of binary refrigerant mixtures." S. Koyama et aL 10 3
' o.
I/'./'
'
xl 0 2
' V/ -!
R22
/-
71
,
i
I
'
r d . 50molXR22+50rnol%R114
51-
#yo
.
-
~5
#~
/
o :,o.,
= "
,
I
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4 5 (Nu)cal
'
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x"~
2 ~°z~
z "J4
/- ~ ' / ~
-
.
A
: 47.7 '~ : 50.7 ¢. : 51.3
/ /
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I
sl/ ~" 4t-
: 28.1
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,, :31.3 0 : 32.4 • : 37.6 ,t : 39.3
I ~"
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-
./
×Zo v
I/
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e. 25molgR'22+75molgR114
An O . /
r kg/h o:39,9
p # /
40.2
o : 46.8
I
103
(Nu)cel xl 0 2 7
I
4
o:30.2
/
/
I
3
xlO 2
7
o:26.1-
/
~z'~"
/
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//
/
~/~1--
o 57.5
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__;~¢'~ ./ Wr kg/h 4~[~
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:Yl
•
i/ -
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~,/o/. /
-
..-r.~ 5
o
:~ .
I #Z I p ~ "~ 10 3
'
b. R 1 1 4 10 3
:68.1
•
A : 70.0 m : 70.3 ,,,,
•
/
I/~:~
2I~
I
,
I ,
2
3
4
5
7xi02
(Nu)cal i
,
I
i
-- c. 75molgR22+25molgR114
5 ~4
-
-
--
/~_/<~.~'~t/
Wr kg/h
o%'~
o: ~o.o
~s
:36.5
o ;:?5 L-/
2
• :58.1 ,t : 66.2 I
3
I
]
4
5
Figure 7 Comparison between (Nu)cxp and (Nu)c,r (a) R22; (b) Rl14; (c) 7 5 m o l % R 2 2 - 2 5 m o l % R l 1 4 ; (d) 5 0 m o l % R 2 2 - 5 0 m o l % R l 1 4 ; and (e) 25mol oVoR22-75mol oVoR114. Figure 7 Comparaison entre (Nu),,p et (Nu)c, l
I
ix10 2
(Nu)co,
grooves are effective in enhancing condensation. The experimental results for the pure refrigerants can be correlated well by the following equation, the functional form of which is the same as Equation (19): N
.
,.,I/ l'~-°4n,_o 6['Re, PrL'~°-a
"--°"W
"" t - - F - )
(25)
Figure 9 compares the experimental results for the refrigerant mixtures with Equation (25). The experimental data for the mixed refrigerants are lower than the values calculated with Equation (25). Fioure 10 shows the relationship between the ratio (Nu),,~/(Nu)ca~ and bulk molar fraction of Rll4, yb, where (Nu)=,p is the measured value and (Nu)¢al is the value predicted by Equation (25). These data can be
approximated by the following equation: (Nu)exp fy = 1 - 0.73y b + 0.37y 2 + 0.36y 3
(Nu)o,~
(26)
All of the experimental data for the pure and mixed refrigerants can be correlated well by the combined equation of Equations (25) and (26), as shown in Figure
11. Pressure drop Figures 12(a)-(e) shows the experimental results for the frictional pressure drop plotted in the coordinate of Lockhart-Martinelli parameters @v and X., which are defined as:
=[APEIAz)~ u2 ,, L(APIAz),, j
Rev. Int. Froid 1990 Vol 13 Juillet
(27)
261
Condensation heat transfer of binary refrigerant mixtures: S. Koyama et aL 5
o
'
I
'
'
'
' 'I
4
0 R22 A Rl14
o Z2 ~co c5 A
~
I
'
E~l'"q
1
R22
10 -
co
' 'I'"'I
(a)
l
'
~ /
"
,R,,',.,/' / >
2
#
J
£L ~'~ I0 6
• present
-
1
~0~
A Hash, z u m e
0.01
0.1
Z
1.0
5
Xtt I
53
,
,
,/~1
4- 5
10 8
2
3
Re~ ' '
Figure 8 Average Nusselt n u m b e r (pure refrigerants)
I .... I
' '-'1 .... t
-L
*
5
o A
[
4
~
=
i
~
.~'~
1
= I ll=14
_
U present
,
,~ Hashizume ,I,,,,I L t
,
0.1
0.01
:2f
CL ~'~-106
.
~50~ =
2
o
(b)
I
E:] 75mol%R22+25mol%R114 0 50mol%R22+50mol%R114 ~7 25mol%R22+75mol%R114
c5 co
i
' '
10 __- R114
Nombre de Nusselt moyen (frigorigbnes puts)
Figure 8
1.0
5
Xtt
~k~/~-/djo O v
Iz
(c) I , ~ ~1
5-
5
75mol%R22+25mol%R114
10
108
Figure 9
Average Nusselt n u m b e r (refrigerants mixtures)
Figure 9
Nombre de Nusseh moyen (mblanges de frigorigbnes)
> ....~ -~
1.2
. . . .
I
£~ Hashi zurne I
v
z~
-
i
;
0.1
0.01
o 5o-5o (~u)o~p/(~u)~o,
• present
o
I /IIIHi
I
R22+R114 (mot%) 0 100- 0 {3 7 5 - 25
1.1 "~1.0
. . . .
~0~
lIH]
~
I
1.0
5
Xtt
2 5 - 75 0-100
: 1 _ 0 . 7 3 yb +0.37 yt? + 0 , 3 6 yb-~
-
(d)
~o.9 %-
10
50mol%R22+50mol%R114/p
iz > 0.7
,
,
,
R22
, (~ . . . . O.5
Rt 14
)
Molar Fraction
Figure 1O R e l a t i o n s h i p between (Nu)~p/(Nu)~,~ a n d y~ Figure 10
-
• present A Hashizume
-~-~ 50% I
I I]1111l
I
I ,I,,,,I J
0.1
0.0
Relation entre (Nu)~v/(Nu)~.l et Yb
,
1.0
5
Xtt 5
9--.
'
4-
I ' ' ''1
R22+R114(molX)
' "i
....t
'
" I ....I
®
/z9" Z~
co
[2-
~
10 6
0--100
£3~,
~
-
12 53
I
4- 5
0.01 ,
,r,l
108 Re~
½
• present
Figure
Average Nusselt n u m b e r pure a n d m i x e d refrigerants
F i g u r e 11
Nombre de Nusselt moyen des frigorigbnes purs et des
mblanges
1990
Vol
13
July
0.1
_
z~ H o s h i z u m e
-~
1.0
5
Xtt
.3
Figure 11
Int. J. Refrig.
.../~ ~
1 --~30%
7.
262
(e)
' "
10 : 25mol%R22+75mol%R114~....
5
12 F r i c t i o n a l pressure drop. (a) R22; (b) R l l 4 ; (e) 75mo1% R 2 2 - 2 5 m o l % R114; (d) 50mo1% R 2 2 - 5 0 m o l % R114; a n d (e) 2 5 m o 1 % R 2 2 - 7 5 m o l % R114 F i g u r e 12 Chute de pression par frottement
Condensation heat transfer of binary refrigerant mixtures: S. Koyama et al. 1
are correlated fairly well by the Lockhart-Martinelli parameters independently of molar fraction.
X 0.9 Pv 0.5 flL 0.1
The frictional pressure gradient ( A p v / A z ) is calculated by Equation (14) and the pressure gradient for vapour How (Ap/Az)v is calculated by the following equation: ~-z v -
diPv
(29)
where fv is the friction coefficient estimated by the Colburn's equation as follows: fv = 0.046/{ Grxdi/#v} 0.2
(30)
Figure 12(a)-(e) corresponds to the results of R22, R 114, 75 mol% R22-25 mol% R 114, 50mo1% R22-50mol% R 114 and 25mo1% R22-75mol% Rl14, respectively. (O) denotes the present results and (A) denotes the experimental results by Hashizume s for the adiabatic two-phase flow of R22 in a horizontal smooth tube, rearranged by the authors, and the solid lines denote the Lockhart-Martinelli relation 9. The present results are correlated fairly well by the Lockhart-Martinelli parameters independently of bulk concentration, and are about 40% higher than Hashizume's results s due to the effect of grooves.
Acknowledgements
This study was supported by the Grant-in-Aid of Scientific Research of the Ministry of Education, Science and Culture, Japan, and was also sponsored by the Japanese Association of Refrigeration. References
1 2 3 4
5 6 7
Bokhanovskiy,YU. G. Heat transfer for Freon-12, Freon-22and their mixtures in a coiled tube condenser Heat Transf Soy Res (1980) 12 43-45 Stoecker,W. F., Kornota,E. Condensingcoefficientswhen using refrigerant mixture ASHRAE Trans (1985) 91 1351-1367 Tandon,T. N., Varnma, H. IL, Gnpta, C. P. Prediction of flow patterns during condensation of binary mixtures in a horizontal tube Trans ASME J Heat Transf (1985) 107 424-430 Tandon, T. N., Varuma, H. K., Gnpta, C. P. Generalized correlation for condensationof binary mixtures inside a horizontal tube Int J Refrio (1986) 9 134-136 Mochizuki, S., Inone, T., Tomiuga, M. Condensation of nonazeotropicbinarymixtures in a horizontal tube TransJSME Part B (1988) 54(503) 1796-1801 Fujii,T., Honda, I-L, Nozu, S. Condensation of fluorocarbon refrigerants inside a horizontal tube Refrigeration JAR (1980) 55(627) 3-20 Fujli,T., Nagata, T. Condensation of vapor in a horizontal tube Report of Research Institute of Industrial Science K yushu University (1973) 52 35-50
Conclusions
An experimental study of condensatin of refrigerant mixtures of R22 and R114 inside a horizontal tube with internal spiral grooves for a double-tube counterflow condenser was performed. The conclusions are as follows. 1. The axial distributions of saturation temperature and tube wall temperature for the refrigerant mixtures are different from those for the pure refrigerants. 2. In the case of the pure refrigerants, the local Nusselt number for a grooved tube is about 60% higher than that for a smooth tube. The distribution of local Nusselt number for the refrigerant mixtures is different from that for the pure refrigerants and depends on the bulk composition and flow-rate of the refrigerant. 3. The average heat transfer coefficient for the mixed refrigerants is lower than that for pure refrigerants, and the degree of maximum decrease is about 20%. The average heat transfer coefficient for the pure and mixed refrigerants can be correlated well by Equations (25) and (26). 4. The present results for the frictional pressure drop
8 9 10 11
12 13 14 15
Hashizmne,K. Flow pattern, void fraction and pressure drop of refrigerant two-phaseflowin a horizontal pipe-IInt J Multiphase Flow (1983) 9 399-410 Lockhart,R. W., MartinelU, R. C. Proposed correlation of data for isothermal two-phase, two-component flow in pipes Chem En9 Pro9 (1949) 45 39 Nishinmi,H., Saito, S. An improved generalized BWR equation of state applicable to low reduced temperatures J Chem Eng Jpn (1975) 8 356-360 Fukuzato,R., Tomisaka, Y., Arai, K., Saito, S. The prediction of binary vapor-liquid equilibrium for fluorocarbon mixtures by use of a modifiedBWR equation of state J Chem Eno Jpn (1983) 16 147-149 Hasegawa,N., Uematsu, M., Watanabe, K. Measurements of PVTx for the R22+RI14 system J Chem En0 Data (1985) 30 32-36 Smith,S. L. Void fraction in two-phaseflow:a correlation based upon an equal velocity head model. Heat Fluid Flow (1971) 1 22-39 Reid,R. C., Prausnitz, J. M., Sherwood,T. K. The Properties of Gases and Liquids McGraw-Hill New York (1977) Chen,Z., Fujii, T., Fujii, M. A methodfor evaluating the mixture constant in the expression for thermal conductivity of binary liquid mixture and its extension to multi-component liquid mixtures Report of Research Institute of Industrial Science Kyushu University (1987) 82 173-187
Rev. Int. Froid 1990 Vol 13 Juillet
263
An experimental study of the condensation of pure and mixed refrigerants of R22 and R 114 inside a spirally grooved horizontal copper tube has been carried out. A double-tube counterflow condenser in the pressure range 3-21 bar and at a mass flow-rate 2670 kg h- 1 was used. The axial distributions of refrigerant, tube wall and cooling water temperatures, wall heat flux density and vapour quality are shown graphically. The variation of tube wall temperature around the circumference of the tube is also shown. The local Nusselt number depends on the molar fraction, whereas the average Nusselt number can be correlated by an equation which is modified from a previously established equation for pure refrigerants inside a horizontal smooth tube. The frictional pressure drop evaluated is correlated well by the Lockhart-Martinelli parameters and is independent of the concentration of the mixture.
(Keywords:condensation;heat transfer; pressuredrop; horizontaltube; refrigerantmixtures)
Transfert de chaleur par condensation des mrlanges binaires des frigorigrnes R22 et R114, dans un tube horizontal/t rainures internes en spirale On a effectuk une ktude expkrimentale sur la condensation des frigorigbnes purs et en mklanyes de R22 et de Rl14, clans un tube horizontal en cuivre h rainures en spirale. On a utilisb un condenseur h contre-courant h double tube, ~ des pressions allant de 3 h 21 bars et hun dkbit masse de 26 b 70 kg h i On montre, sousJorme graphique, les distributions axiales du frigorigkne, les tempbratures de la paroi du tube et de l'eau de refroidissement, la densit~ de flux thermique de la paroi et la qualit~ de la vapeur. On montre kgalement la variation de la temperature de la paroi du tube le Ion9 de sa circonfkrence. Le nombre de Nusselt local dkpend de la fraction molaire, alors que le nombre de Nusselt moyen peut ktre corrklk par une kquation dbrivbe d'une kquation prkalablement ktablie pour des frigoriyknes purs clans un tube lisse horizontal. La chute de pression par frottement qui est bvaluke, est en bonne correlation avec les paramktres de Lockhart-Martinelli et est indbpendante de la concentration du mblange.
(Mots cl+s: condensation; transfert de chaleur; chute de pression; tube horizontal; mrlange de frigorigrne)
In recent years increasing attention has been paid to improving the thermodynamic performance of vapour compression heat p u m p systems by using non-azeotropic refrigerant mixtures as the working fluid. In the development of an improved system, the accurate prediction of the heat transfer characteristics in condensers and evaporators is required, in addition to the thermodynamic properties of refrigerant mixtures. There have been several studiesX-5 on the condensation of refrigerant mixtures inside a horizontal tube. Bokhanovskiy x carried out condensation inside a double-tube coiled heat exchanger with R22-R12 mixtures and indicated that in the region of low heat flux density the average heat transfer coefficients of R22-R 12 mixtures are lower than those of the pure refrigerants. Stoecker and K o r n o t a / experimentally investigated the heat transfer and flow characteristics of the condensation of R 12-R 114 mixtures inside a horizontal glass tube. No remarkable effect of composition on the flow pattern is observed in their photographs. Tandon et al. 3 proposed a flow
regimes m a p for the condensation of R22 RI2 mixtures inside a horizontal glass tube. In their m a p the flow pattern is not affected by composition. They also proposed 4 a correlation of the local heat transfer coefficient. Mochizuki et al. 5 presented a correlation of the average heat transfer coefficient of condensation of R1 I-R113 mixtures. These correlations are independent of composition; this is caused by a small diffusion resistance between the vapour and the vapour-liquid interface because of small differences between the compositions in vapour and condensate phases. As there are many possible combinations between the components and the mass fraction and heat transfer surface configuration, more work on the mechanism of heat transfer is required. This paper reports an investigation of the condensation of mixtures of R22 and R 114 inside a horizontal tube with internal spiral grooves. The results for heat transfer and pressure drop are compared with those reported 6-9 for pure refrigerants inside a horizontal smooth tube.
Experimental apparatus and measurement * Based on a paper presented at the Purdue/IIR Conference, Purdue University, West Lafayette, IN, USA, 18-21 July 1988 0140-7007/90/040256-08 © 1990 Butterworth-Heinemann Limited and IIR
256
Int. J. Refrig. 1990 Vol 13 July
Figure 1 is a schematic diagram of the experimental
Condensation heat transfer of binary refrigerant mixtures." S. Koyama et aL
Nomenclature
do
f,
G h 1 Nu P Ph Pr
g R
Re t T U W x
Y Y z
Isobaric specific heat capacity (J kg-a K-x) Mean inner diameter of inner tube (m) Outer diameter of inner tube (m) Function of molar fraction Mass velocity (kg m- 2 s- 1) Enthalpy (kJ kg- 1) Total condensing length (m) Nusselt number Pressure (bar) Phase change number Prandtl number Heat transfer rate (W) p-# ratio Two phase Reynolds number Temperature (K or °C) Velocity (m s- ~) Mass flow-rate (kg s- x or kg h- 1) Vapour quality Molar fraction Mass fraction Tube length
2 /~
Thermal conductivity (W m- 1 K - 1) Dynamic viscosity (Pa s) Void fraction Density (kgm -3) Lockhart-Martinelli parameter Lockhart-Martinelli parameter
p @v
Xtt
Subscripts b Bulk cal Calculated value exp Experimental value F Friction L Liquid M Momentum r Refrigerant s Heat sink water sat Saturation state v Vapour wi Inside of tube wall wo Outside of tube wall 1 Inlet of subsection 2 Outlet of subsection
Greek letters A
Heat transfer coefficient (W m- 2 K - 1) Change in physical quantity
Superscript Average
t"---'l I
:_ . . . .
r . . . . .
< . . . . .
r" - D . O - , ~ - -
i I~
400x12~4-800
1
"
I •,
" 4o0
II II ,~
Figure 1 Schematic diagram of experimental apparatus, (1) Compressor; (2) oil separator; (3) condenser; (4) sight glass; (5) subcooler; (6) turbine flow meter; (7) expansion valve; (8) evaporator; (9) accumulator; (10) reservoir; (11) inverter; (12) heat source unit; (13) heat sink unit; (14) float flow meter Figure 1 Schbma de l'appareil expbrimental. (1) Compresseur; (2) sbparateur d'huile; (3) condenseur; (4) voyant; (5) sous-refroidisseur; (6) dbbitmbtre it turbine; (7) dbtendeur; (8) bvaporateur; (9) accumulateur; (10) rbservoir; (11) convertisseur de frbquence de courant; (12) dispositif de source de chaleur; (13) dispositif de puits de chaleur; (14) dbbitmbtre it flotteur
Figure 2 Test condenser (length unit: mm). (1) Mixing chamber (refrigerant); (2) mixing chamber (heat sink water); (3) thermocouple (refrigerant); (4) pressure tap; (5) thermocouple (heat sink water); (6) thermocouple (outside surface of inner tube); (7) partition plate; (8) O-ring Figure 2 Condenseurexpdrimental (longueur en ram). (I) Chambre de mdlange (frifforiff~ne); (2) chambre de m~lange (eau comme puits de chaleur); (3) thermocouple (frigorig~ne); (4) robinet de pression; (5) thermocouple (eau comme puits de chaleur); (6) thermocouple (surface externe du tube int~rieur); (7) plaque de s~paration; (8) joint torique
apparatus. It is a vapour compression heat pump loop composed of a 0.75 kW hermetic compressor (1), an oil separator (2), a test condenser (3), a subcooler (5), an expansion valve (7) and an evaporator (8). The heat source and heat sink units (12 and 13) are used to supply heat source/sink water at constant temperature to the evaporator and condenser, respectively. The refrigerant flow-rate and pressure level in the loop are regulated by varying the rotating speed of the compressor and opening the expansion valve. The pressure level can also be modified by varying the flow-rate and/or temperature of
the water supplied to the condenser and evaporator. The concentration of refrigeration oil (SUNISO 3GS) is estimated to be about 0.1% at the oil separator outlet. Schematic views of the test condenser and its subdivided section are shown in Figure 2. The test condenser is a double-tube counterflow condenser, in which the refrigerant flows inside the inner tube and the heat sink water flows countercurrently in the outer annulus. The inner tube shown in Figure 3 is made of copper with dimensions as follows: outer diameter, 9.52 mm; mean wall thickness, 0.60 mm; mean inner diameter, 8.32 mm;
Rev. Int. Froid 1990 Vol 13 Juillet
257
Condensation heat transfer of binary refrigerant mixtures: S. Koyama et al. groove depth, 0.15 mm; number of grooves, 60; lead angle of grooves, 30 °. The outer tube is made of polycarbonate resin with an inner diameter of 16 mm. The total length of the condenser is 4.8 m. To measure the local heat transfer characteristics, the outer annulus is subdivided into 12 subsections 0.4 m in length and the effective heat transfer length in each subsection is estimated to be 0.37 m. The flow-rates of the refrigerant and heat sink water are measured with a turbine flow meter and a float flow meter, respectively, as shown in Figure 1. At the inlet and outlet of each subsection the refrigerant temperature and the heat sink water temperature are measured with sheathed chromel-alumel thermocouples (0.5mm o.d.) and the refrigerant pressure is measured with a pressure transducer and a differential pressure transducer. At the central positions of each subsection the circumferential temperature distribution of the outer surface of the inner tube is measured with four copper--constantan thermocouples, where four constantan wires (0.1 mm o.d.) are welded on the top, bottom, right and left sides, and a copper wire (0.1 mm o.d.) is welded 50mm away from the central position. The refrigerant mixture is sampled at the discharge port of the compressor and the bulk mass fraction is measured with a gas chromatograph. The range of test conditions is given in Table 1. The test fluids are R22, R114 and three mixtures containing about 25, 50 and 75% bulk molar fractions of Rl14.
expressed as:
W~[(1 --x,)h,., + x,hv, ] - ( 2 = Wr[(1 -x2)hL2+X2hv2 ]
(1)
where W, is the mass flow-rate of refrigerant, Q is the heat transfer rate calculated from the flow-rate and temperature rise of the heat sink water, h v and h E are the specific enthalpies of the vapour and liquid, respectively, and the subscripts I and 2 denote the inlet and outlet of the subsection, respectively. The vapour quality x is given by: x-
(2)
~L--L
where .vb is the bulk mass fraction of R114. The local mass fractions Yvand YLare determined from the vapour-liquid equilibria functionally expressed as: Yv = Yv(P, 'Tsar)
(3)
);L
(4)
=
YL(P, T~a,)
where P is the pressure. The specific enthalpy of the vapour hv is functionally expressed as: hv=Hv(P, T, Yv)
(5)
where the measured refrigerant temperature T~ is employed as T when T~> T~at,and T~a,is employed as T when T~< T~at. The specific enthalpy of liquid hL is also expressed functionally as: h E = HL(P , T, YL)
D a t a reduction
Fi#ure 4 shows the physical model employed in the data reduction. The axial distributions of the vapour quality x, the saturation temperature Ts~t and the vapour and liquid mass fractions of R114 Yv and YL are calculated assuming that the radial distributions of temperature and concentration in the vapour and liquid phases can be ignored. The energy balance in each subsection of length Az is
(6)
where T~.t is employed as T under the assumption that the effect of subcooling of the liquid is negligible. The axial distributions of x, T~.1, Yv, and YL can be obtained by solving the simultaneous equations composed of Equations (1)-(6) with measured values of Wr, P, Q and Yb under the condition that x at the refrigerant inlet is equal to 1. The phase equilibria and the enthalpies are cal-
•
AZ
,//// P1 ,Trl hvl, yvl Xl, Tsar1 J
hLI,~L
Vapor ~
P2,Tr2 ' hv2, .Yv2
~ ~ - - - ,
x2, Tsar2
Liquid
hL2,YL2
0 Figure 3 Schematic diagram of inner test tube (length unit: mm) Figure 3 Tube experimental (tube intbrieur, longueur en mm)
Figure 4 Physical model used in calculations Figure 4 ModOle physique
Table 1 Experimental conditions Tableau 1 Conditions exp~rimentales
Refrigerant
R22
Molar fraction of R114 Mass flow-rate (kg h- 1) Pressure (bar) Inlet saturation temperature (°C) Inlet superheat (°C) Heat transfer rate (kW)
0 36-70 15-21 40-54 20-41 2.0-3.7
258
Int. J. Refrig. 1990 Vol 13 July
R22-R 114 0.23-0.24 30-66 11-17 44-63 19-25 1,4-2.8
0.464).49 38-58 8-12 52~5 11-18 1.7-2.4
R114 0.73-0.75 40-52 6-7 51-58 11-23 1.6-2.0
1 26-51 3-5 35-59 2-26 1.0-1.9
Condensation heat transfer of binary refrigerant mixtures." S. Koyama et aL
culated by using the modified Benedict-Webb-Rubin equation of state ~°'~t, in which the value of a binary interaction parameter for the pair R22 and R114 is 0.976, referring to measured values of vapour-liquid equilibria of R22-R114 x2. The local heat transfer coefficient a is defined as: q
(7)
~a,-- Li where
(8)
Q
q = 7zdiAz T.
Twi =
' ^ ln(d°/di)
wo -t- ~ ~
(9)
where q is the heat flux density in the subsection Az(= 0.37 m), d~ and d o are the inner and outer diameters of the test tube, respectively, Twois the arithmetic mean of four outer wall temperatures measured in each subsection and 2,, is the thermal conductivity of the tube. The average heat transfer coefficient ~ is defined as:
where @=
{ (hv)z=o-(hL)==,} VVr udil
(11)
fraction does not affect the calculation of frictional pressure drop. In the dimensionless expressions of the local heat transfer coefficient and the frictional pressure drop the thermophysical properties of saturated vapour and liquid at each local position are used as reference properties. In the dimensionless expression of average heat transfer coefficient the thermophysical properties of superheated vapour at the inlet and those of saturated liquid at the end-point of condensation are used. The estimation of the thermophysical properties of mixtures are made as follows: (1) the dynamic viscosity of the liquid is calculated from the equation t4 expressed as In ]/L ~- Y~YLi(In #Li), and that of vapour is calculated from the Wilke's equationX4; (2) the thermal conductivity of liquid is calculated from the equation proposed by Chen et al.t 5 and that ofvapour is calculated from Wassilijewa's equation 14 and Lindsay-Bromley's equationt4; and (3) the isobaric specific heat capacities of liquid and vapour are calculated from the equation Cp= ZYicm.
Results and discussion
Local heat transfer Figure 5(a) and (b) shows the examples of the experimental results for pure refrigerant R22 and the refrigerant mixture 52 mol% R22-48 mol% R114, respectively. The section number on the abscissa is counted from the 100
(12)
~ . , _ (T~.t)z = o + (T,.,)= =, 2 "Fwi --
Z
'
'
/ 9O 80
TwiAz l
' ' o. R22 Wr = Pin = Ws=
I
'
AP F = A P - AP u
(14)
AP=PE-Pz
I
'
'
o : Ts o
6 X 104
: Twi
a :Tr
5¢,1
E
4~
0~= A
t 2r
4O
1.0
t~
30
0.8 0.6
20
0.4
v ~
1 2
3
4- 5
6
7
8
0.2 0
9 10 11 12
Section No.
(15)
rG,ex2 G,2(1-x)~-I l-G~x2 G,~(1-x)~] APM = I - - ~ + L ¢Pv O:~-~L
I,-LT : H
(1--~)PL
_]2
100
(16)
~={1 + ( p ~ ( 1 - - x ~ X
I
go 80
AP is the measured static pressure drop, APM is the pressure recovery due to the momentum change, Gr[ = 4 Wr/(ltd 2)] is the mass velocity of refrigerant, Pv and PL are the densities of vapour and liquid, respectively, and is the void fraction. As there is no available equation of for an internally grooved tube, it is estimated by the following equation x3 for two-phase flow in a smooth tube: \PL.]\"
eA
50
10
where
'
v : Ts~t o : 1-X • :q
7O
p 6o
'
70.0 kg/h 2 . 1 5 6 MPo 216.1 k g / h
(13)
I is the total condensing length, the superscript - denotes the mean value, and the subscripts z = 0 and z = l denote the refrigerant inlet and the position where the condensation has just completed, respectively. The frictional pressure drop of each subsetion AP F is calculated by the following equation:
'
a
. . . . I . . . . I ' b. 5 2 r n o l ~ R 2 2 + 4 8 m o l g R 1 1 4 Wr = 4 5 . 8 k g / h Pin = 1.173 MPo W= = 8 6 . 5 k g / h v : Tsot O:l-X
70 60
o:Tg o:Twi ~ : Tr
~
3
(17)
It is confirmed that the use of the other equations for void
~
t ~
J 1 2
x(O.4+O.6FX(OL/PO+O'g(1--X)]I/E~ -1 \ L x+O.4(1-x) J /J
5o4 4~
40:
10
6 X 10 4
E
~50
20
,/
'
~ 3
4
5
6
7
8
9 10 11 12
1,0 0.8
0.6 0.2 0
Secti on No. Figure 5 Examples of experimental results. (u) 1122; (b) 52mol%R22-48mol%R114 Figure 5 Exemples de r~sultats exp~rimentaux
Rev. Int. Froid 1 9 9 0 Vol 13 Juillet
259
Condensation heat transfer of binary refrigerant mixtures. S. Koyama et al. 55
'
'
i
'
I
'
'
'
R22
o.
'
I
'
o :Top
Wr=70.Okg/h
/, :R-side u :Bottom v :L-side
0
45
e~ 0e
~- 40 35
30
60
,
,
~
I
~
,
T
,
I
,
2 3 4 5 6 7 8 9101112 Section No. '
'
'
I
'
'
'
'
I
'
b. 52mol%R22+48mol%R114 55
5O
$
Wr=45.8kg/h
0 45
N u = ~di o :Top A :R-side u :Bottom v :L-side
0
e
4O
35
0
2 3 4~ 5~ 6I 7 ~ 8 ~ Section No.
~9=10111~1
2
Figure 6 Examples of axial distribution of tube wall temperature. (a) R22; (b) 52mo1% R22-48mol% R 114 Figure 6 Exemples de la distribution axiale de la temp@rature de la paroi du tube
refrigerant inlet. In Figure 5(a), for pure refrigerant, superheated vapour entering the test section is cooled down to the saturated vapour state with accompanying condensation, and the condensation proceeds at almost constant temperature until complete. The condensate is then subcooled. Accordingly, the temperature difference (T~,,-T~) in the condensing region increases in the direction of flow of refrigerant and reaches a maximum at x - 0 . In the condensation region of superheated vapour the heat flux density q decreases in the flow direction of the refrigerant, whereas in the condensation region of saturated vapour it remains almost constant. In Figure 5(b), for refrigerant mixtures, the saturation temperature T~= gradually decreases in the direction of flow of refrigerant, and the temperature difference (T,,~-Ts) is kept nearly constant in the condensing region by regulating the flow-rates of refrigerant and heat sink water. The heat flux density q gradually decreases in the direction of flow of refrigerant. Figure 6(a) and (b) shows the axial distributions of the outer wall surface temperatures for pure R22 and the mixture 52 mol% R22-48mol% R114, respectively. In Figure 6(a), which corresponds to the test run shown in Figure 5(a), the circumferential temperature variation is small near the refrigerant inlet and gradually increases in the direction of flow of refrigerant. This suggests that the flow pattern is annular near the inlet and changes into a
260
semi-annular flow as condensation proceeds. In the semi-annular flow region (Sections 6-9), the wall temperature decreases circumferentially in the order right side, top, bottom and left side. This fact allows the inference that the liquid film is deformed circumferentially by the grooves cut like a right hand screw. In Fi,qure 6(b), which corresponds to the test run shown in Fiyure 5(b), the circumferential temperature difference in the condensing region is small over the whole condensation region. Stoecker et al. 2 and Tandon et al? reported that the flow pattern is not affected by the composition of the refrigerant mixture. In the semi-annular flow region of refrigerant mixtures the temperature at the vapour liquid interface probably varies circumferentially. Figure 7(a) (e) shows a comparison between the local Nusselt number (NU)exp obtained experimentally and (Nu)ca I estimated from an empirical equation for pure refrigerants inside a horizontal smooth tube 6. The local Nusselt number is defined as:
Int. J. Refrig. 1990 Vol 13 July
(18)
where )-L is the thermal conductivity of liquid. In Fiyure 7(a) and (b) the (NU)exp values for R22 and Rl14 are higher than the (NU)cal values by 40-70% and 50-80%, respectively. These facts reveal the effectiveness of condensation enhancement due to the internal spiral grooves. The ratios (Nu)exp/(Nu)ca~ for the refrigerant mixtures are smaller than those for the pure refrigerants and the axial variations of the ratio are affected by both the bulk concentration and the flow-rate of refrigerant as shown in Fiqure 7(c)-(e).
Average heat transfer Figure 8 shows a comparison between the measured average Nusselt number of the pure refrigerants and the following empirical equation, which was obtained for the condensation of pure organic substances inside a horizontal smooth tube by Fujii and Nagata 7.
where
Nu = .
(20)
AL
Ph = Re z-
CpL(Tsat- "Fwi) (hvsat)z = 0 - (hLsa,)z= I
PL Uvl /~L
(21)
(22)
R --[ DLIIL] 1/2 - LP~_I
(23)
p r L = #LCpL
(24)
2L
Nu is the average Nusselt number, Ph is the phase change number, Re~ is the two-phase Reynolds number, R is the p-/~ ratio, Pr L is the Prandtl number of liquid, Uv is the mean vapour velocity at the refrigerant inlet, and the subscript sat denotes the saturation state. The measured values are about 100% higher than the values predicted by Equation (19). This shows that the internal spiral
Condensation heat transfeY of binary refrigerant mixtures." S. Koyama et aL 10 3
' o.
I/'./'
'
xl 0 2
' V/ -!
R22
/-
71
,
i
I
'
r d . 50molXR22+50rnol%R114
51-
#yo
.
-
~5
#~
/
o :,o.,
= "
,
I
3
4 5 (Nu)cal
'
'
'
v
'
x"~
2 ~°z~
z "J4
/- ~ ' / ~
-
.
A
: 47.7 '~ : 50.7 ¢. : 51.3
/ /
,
I
I
2
3
4
5
7~
I
5
7x102
I
sl/ ~" 4t-
: 28.1
#!y
I
I
''/~rl~/'/W
,, :31.3 0 : 32.4 • : 37.6 ,t : 39.3
I ~"
I
I
.
-
./
×Zo v
I/
I
I
e. 25molgR'22+75molgR114
An O . /
r kg/h o:39,9
p # /
40.2
o : 46.8
I
103
(Nu)cel xl 0 2 7
I
4
o:30.2
/
/
I
3
xlO 2
7
o:26.1-
/
~z'~"
/
2
(Nu)col
//
/
~/~1--
o 57.5
/
__;~¢'~ ./ Wr kg/h 4~[~
o 45.6 ~ 53.4
:Yl
•
i/ -
37.8
~,/o/. /
-
..-r.~ 5
o
:~ .
I #Z I p ~ "~ 10 3
'
b. R 1 1 4 10 3
:68.1
•
A : 70.0 m : 70.3 ,,,,
•
/
I/~:~
2I~
I
,
I ,
2
3
4
5
7xi02
(Nu)cal i
,
I
i
-- c. 75molgR22+25molgR114
5 ~4
-
-
--
/~_/<~.~'~t/
Wr kg/h
o%'~
o: ~o.o
~s
:36.5
o ;:?5 L-/
2
• :58.1 ,t : 66.2 I
3
I
]
4
5
Figure 7 Comparison between (Nu)cxp and (Nu)c,r (a) R22; (b) Rl14; (c) 7 5 m o l % R 2 2 - 2 5 m o l % R l 1 4 ; (d) 5 0 m o l % R 2 2 - 5 0 m o l % R l 1 4 ; and (e) 25mol oVoR22-75mol oVoR114. Figure 7 Comparaison entre (Nu),,p et (Nu)c, l
I
ix10 2
(Nu)co,
grooves are effective in enhancing condensation. The experimental results for the pure refrigerants can be correlated well by the following equation, the functional form of which is the same as Equation (19): N
.
,.,I/ l'~-°4n,_o 6['Re, PrL'~°-a
"--°"W
"" t - - F - )
(25)
Figure 9 compares the experimental results for the refrigerant mixtures with Equation (25). The experimental data for the mixed refrigerants are lower than the values calculated with Equation (25). Fioure 10 shows the relationship between the ratio (Nu),,~/(Nu)ca~ and bulk molar fraction of Rll4, yb, where (Nu)=,p is the measured value and (Nu)¢al is the value predicted by Equation (25). These data can be
approximated by the following equation: (Nu)exp fy = 1 - 0.73y b + 0.37y 2 + 0.36y 3
(Nu)o,~
(26)
All of the experimental data for the pure and mixed refrigerants can be correlated well by the combined equation of Equations (25) and (26), as shown in Figure
11. Pressure drop Figures 12(a)-(e) shows the experimental results for the frictional pressure drop plotted in the coordinate of Lockhart-Martinelli parameters @v and X., which are defined as:
=[APEIAz)~ u2 ,, L(APIAz),, j
Rev. Int. Froid 1990 Vol 13 Juillet
(27)
261
Condensation heat transfer of binary refrigerant mixtures: S. Koyama et aL 5
o
'
I
'
'
'
' 'I
4
0 R22 A Rl14
o Z2 ~co c5 A
~
I
'
E~l'"q
1
R22
10 -
co
' 'I'"'I
(a)
l
'
~ /
"
,R,,',.,/' / >
2
#
J
£L ~'~ I0 6
• present
-
1
~0~
A Hash, z u m e
0.01
0.1
Z
1.0
5
Xtt I
53
,
,
,/~1
4- 5
10 8
2
3
Re~ ' '
Figure 8 Average Nusselt n u m b e r (pure refrigerants)
I .... I
' '-'1 .... t
-L
*
5
o A
[
4
~
=
i
~
.~'~
1
= I ll=14
_
U present
,
,~ Hashizume ,I,,,,I L t
,
0.1
0.01
:2f
CL ~'~-106
.
~50~ =
2
o
(b)
I
E:] 75mol%R22+25mol%R114 0 50mol%R22+50mol%R114 ~7 25mol%R22+75mol%R114
c5 co
i
' '
10 __- R114
Nombre de Nusselt moyen (frigorigbnes puts)
Figure 8
1.0
5
Xtt
~k~/~-/djo O v
Iz
(c) I , ~ ~1
5-
5
75mol%R22+25mol%R114
10
108
Figure 9
Average Nusselt n u m b e r (refrigerants mixtures)
Figure 9
Nombre de Nusseh moyen (mblanges de frigorigbnes)
> ....~ -~
1.2
. . . .
I
£~ Hashi zurne I
v
z~
-
i
;
0.1
0.01
o 5o-5o (~u)o~p/(~u)~o,
• present
o
I /IIIHi
I
R22+R114 (mot%) 0 100- 0 {3 7 5 - 25
1.1 "~1.0
. . . .
~0~
lIH]
~
I
1.0
5
Xtt
2 5 - 75 0-100
: 1 _ 0 . 7 3 yb +0.37 yt? + 0 , 3 6 yb-~
-
(d)
~o.9 %-
10
50mol%R22+50mol%R114/p
iz > 0.7
,
,
,
R22
, (~ . . . . O.5
Rt 14
)
Molar Fraction
Figure 1O R e l a t i o n s h i p between (Nu)~p/(Nu)~,~ a n d y~ Figure 10
-
• present A Hashizume
-~-~ 50% I
I I]1111l
I
I ,I,,,,I J
0.1
0.0
Relation entre (Nu)~v/(Nu)~.l et Yb
,
1.0
5
Xtt 5
9--.
'
4-
I ' ' ''1
R22+R114(molX)
' "i
....t
'
" I ....I
®
/z9" Z~
co
[2-
~
10 6
0--100
£3~,
~
-
12 53
I
4- 5
0.01 ,
,r,l
108 Re~
½
• present
Figure
Average Nusselt n u m b e r pure a n d m i x e d refrigerants
F i g u r e 11
Nombre de Nusselt moyen des frigorigbnes purs et des
mblanges
1990
Vol
13
July
0.1
_
z~ H o s h i z u m e
-~
1.0
5
Xtt
.3
Figure 11
Int. J. Refrig.
.../~ ~
1 --~30%
7.
262
(e)
' "
10 : 25mol%R22+75mol%R114~....
5
12 F r i c t i o n a l pressure drop. (a) R22; (b) R l l 4 ; (e) 75mo1% R 2 2 - 2 5 m o l % R114; (d) 50mo1% R 2 2 - 5 0 m o l % R114; a n d (e) 2 5 m o 1 % R 2 2 - 7 5 m o l % R114 F i g u r e 12 Chute de pression par frottement
Condensation heat transfer of binary refrigerant mixtures: S. Koyama et al. 1
are correlated fairly well by the Lockhart-Martinelli parameters independently of molar fraction.
X 0.9 Pv 0.5 flL 0.1
The frictional pressure gradient ( A p v / A z ) is calculated by Equation (14) and the pressure gradient for vapour How (Ap/Az)v is calculated by the following equation: ~-z v -
diPv
(29)
where fv is the friction coefficient estimated by the Colburn's equation as follows: fv = 0.046/{ Grxdi/#v} 0.2
(30)
Figure 12(a)-(e) corresponds to the results of R22, R 114, 75 mol% R22-25 mol% R 114, 50mo1% R22-50mol% R 114 and 25mo1% R22-75mol% Rl14, respectively. (O) denotes the present results and (A) denotes the experimental results by Hashizume s for the adiabatic two-phase flow of R22 in a horizontal smooth tube, rearranged by the authors, and the solid lines denote the Lockhart-Martinelli relation 9. The present results are correlated fairly well by the Lockhart-Martinelli parameters independently of bulk concentration, and are about 40% higher than Hashizume's results s due to the effect of grooves.
Acknowledgements
This study was supported by the Grant-in-Aid of Scientific Research of the Ministry of Education, Science and Culture, Japan, and was also sponsored by the Japanese Association of Refrigeration. References
1 2 3 4
5 6 7
Bokhanovskiy,YU. G. Heat transfer for Freon-12, Freon-22and their mixtures in a coiled tube condenser Heat Transf Soy Res (1980) 12 43-45 Stoecker,W. F., Kornota,E. Condensingcoefficientswhen using refrigerant mixture ASHRAE Trans (1985) 91 1351-1367 Tandon,T. N., Varnma, H. IL, Gnpta, C. P. Prediction of flow patterns during condensation of binary mixtures in a horizontal tube Trans ASME J Heat Transf (1985) 107 424-430 Tandon, T. N., Varuma, H. K., Gnpta, C. P. Generalized correlation for condensationof binary mixtures inside a horizontal tube Int J Refrio (1986) 9 134-136 Mochizuki, S., Inone, T., Tomiuga, M. Condensation of nonazeotropicbinarymixtures in a horizontal tube TransJSME Part B (1988) 54(503) 1796-1801 Fujii,T., Honda, I-L, Nozu, S. Condensation of fluorocarbon refrigerants inside a horizontal tube Refrigeration JAR (1980) 55(627) 3-20 Fujli,T., Nagata, T. Condensation of vapor in a horizontal tube Report of Research Institute of Industrial Science K yushu University (1973) 52 35-50
Conclusions
An experimental study of condensatin of refrigerant mixtures of R22 and R114 inside a horizontal tube with internal spiral grooves for a double-tube counterflow condenser was performed. The conclusions are as follows. 1. The axial distributions of saturation temperature and tube wall temperature for the refrigerant mixtures are different from those for the pure refrigerants. 2. In the case of the pure refrigerants, the local Nusselt number for a grooved tube is about 60% higher than that for a smooth tube. The distribution of local Nusselt number for the refrigerant mixtures is different from that for the pure refrigerants and depends on the bulk composition and flow-rate of the refrigerant. 3. The average heat transfer coefficient for the mixed refrigerants is lower than that for pure refrigerants, and the degree of maximum decrease is about 20%. The average heat transfer coefficient for the pure and mixed refrigerants can be correlated well by Equations (25) and (26). 4. The present results for the frictional pressure drop
8 9 10 11
12 13 14 15
Hashizmne,K. Flow pattern, void fraction and pressure drop of refrigerant two-phaseflowin a horizontal pipe-IInt J Multiphase Flow (1983) 9 399-410 Lockhart,R. W., MartinelU, R. C. Proposed correlation of data for isothermal two-phase, two-component flow in pipes Chem En9 Pro9 (1949) 45 39 Nishinmi,H., Saito, S. An improved generalized BWR equation of state applicable to low reduced temperatures J Chem Eng Jpn (1975) 8 356-360 Fukuzato,R., Tomisaka, Y., Arai, K., Saito, S. The prediction of binary vapor-liquid equilibrium for fluorocarbon mixtures by use of a modifiedBWR equation of state J Chem Eno Jpn (1983) 16 147-149 Hasegawa,N., Uematsu, M., Watanabe, K. Measurements of PVTx for the R22+RI14 system J Chem En0 Data (1985) 30 32-36 Smith,S. L. Void fraction in two-phaseflow:a correlation based upon an equal velocity head model. Heat Fluid Flow (1971) 1 22-39 Reid,R. C., Prausnitz, J. M., Sherwood,T. K. The Properties of Gases and Liquids McGraw-Hill New York (1977) Chen,Z., Fujii, T., Fujii, M. A methodfor evaluating the mixture constant in the expression for thermal conductivity of binary liquid mixture and its extension to multi-component liquid mixtures Report of Research Institute of Industrial Science Kyushu University (1987) 82 173-187
Rev. Int. Froid 1990 Vol 13 Juillet
263