Heat transfer, pressure drop, and flow pattern recognition during condensation inside smooth,...

International Journal of Refrigeration 30 (2007) 609e623www.elsevier.com/locate/ijrefrig

Heat transfer, pressure drop, and flow pattern recognitionduring condensation inside smooth, helical micro-fin,

and herringbone tubes

Jonathan A. Oliviera, Leon Liebenberga, John R. Thomeb, Josua P. Meyera,*

aDepartment of Mechanical and Aeronautical Engineering, University of Pretoria, Pretoria, South AfricabLaboratory of Heat and Mass Transfer, Faculty of Engineering Science, Swiss Federal Institute of Technology-Lausanne (EPFL),

Lausanne 1015, Switzerland

Received 19 May 2006; received in revised form 31 October 2006; accepted 1 November 2006

Available online 17 January 2007

Abstract

This paper presents a study of flow regimes, pressure drops, and heat transfer coefficients during refrigerant condensationinside a smooth, an 18� helical micro-fin, and a herringbone tubes. Experimental work was conducted for condensing refrigerantsR-22, R-407C, and R-134a at an average saturation temperature of 40 �C with mass fluxes ranging from 400 to 800 kg m�2 s�1,and with vapour qualities ranging from 0.85 to 0.95 at condenser inlet and from 0.05 to 0.15 at condenser outlet. These test con-ditions represent annular and intermittent (slug and plug) flow conditions. Results showed that transition from annular flow tointermittent flow, on average for the three refrigerants, occurred at a vapour quality of 0.49 for the smooth tube, 0.29 for thehelical micro-fin tube, and 0.26 for the herringbone tube. These transition vapour qualities were also reflected in the pressuregradients, with the herringbone tube having the highest pressure gradient. The pressure gradients encountered in the herringbonetube were about 79% higher than that of the smooth tube and about 27% higher than that of the helical micro-fin tube. A widelyused pressure drop correlation for condensation in helical micro-fin tubes was modified for the case of the herringbone tube. Themodified correlation predicted the data within a 1% error with an absolute deviation of 7%. Heat transfer enhancement factors forthe herringbone tube against the smooth tube were on average 70% higher while against the helical micro-fin tube it was 40%higher. A correlation for predicting heat transfer coefficients inside a helical micro-fin tube was modified for the herringbonetube. On average the correlation predicted the data to within 4% with an average standard deviation of 8%.� 2006 Elsevier Ltd and IIR. All rights reserved.

Keywords: Refrigeration; Air conditioning; Condensation; R-22; R-407C; R-134a; Smooth tube; Micro-finned tube; Ribbed tubes;

Heat transfer; Pressure drop; Flow

* Corresponding author. Tel.: þ27 012 420 3104; fax: þ27 012 362 5124.

E-mail address: [email protected] (J.P. Meyer).

0140-7007/$35.00 � 2006 Elsevier Ltd and IIR. All rights reserved.

doi:10.1016/j.ijrefrig.2006.11.003

mailto:[email protected]

http://www.elsevier.com/locate/ijrefrig

610 J.A. Olivier et al. / International Journal of Refrigeration 30 (2007) 609e623

Transfert de chaleur, chute de pression et comportement del’ecoulement lors de la condensation a l’interieur

des tubes lisses, helico€ıdaux, a micro-ailettes et a chevrons

Mots cles : Froid ; Conditionnement d’air ; Condensation ; R-22 ; R-407C ; R-134a ; Tube lisse ; Tube micro-ailete ; Tube nervure ;

Transfert de chaleur ; Chute de pression ; Ecoulement

Nomenclature

A area (m2)B fin base thickness (m)Bo Bond numberC constantcp specific heat (J kg�1 K�1)D diameter (m)e fin height (m)f friction factorFr Froude numberFt Froude rateG mass flux (kg m�2 s�1)g gravitational acceleration (m s�2)h enthalpy (J kg�1 K�1)L length of test section (m)_m mass flow (kg s�1)n number of finsDp pressure drop (Pa)Pr Prandtl numberQ heat transferred (W)Re Reynolds numberT temperature (�C)U overall heat transfer coefficient

(W m�2 K�1)u velocity (m s�1)Xtt Martinelli parameterx quality

Greek symbolsa heat transfer coefficient (W m�2 K�1)b helix angle (degrees)f two-phase multiplierg apex angle (degrees)l conduction coefficient (W m�1 K�1)m viscosity (Pa s)r density (kg m�3)s surface tension (Pa)

SubscriptsCu copperdew dew pointeq equivalentf frictionfa actual flow areaglide temperature glidei inner/inletia intermittent-annularL liquidlmtd logemean temperature differencelo liquid onlyn nominalsat saturationV vapourvo vapour onlyw wall

1. Introduction

1.1. Enhanced tubes

The use of heat transfer enhancement devices for con-densers is common, and the use of helical micro-fin tubes(cf. Fig. 1) is the most prevalent passive enhancement devicein use today. A study conducted by Liebenberg et al. [1,2]using helical micro-fin tubes with an inner diameter of8.9 mm, showed that these tubes have a heat transfer coeffi-cient increase of about 200% compared to that of a smoothtube. However, this increase in heat transfer had to contendwith a trade-off, namely a corresponding increase in

pressure drop of about 100% compared to a smooth tube.These pressure drop increases were attributed to the in-creased vapour velocities in a helical micro-fin tube con-denser, brought about by the greater regions of annularflow, which in turn increases the turbulence inside the tubecompared to a smooth tube. The fins redistribute the liquidlayer around the circumference of the tube, forcing theflow to become annular rather than intermittent or stratified.

In the mid-1990s a new generation of micro-fin tube wasdeveloped, most notably the herringbone tube. This tubeconsists of a double chevron-shaped tube, with embossedmicro-fins, as shown in Fig. 1. Due to the effect of gravity,especially at low velocities, the distribution of liquid at the

611J.A. Olivier et al. / International Journal of Refrigeration 30 (2007) 609e623

bottom of the tube will be higher than at the top [3]. The heattransfer enhancement, which can be as high as 350%, can beattributed to the presence of a thin film layer on the tubesides and the mixing of the converging liquid at the topand the bottom of the tube [4]. The effect of tube orientationon heat transfer is in fact negligible while its effect onpressure drops is very small [4]. Miyara et al. [4] also notedthat fin height and helix angle have a large effect on theheat transfer and pressure drop. The greater the helix anglethe greater the heat transfer and pressure drop. Fin heightswere shown to increase heat transfer up to 0.18 mm withfurther enhancement being negligible with height increase.The increase of fin height also increased the pressuregradient.

Previous researchers [5e7] worked in the mass fluxrange of 100e400 kg m�2 s�1 during condensation research

Fig. 1. Pictures of a herringbone tube (top) and a helical micro-fin

tube (bottom).

investigations, which is lower than those typical for someheat pump water heater applications, i.e. from 500 to1000 kg m�2 s�1. Although smaller tubes are being investi-gated by other researchers, especially in countries such asthe USA and in Asia where tube diameters as low as 4 mmare being introduced, 3/8’’ (9.5 mm) outside diameter tubesare still the most common tubes used in residential andcommercial air conditioners, heat pumps and refrigerationsystems [8].

1.2. Flow pattern maps

Numerous flow pattern maps have been proposed overthe years for predicting two-phase flow regime transitionsin horizontal tubes under adiabatic and diabatic conditions[1,2]. Kattan et al. [9e12] proposed the first comprehensiveflow-boiling model for evaporation inside horizontal tubesbased on the local flow pattern and a newly developed dia-batic flow pattern map. Their new approach resulted in sig-nificant improvements in the accuracy and reliability of heattransfer predictions compared to previous methods. As a con-sequence, several new condensation heat transfer modelsbased on local flow patterns have been proposed by Shao andGranryd [13] and Cavallini et al. [14]. El Hajal et al. [15]adapted the flow-boiling two-phase flow pattern map, origi-nally developed by Kattan et al. for condensation inside hor-izontal tubes. El Hajal et al. [15] and Collier and Thome [16]proposed a new logemean method for predicting void frac-tions for pressures ranging from atmosphere up to near thecritical pressure. This was prompted by the high reducedpressures during condensation and because of the sensitivityof flow pattern transitions, heat transfer coefficients andpressure drop to void fraction. The Thome map [17] isa modification of the Steiner [18] map, which in turn isa modification of the original Taitel and Dukler [19] map.El Hajal et al. [15] simplified the implementation of themap by bringing the Rouhani and Axelsson [20] voidfraction equation into the method to eliminate its iterativesolution scheme. Also, the changes proposed by Zurcheret al. [21] were implemented to ensure better prediction ofthe transition curves from annular to stratified/wavy flow,and from stratified/wavy to wavy flow.

El Hajal et al. [15] and Thome [17] show that for conden-sation, saturated vapour enters a condenser tube and formseither (i) a thin liquid film around the perimeter of thetube as an annular flow or (ii) a liquid layer in the bottomof the tube and a gravity-controlled condensing film aroundthe upper perimeter as a stratified or stratified-wavy flow.Therefore, the transition curve from annular and intermittentflow to stratified/wavy flow can be supposed to reach its min-imum value and then continue on horizontally to the vapourquality of 1.0. This means that a saturated vapour enters atx¼ 1.0 and goes directly into either the annular flow regimeor the stratified-wavy flow regime, depending on whether themass flux is greater or less than the transition curve (cf.Fig. 3).


1.3. Purpose of this paper

By making use of the flow pattern maps of Thome [17], itwill be shown from the results of condensation experiments,that flow patterns have a strong influence on heat transferand pressure drop of smooth and enhanced tubes. Themain purpose, therefore, is to present data obtained fromin-tube condensation experiments for mass fluxes rangingfrom 400 to 800 kg m�2 s�1. The flow regimes, pressuredrops, heat transfer coefficients, and the overall performanceof three different tubes, namely a smooth, a helical micro-fin, and a herringbone tube (each having a nominal diameterof approximately 9.51 mm), will be presented and com-pared. The refrigerants used during all the experimentswere R-22, R-134a, and R-407C all at an average saturationtemperature of 40 �C.

2. Experimental facility

The experimental test facility consisted of two main sub-systems: the vapour-compression loop, and the water loops.Each of the sub-systems was of the closed loop type. Thevapour-compression loop consisted of a hermetically sealedreciprocating compressor, an oil separator, a manuallyadjustable expansion valve, a water-heated evaporator anda water-cooled test condenser. Three test condensers wereused, namely a smooth tube, an 18� helical micro-fin tube,and a herringbone micro-fin tube. Geometric parameters ofthe tubes as well as their lengths are given in Table 1. A sche-matic layout of the system is given in Fig. 2. Steady stateconditions were assumed when the energy balance errorwas less than 1% and temperatures and pressures wereconstant for a period of at least 10 min. Further, experimen-tal conditions were also maintained so that the inlet qualitieswere between 0.85 and 0.95 and outlet qualities werebetween 0.05 and 0.15.

The orientation of the herringbone micro-fin tube waschosen such that the liquid converged at the top and the

bottom of the tube and diverged at the sides. Visual inspec-tion ensured that the orientation of each tube was maintainedduring the manufacture of the condenser. Short-sightglasses, cylindrical in shape, were mounted in close proxim-ity to the temperature and pressure sensors and were used tovisualise the flow patterns inside the tubes. The inner diam-eter of these sight glasses were the same as the inner diam-eter of the condenser tubes, so that the flow inside the tubeswould not be affected. The sight glasses were mounted justbefore the U-bends connecting the condenser sub-sections,so that there was at least a 1-m spacing between sight glassesin order to negate the flow effects that U-bends induce due togeneration of non-negligible centrifugal forces. (Tube bendstypically affect the local flow conditions, but flow typicallyreturns to its original (pre-bend) characteristic, approxi-mately 50 tube diameters downstream [22].)

The test condenser was of the tube-in-tube type withrefrigerant flowing in the inner tube and water flowing ina counterflow direction in the annulus. The cold- and hot-water loops were connected to the condenser and evaporator,respectively. On the condensing side, the cold water wasused as a heat sink, removing the latent heat from the con-densing refrigerant. The water temperature was kept con-stant in the range of 20e25 �C (depending on theexperiments conducted) in a 1000 L insulated reservoir con-nected to a 15 kW chiller. The water gauge pressure in theannulus was maintained between 70 and 140 kPa to preventthe formation of air bubbles that could affect temperaturereadings and the heat transfer from the refrigerant to thewater. A centrifugal pump pumped the chilled water to thedouble-tube condenser. A hand-controlled valve controlledthe water flow rate through the test sections. After passingthrough the condenser, the water returned to the reservoirof the chiller unit.

A similar hot-water flow loop was used on the evaporat-ing side, also with an insulated 1000 L reservoir, but con-nected to a 12 kW electric resistance heater. The reservoirwater temperature was varied between 30 and 40 �C

Table 1

Inner tube geometric parameters of the test condenser

Smooth Helical Herringbone

Material (Copper) Hard-drawn Hard-drawn Soft-drawn

Helix angle, b (degrees) e 18 16

Apex angle, g (degrees) e 40 25

Number of fins e 60 70

Fin base thickness, b (mm) e 1.672 0.887

Outside diameter, Do (mm) 9.52 9.55 9.51

Inside diameter, Di (mm) 8.11 8.94 8.52

Tube wall thickness, tw (mm) 1.4 0.307 0.3

Equivalent diameter, Deq (mm) 8.11 8.79 8.82

Fin height, e (mm) e 0.209 0.2

Actual flow area, Afa (�10�6 m2) 51.7 60.64 61.16

Condenser sub-section length (m) 1.5 1.13 0.563

Condenser total length (m) 12 9 4.5


Fig. 2. Schematic of the experimental layout for the herringbone tube.

depending on the experiments conducted. Increasing or de-creasing the temperature of the water through the evaporatoraltered the refrigerant density at the compressor inlet andthus the refrigerant mass flow.

For the smooth and the helical micro-fin tubes, tworesistance temperature detectors (RTDs) were used priorto each sub-section, placed at the top and the bottom ofthe inner tube. This was done to obtain an average repre-sentative refrigerant temperature since the distribution ofthe liquid layer inside the tube would under- or overesti-mate the temperature measurement if only one RTD wasused. For the herringbone tube, however, the RTDs wereplaced at the top and on the side of the inner tube sincethe liquid film thickness on the top is much thicker thanon the side.

The absolute pressures of the condensing refrigerantwere measured with piezoelectric pressure transducers,which were positioned at the inlet of each condenser sub-section. Two Coriolis mass flow metres were used for thevapour-compression loop and the cold-water loop. The

uncertainties of the instruments are given in Table 2and were calculated by using the method of Kline andMcClintock [23].

3. Data reduction

The properties of the refrigerant at the inlet and outlet ofthe condenser were determined by temperature and pressuremeasurements. From these measurements, the thermo-physical properties of the condensing refrigerant were deter-mined by interpolating the super-heated (at the inlet of thecondenser) and sub-cooled (at the outlet of the condenser)tables that were obtained from a refrigerant propertydatabase [24]. The refrigerant properties for the rest of thecondenser (two-phase sections) were determined by assum-ing that the calculated values of the heat transferred to thewater side was more accurate than the heat transferredfrom the refrigerant side. With this assumption the enthalpyvalues of the refrigerant could be deduced. In the first testsub-section, the water heat flux was equated to the


refrigerant heat flux (due to the refrigerant enthalpy change)to obtain the outlet enthalpy, ho. This outlet enthalpy wasthen used as the inlet enthalpy, hi, for the next sub-section.This procedure was repeated for all the eight sub-sections.

The average sectional vapour quality was thus obtainedby:

xo ¼ho � hL

hV � hL

with hL and hV measured at To

xi ¼hi � hL

hV � hL

with hL and hV measured at Ti

The average vapour quality of each test sub-section wasthen determined by the arithmetic mean of xi and xo. Althougha small change in vapour quality in the test sub-sections wouldhave been preferred, the authors also attempted to simulatea realistic condenser operation by using realistically longtube lengths, which introduced a 16e24% change in averagevapour quality in the two-phase sub-sections.

The pressure drops were determined by subtracting thepressure readings before and after each test section. Thepressure gradients were calculated by dividing the pressuredifference with the sub-section length. With these pressuregradients the penalty factors could be determined. The pen-alty factor is defined as the ratio of the pressure gradient ofthe tube being examined to that of a reference tube.

Flow regime observations were made with the aid ofvideo cameras. Of interest was the transition from annularto intermittent flow which was deduced with the aid of theThome maps [17], the Froude rate, and the Martinelli para-meter. The Froude rate is given by

Ft ¼�

G2x3

ð1� xÞr2VgDi

�1=2

ð1Þ

which is a ratio of gravitational energy dissipation to inertialenergy. The Martinelli parameter, which is given by

Xtt ¼�

1� x

x

�0:9�rV

rL

�0:5�mL

mV

�0:1

ð2Þ

Table 2

Experimental uncertainties

Measurement Uncertainty

Refrigerant temperature 0.14 K

Water temperature 0.11 K

Saturation temperature 0.12 K

Pressure 0.23%

Refrigerant mass flow 0.23%

Water mass flow 0.28%

Average quality 3.0%

Viscosity 0.10%

Density 0.03%

Reynolds number 1.0%

Heat transfer coefficient 9e11.30%

Pressure drop 6.80%

is the ratio of viscous dissipation to inertial energy. Thesetwo parameters were combined to predict the transitionfrom annular flow (viscous forces predominate) to intermit-tent flow (gravitational forces predominate) [2].

By making use of the modified Wilson Plot technique[25], the unknown Reynolds number coefficient as well asthe annulus side heat transfer coefficient expressed in termsof a Dittus and Boelter-type equation could be determined.The annulus side heat transfer coefficient was found to be

ao ¼ 0:0936lL

Di

Re0:742lo Pr1=3

L

�mL

mL;w

�0:14

ð3Þ

With this the average refrigerant side heat transfercoefficient could be calculated:

ai ¼1

Ai

�1

UA� 1

aoAo

� lnðDo=DiÞ2plCuL

��1

ð4Þ

The overall heat transfer coefficient, U, based on the out-side area of the inner tube, was determined by:

U ¼ Q

AoTlmtd

ð5Þ

The value of Q was taken from the water side due to thefact that the temperature differences were much greater frominlet to outlet than for the refrigerant side. This gave rise toa much lower uncertainty in the calculation of Q.

For pure refrigerants, the saturation temperature was ob-tained from the average of the refrigerant bulk temperaturesat the inlet and outlet of the condenser test section. For zeo-tropic mixtures, the correlation of Cavallini et al. [8] wasused to determine the saturation temperature:

Tsat ¼ Tdew �DTglideð1� xÞ ð6Þ

where DTglide is the temperature glide, while x is an empir-ical function of the vapour quality that varies from 0 to 1.The empirical function was determined by a curve fitthrough data generated from a refrigerant database [24].

4. Flow regime observations and the effects of massflux and vapour quality variations

In Fig. 3, the vertical line for the smooth tube and the for-mat of the flow regime map were first published by Thome[17]. It was visually observed that the micro-fins (in the he-lical micro-fin and herringbone tubes) delay the transitionfrom annular flow to intermittent flow, which enhances theheat transfer relative to the smooth-tube counterpart. Thevisual observations showed that the intermittent flow regimeis neither uniform nor steady, and it is easy to deduce thatempirical methods will be unable to completely definethis, as they would only generate temporal and spatial aver-ages. Video images were superimposed on the flow patternmap shown in Fig. 3, and this corroborated the annular/intermittent flow transition criteria. Fig. 4 shows the resultsfor the herringbone tube (with only a few flow regimeimages) using R-407C.


Mas

s F

lux,

kg

m−2

s−1

Fig. 3. Comparison of experimental flow pattern transitions for condensation of R-22, R-134a and R-407C in smooth, helical micro-fin, and

herringbone tubes. The annular-to-intermittent transition line for R-407C (herringbone) lies directly on top of those for R-22 (herringbone) and

R-134a (helical). The transition lines for R-407C (helical) and R-22 (helical) are much closer but the discrepancy is exaggerated in this

diagram for convenience.

Thome [17] and the current authors obtained the transitionbetween intermittent flow and annular flow, xia, in a smoothtube as a vertical line (Fig. 3) which was determined bysetting the Martinelli parameter Xtt equal to 0.34, as:

xia ¼��

C

�rL

rV

�0:5�mV

mL

�0:1�þ 1

��1

ð7Þ

with C¼ 0.3218 (smooth tubes).

Liebenberg et al. [2] adapted the Thome [17] smooth-tube transition criteria for condensation of R-22, R-407Cand R-134a in a range of helical micro-fin tubes, and the cur-rent authors did the same for refrigerant condensation insideherringbone tubes; the resulting transition criteria featureonly different coefficients:

Helical micro-fin tubes: C¼ 0.6778Herringbone tubes: C¼ 0.7898.

Fig. 4. New flow regime map showing transition qualities for condensation in the herringbone tube for R-134a with superimposed video

images at 300, 500, and 800 kg m�2 s�1.


The combined flow transition criteria for refrigerant con-densation inside horizontal tubes are superimposed on theThome map [17] (which is only intended for smooth tubes)as given in Fig. 3. It is apparent that the transition linesbetween annular and intermittent flows are close for helicalmicro-fin and herringbone tubes, and far removed from theirsmooth-tube counterpart, which implies that the finnedconfigurations bring about a delay in transition from annularto intermittent flow.

The flow patterns at a mass flux of 400 kg m�2 s�1 weresimilar to those at 500 kg m�2 s�1 except that for the formerthere is no apparent redistribution of the condensed liquid atthe top of the tube. However, for the herringbone tube thethick annular film does become unstable at vapour qualitiesaround 0.45, as compared to 0.75 for helical micro-fin tubesand 0.85 for smooth tubes. Transition from annular to inter-mittent flow occurs at around x¼ 0.3 for herringbone tubescompared to 0.45 for helical micro-fin tubes. Stratifiedwaves occur at very low vapour qualities of about 0.1, afterwhich the flow becomes fully condensed, but with theappearance of an occasional elongated slug, and chaoticbubbles. It is evident that the micro-fin tube condensationprocess becomes stretched out with a much greater portionof the herringbone tube subjected to annular and intermittentflows, in comparison to the other two tubes [2]. These resultsfor the herringbone tube compare well with the data ofMiyara et al. [4] who reported that at a mass flux of300 kg m�2 s�1 liquid was still being collected at the topand bottom of the tube at qualities as low as 0.2 due to therelatively high vapour velocity.

At 800 kg m�2 s�1, the lower vapour quality regions areseen to experience turbulent slugs for herringbone tubes,with swirling bubbles that eventually become elongatedslugs for helical micro-fin tubes and just occasional bubblesfor smooth tubes. In the case of helical micro-fin tubes, theseslugs become very thin, swirling, vapour streaks just beforethe flow becomes fully condensed whereas in the herring-bone tube, more violent slugs are still witnessed [2].

At mass fluxes greater than 400 kg m�2 s�1, the visualobservations reveal that flow regimes included intermittent(slug, plug) flow at vapour qualities below approximately0.27, and annular flow for approximately 0.27< x< 0.8.

The analysed images reveal insensitivity of flow patternsto the three used refrigerants at any given vapour quality.This could be attributed to the fact that the transition points(from intermittent to annular flow) for the different refriger-ants are very close.

5. Experimental pressure drop results

Prior to helical micro-fin and herringbone tube tests, con-densation experiments were conducted on a smooth-tube testsection. The pressure drop data were compared to data ob-tained from Cavallini [26]. Confidence in the experimentalsystem was confirmed due to the fact that the experimentaldata did not deviate from the given data by more than 12%.

For the experimental work performed on the helical micro-fin tube, the data were compared to the correlation ofCavallini et al. [8] that did not deviate from this correlationby more than 13%. For the experimental work performed onthe herringbone tube, the data were compared to the correla-tion of Miyara et al. [6]. Although Miyaras correlation wasbased on experimental work performed at mass fluxes between100 and 400 kg m�2 s�1, the current authors experimentaldata underpredicted the correlation by no more than 7%.The current author’s experimental work was performed atmass fluxes between 300 and 800 kg m�2 s�1 that confirmedthat Miyaras correlation was valid at higher mass fluxes.

5.1. Pressure gradients

Figs. 5, 6, and 7 give the experimental pressure gradientsfor refrigerants R-22, R-134a, and R-407C, respectively, forall the three tubes. The overall trend is a decrease in pressuregradient as the vapour quality decreases. With respect to thetransition qualities, the pressure gradients in the annular flowregime are higher than those in the intermittent flow regime.At high qualities the pressure gradients are high due to thefriction generated by the high vapour velocities (shear-dominated flow), while at low qualities the flow is gravity-dominated, and hence the gradients being lower than thoseencountered at high qualities. It is interesting to note fromthese figures that the pressure gradients at low qualities forall three tubes are almost the same. This implies that thefins have little or no effect on friction at low qualities.

The average pressure gradients for the helical micro-finand herringbone tubes are also much higher than that ofthe smooth tube. This is mainly caused by the increase infriction the fins bring about. These fins, however, also havethe effect to prolong the annular flow regime, delaying tran-sition to much lower vapour qualities. This is visually shownin Fig. 8 which depicts images throughout the condensationprocess for each of the tubes. To prove this point, at a qualityof 0.46, for the smooth tube the flow is intermittent, withslugs and plugs forming at the top of the tube. For the helicalmicro-fin and herringbone tubes, the flow is still annular witha liquid film layer around the circumference of the tube.

5.2. Penalty factors

The penalty factors were defined as the ratio of theaverage pressure gradients of the herringbone tube to thatof the smooth and helical micro-fin tubes, respectively. Thesepenalty factors are given in Fig. 9 for the smooth and helicalmicro-fin tubes. For R-22, the herringbone tube has on aver-age an 84% higher pressure drop than that of the smooth tube,while for R-134a and R-407C the pressure drops are 72% and80% higher, respectively. This increase is expected due to theincrease in turbulence generated by the fins.

Compared to the helical micro-fin tube, the herringbonetube has on average a 41% higher pressure drop for R-22.This penalty factor is consistent with those obtained by other


0 0.2 0.4 0.6 0.8 10

5

10

15

20

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

0 0.2 0.4 0.6 0.8 10

10

20

30

40

50

60

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

0 0.2 0.4 0.6 0.8 10

5

10

15

20

25

30

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

a b

c

G=400 kg/m2.s

G=600 kg/m2.s

G=800 kg/m2.s

Fig. 5. Pressure gradients for R-22 in the smooth tube (a), helical micro-fin tube (b), and the herringbone tube (c).

researchers [4e8]. At lower mass fluxes for R-134a andR-407C, however, the penalty factors are less than 1. An expla-nation could be that the flow over the fins of the herringbonetube induces a viscous sub-layer thickness, buffer layer, andan integral constant in the log-law that is greater than thatfor the helical micro-fin tube (as explained by Wang et al.[27]). This implies that the fins of the herringbone tube mighthave a drag-reducing effect. At higher mass fluxes, the pres-sure drops for these two refrigerants are, respectively, 24%and 17% higher than that of the helical micro-fin tube. Onaverage the herringbone tube has a pressure drop increase ofabout 27% above that of the helical micro-fin tube [28].

5.3. Pressure gradient correlation

A two-phase multiplier-type of pressure correlation wasdeveloped for the helical micro-fin and herringbone tubesof which the frictional pressure drop is given by

Dpf ¼ Dplof2L ð8Þ

with the two-phase multiplier, f, being that of Souza et al.[29], given by

f2L ¼ 1:376þ 7:242

X1:655tt

ð9Þ

The liquid-only pressure drop is the normal DarcyeWeisbach equation, modified by Friedel [30] for two-phaseflow:

Dplo ¼2floG2ð1� xÞ2L

rLDi

ð10Þ

The liquid-only friction factor, flo, is of the formf¼ 0.046 Re�0.2, and given in the modified form to incorpo-rate the effects of the fins of the herringbone tube as

flo ¼ 0:046Re�0:2L

�Di

Deq

��1�A

An

��0:5

ð2sec bÞ1:1 ð11Þ

with the area ratio, A/An, obtained from Azer and Said [31]and given in the modified form as

A

An

¼ 1� 2ent

pD2i cos b

ð12Þ

The friction factor in Eq. (11) is thus not only a functionof the fluid Reynolds number alone, but also the physicalgeometry of the tube. The second and third terms accountfor the geometric aspects of the tube while the last termaccounts for the swirl induced by fins, being a function of


0 0.2 0.4 0.6 0.8 10

5

10

15

20

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

0 0.2 0.4 0.6 0.8 10

10

20

30

40

50

60

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

0 0.2 0.4 0.6 0.8 10

10

20

30

40

50

60

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

a b

c

G=400 kg/m2.s

G=600 kg/m2.s

G=800 kg/m2.s

Fig. 6. Pressure gradients for R-134a in the smooth tube (a), helical micro-fin tube (b), and the herringbone tube (c).

the helix angle. The constants were obtained by curve-fittingthe experimental data. The experimental pressure drop datausing the modified correlation were predicted to within anerror of 1%, having an absolute mean deviation of 7%,and 94% of the data were predicted within 20%.

6. Heat transfer coefficients

The experimental set-up as used in the current study[1e3] was in excellent agreement with previous research per-formed on the same system for smooth, helical micro-fin andtwisted tape inserts viz Refs. [32] and [33]. Further, the re-sults obtained for this study were compared to thecorrelation of Shah [34] which predicted the R-22-data andR-134a-data on average to within 24%. For R-407C, the cor-relation of Tang [35] was used which predicted the data towithin 21%. For the helical micro-fin tube, the correlationof Cavallini et al. [8] was used. On average the correlationpredicted the data for all three refrigerants to within 11%.

6.1. Semi-local heat transfer coefficients

For all three test-condensers tested, the overall trend wasa decrease in heat transfer coefficient with a decrease in

vapour quality. Fig. 10 shows a plot of the semi-local heattransfer coefficients for the herringbone tube for condensingrefrigerants R-22, R-134a, and R-407C. It follows from theflow regime maps, Figs. 3 and 4, and Fig. 10 that as the qual-ity decreases (annular to intermittent flow) the heat transfercoefficients also decrease. This is due to the heat transferdegradation that takes place at lower qualities. For annularflow, a thin liquid layer is spread along the periphery ofthe tube. As the refrigerant starts to condense, the liquidfilm layer becomes thicker, with the resistance to heat trans-fer increasing. At very low qualities (and low mass fluxes),stratified/wavy flow (a thick liquid layer at the bottom ofthe tube with intermittent waves wetting the top of thetube) prevails. The main heat transfer in this type of flowis through heat conduction through the top part of the tubewhere the liquid layer is very thin.

Fig. 11 gives the data for the average heat transfer coef-ficients for the three different tubes and refrigerants tested.The overall trend is an increase in heat transfer coefficientwith an increase in mass flux, which is to be expected dueto the increase in turbulence, and hence the increase in mix-ing. When compared to the smooth tube, the herringbonetube had an average heat transfer coefficient increase of280% for R-22, 320% for R-134a, and 280% for R-407C.


0 0.2 0.4 0.6 0.8 10

5

10

15

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

0 0.2 0.4 0.6 0.8 10

5

10

15

20

25

30

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

0 0.2 0.4 0.6 0.8 10

10

20

30

40

Vapour Quality

Pre

ssur

e G

radi

ent,

kP

a m

−1

a b

c

G=400 kg/m2.s

G=600 kg/m2.s

G=800 kg/m2.s

Fig. 7. Pressure gradients for R-407C in the smooth tube (a), helical micro-fin tube (b), and the herringbone tube (c).

When compared to the helical micro-fin tube, the herring-bone tube had an average increase of approximately 170%for R-22, 190% for R-134a, and 200% for R-407C.

6.2. Heat transfer enhancement factors

The area enhancement ratio is defined as the ratio of thesurface area of the enhanced tube to that of an equivalentsmooth tube. The area enhancement factors for the helical mi-cro-fin and herringbone tube are 1.52 and 1.80, respectively.

The heat transfer enhancement factor is defined as the ra-tio of the average heat transfer coefficient of the enhancedtube to that of the smooth tube with consideration of thearea enhancement ratio. The average enhancement factorsof the herringbone tube against the smooth tube were foundto be 1.6, 1.9, and 1.6 for the refrigerants R-22, R-134a, andR-407C, respectively. The enhancement factors of theherringbone tube against the helical micro-fin tube wereon average 1.4, 1.5, and 1.4 for the three refrigerants, respec-tively. These enhancement factors are shown in Fig. 12.

The overall performance of the two enhanced tubes withrespect to the smooth tube may be defined as the ratio ofthe average enhancement factor to the average penalty factorfor each tube. The average performance factors for the

herringbone tube against the smooth tube are 1.0 for R-22,1.1 for R-134a and 0.9 for R-407C. This implies that thegain in heat transfer enhancement over the power consumedis approximately one to one. For the herringbone tubeagainst the helical micro-fin tube, the average performancefactors are 1.0 for R-22, 1.4 for R-134a, and 1.6 for R-407C. The performance factors are given in Fig. 13.

6.3. Heat transfer correlation

By adapting the correlation of Cavallini et al. [8], whichwas developed for use with helical micro-fin tubes, the heattransfer coefficients for herringbone tubes could be pre-dicted. The adapted correlation applies to heat transfer coef-ficients in the annular and intermittent flow regimes and isgiven by

a¼ lL

Di

0:05 Reseq Pr1=3

L TFðRxÞ2ðBo� lnðFrÞÞ�0:26 ð13Þ

where Reeq is the equivalent Reynolds number given by

Reeq ¼4 _mh1� xþ xðrL=rVÞ

0:5i

pDimL

ð14Þ


Fig. 8. The condensation process of R-134a at a mass flux of 500 kg m�2 s�1 for the three tubes tested.


The exponent to the equivalent Reynolds number isdefined for each tube as:

s¼ 0.715 for herringbone tubess¼ 0.700 for helical micro-fin tubes

TF is defined as a trend factor and is given by

TF¼ ð1� xÞ þ x1:3 ð15Þ

while Rx is given by

Rx ¼�

2enð1� sinðg=2ÞÞpDicosðg=2Þ þ 1

�ðcos bÞ�1 ð16Þ

300 400 500 600 7001

1.5

2

2.5

Pen

alty

Fac

tor

300 400 500 600 700 800 9000.5

1

1.5

2

Mass Flux, kg m−2 s−1

Pen

alty

Fac

tor

a

b

R−22R−134aR−407CR−22 Polyn.R−134a Polyn.R−407C Polyn.

Fig. 9. Penalty factors for the herringbone tube against the smooth

tube (a), and the herringbone tube against the helical micro-fin

tube (b).

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 10

5000

10000

15000

20000

25000

Vapour Quality

Hea

t T

rans

fer

Coe

ffic

ient

, W m

−2 K

−1

R−22, 400 kg m−2.s−1

R−134a, 400 kg m−2.s−1

R−407C, 400 kg m−2.s−1

R−22, 600 kg m−2.s−1

R−134a, 600 kg m−2.s−1

R−407C, 600 kg m−2.s−1

R−22, 800 kg m−2.s−1

R−134a, 800 kg m−2.s−1

R−407C, 800 kg m−2.s−1

Fig. 10. Semi-local heat transfer coefficients for the herringbone

tube using refrigerants R-22, R-134a, and R-407C for the mass

flux range of 400e800 kg m�2 s�1.

The Bond and Froude numbers are given by

Bo¼ grLepDi

8snð17Þ

Fr ¼ u2vo

gDi

ð18Þ

For the herringbone tube, Eq. (13) predicted the ex-perimental heat transfer coefficient data on averagewithin 4% with an average standard deviation of 8%,and 96% of the semi-local heat transfer coefficient data

300 400 500 600 700 800 9000

2000

4000

6000

8000

10000

12000

14000

16000


Ave

rage

Hea

t Tra

nsfe

r C

oeff

icie

nt, W

m−2

K−1

R−22 HerringboneR−134 HerringboneR−407C HerringboneR−22 Helical Micro−finR−134a Helical Micro−finR−407C Helical Micro−finR−22 SmoothR−134a SmoothR−407C Smooth

Herringbone

HelicalMicro−fin

Smooth

Fig. 11. Average heat transfer coefficients for refrigerants R-22,

R-134a, and R-407C.

300 400 500 600 700 800 9001

1.2

1.4

1.6

1.8

2

2.2

Enh

ance

men

t F

acto

r

300 400 500 600 700 800 9001

1.11.21.31.41.51.61.7


Enh

ance

men

t F

acto

r

a

b

R−22R−134aR−407C

Fig. 12. Heat transfer enhancement factors for the herringbone tube

against the smooth tube (a) and helical micro-fin tube (b).


fell within the 20% region. For the helical micro-fin tube,Eq. (13) predicts the data within 6% with an averagestandard deviation of 9%, and 89% of the data fell withinthe 20% region.

7. Conclusion

For the condensation experiments, three different tubeswere tested, namely a smooth tube, an 18� helical micro-fin tube, and a herringbone tube. These experiments wereperformed with refrigerants R-22, R-134a, and R-407C ata saturation temperature of 40 �C and at mass fluxes rangingfrom 400 to 800 kg m�2 s�1.

Flow regime observations showed that the transitionfrom annular flow to intermittent flow occurred at an averagequality of 0.49 for the smooth tube, 0.29 for the helicalmicro-fin tube, and 0.26 for the herringbone tube. Thesetransition qualities had an effect on the total pressure dropsfor the different tubes. For the herringbone tube against thesmooth tube, the pressure drops were about 80% higher,while for the herringbone tube against the helical micro-fin tube, the pressure drops were about 27% higher. Thiswas due to the effects the fins of the tubes have on the fluidflow, prolonging the annular flow regime. This had the effectof higher vapour velocities which in turn implied greaterfriction.

An existing helical micro-fin tube correlation was modi-fied to predict the pressure drops inside the helical andherringbone tubes for both the annular and intermittentflow regimes. The results indicated that this modified corre-lation predicted the data within an average error of 1% andan average mean deviation of 7%.

Heat transfer results showed that the herringbone tubehad an average enhancement factor of 1.7 for the three

300 400 500 600 700 800 9000.60.70.80.9

11.11.21.3

Per

form

ance

Fac

tor

300 400 500 600 700 800 9000.5

1

1.5

2

2.5


Per

form

ance

Fac

tor

a

b

R−22R−134aR−407C

Fig. 13. Performance factors of the herringbone tube versus the

smooth tube (a) and the helical micro-fin tube (b).

refrigerants against the smooth tube. For the herringbonetube against the helical micro-fin tube, the enhancementfactor was in the order of 1.4. This increase in heat transferwas due to the fins increasing the turbulence as well as pro-longing the annular flow regime to much lower qualitieswhen compared to the smooth tube.

A heat transfer correlation from the literature was modi-fied for the herringbone tube. Results showed that the newcorrelation predicted the data on average to within 4%with a standard deviation of 8%. The same correlation wasused to predict heat transfer coefficients inside helical mi-cro-fin tubes. The correlation predicted the data on averageto within 6% with a standard deviation of 9%.

Acknowledgements

Wolverine Inc., Alabama, USA supplied the experimen-tal helical micro-fin tubes, and Mr. Petur Thors is thankedfor arranging the donation. Dr. Axel Kriegsmann of WielandWerke also kindly donated several lengths of micro-fintubing. The experimental herringbone tube was donated byProf. Yasuyuki Takata of Kyushu University, Japan, forwhich the authors are most grateful. The research workwas performed with a South African National ResearchFoundation grant, under grant number 2053278 and undera THRIP grant, number 3257.

References

[1] L. Liebenberg, A.E. Bergles, J.P. Meyer, A review of refriger-

ant condensation in horizontal micro-fin tubes, in: Proceed-

ings of the ASME Advanced Energy Systems Division,

2001, pp. 155e168.

[2] L. Liebenberg, J.R. Thome, J.P. Meyer, Flow visualization

and flow pattern identification with power spectral density

distributions of pressure traces during refrigerant condensa-

tion in smooth and micro-fin tubes, Journal of Heat Transfer

127 (2005) 209e220.

[3] J.A. Olivier, L. Liebenberg, M.A. Kedzierski, J.P. Meyer,

Pressure drop during refrigerant condensation inside horizon-

tal smooth, helical microfin, and herringbone microfin tubes,

Journal of Heat Transfer 126 (2004) 687e696.

[4] A. Miyara, Y. Otsubo, S. Ohtsuka, Y. Mizuta, Effects of fin

shape on condensation in herringbone microfin tubes, Interna-

tional Journal of Refrigeration 26 (2003) 417e424.

[5] T. Ebisu, K. Torikoshi, Experimental Study on evaporation

and condensation heat transfer enhancement for R-407C

using herringbone heat transfer tube, ASHRAE Transactions

104 (2) (1998) 1044e1052.

[6] A. Miyara, K. Nonaka, M. Taniguchi, Condensation heat

transfer and flow pattern inside a herringbone-type micro-fin

tube, International Journal of Refrigeration 23 (2000)

141e152.

[7] N. Goto, N. Inoue, N. Ishiwatari, Condensation and evaporation

heat transfer of R410A inside internally grooved horizontal

tubes, International Journal of Refrigeration 24 (2000) 628e638.

[8] A. Cavallini, D. Del Col, L. Doretti, G.A. Longo, L. Rossetto,

Heat transfer and pressure drop during condensation of


refrigerants inside horizontal enhanced tubes, International

Journal of Refrigeration 23 (2000) 4e25.

[9] N. Kattan, J.R. Thome, D. Favrat, Measurement and Prediction

of two-phase flow patterns for new refrigerants inside horizontal

tubes, ASHRAE Transactions 101 (2) (1995) 1251e1257.

[10] N. Kattan, J.R. Thome, D. Favrat, Boiling of R-134a and

R-123 in a microfin tube, in: Proceedings of the 19th Interna-

tional Congress of Refrigeration, The Hague, Netherlands,

1995, pp. 337e344.

[11] N. Kattan, J.R. Thome, D. Favrat, Flow Boiling in horizontal

tubes. Part 1: Development of a diabatic two-phase flow pat-

tern map, Journal of Heat Transfer 120 (1) (1998) 140e147.

[12] N. Kattan, J.R. Thome, D. Favrat, Flow boiling in horizontal tubes.

Part 3: development of a new heat transfer model based on flow

patterns, Journal of Heat Transfer 120 (1) (1998) 156e165.

[13] D.W. Shao, E. Granryd, Flow pattern, heat transfer and

pressure drop in flow condensation. Part I: pure and azeo-

tropic refrigerants, International Journal of HVAC & R

Research 6 (2) (2000) 175e195.

[14] A. Cavallini, G. Censi, D. Del Col, L. Doretti, G.A. Longo,

L. Rossetto, In-tube condensation of halogenated refrigerants,

ASHRAE Transactions 108 (1) (2002) 146e161.

[15] J. El Hajal, J.R. Thome, A. Cavallini, Condensation in horizon-

tal tubes, Part 1: Two-phase flow pattern map, International

Journal of Heat and Mass Transfer 46 (2003) 3349e3363.

[16] J.G. Collier, J.R. Thome, Convective Boiling and Condensa-

tion, Clarendon Press, Oxford, 1994.

[17] J.R. Thome, Update on KattaneThomeeFavrat flow boiling

model and flow pattern map, in: Proceedings of the Fifth

International Conference on Boiling Heat Transfer, Jamaica,

4e8 May 2003.

[18] D. Steiner, Heat transfer to boiling saturated liquids, in: Ver-

ein Deutscher Ingenieure (Ed.), VDI-Warmeatlas (VDI Heat

Atlas), VDI-Gesellschaft Verfahrenstechnik und Chemieinge-

nieurswesen (GCV), Dusseldorf, 1993.

[19] Y. Taitel, A.E. Dukler, A model for predicting flow regime

transitions in horizontal and near-horizontal gaseliquid

flow, American Institute of Chemical Engineering (AIChE)

Journal 22 (1) (1976) 47e55.

[20] S.Z. Rouhani, E. Axelsson, Calculation of volume void frac-

tion in the subcooled and quality region, International Journal

of Heat and Mass Transfer 13 (1970) 383e393.

[21] O. Zurcher, J.R. Thome, D. Favrat, Evaporation of ammo-

nia in a smooth horizontal tube: heat transfer measurements

and predictions, Journal of Heat Transfer 121 (1993)

89e101.

[22] K. Cho, S.-J. Tae, Condensation heat transfer for R-22 and

R-407C refrigeranteoil mixtures in a microfin tube with

a U-bend, International Journal of Heat and Mass Transfer

44 (2001) 2043e2051.

[23] S.J. Kline, F.A. McClintock, Describing uncertainties in single-

sample experiments, Mechanical Engineering 75 (1953) 3e8.

[24] REFPROP, NIST thermodynamic properties of refrigerants

and refrigerant mixtures (REFPROP), Version 6.0, NIST

Standard Reference Database 23, National Institute of

Standards and Technology, Gaithersburg, MD, 1998.

[25] D.E. Briggs, E.H. Young, Modified Wilson plot techniques for

obtaining heat transfer correlation for shell and tube heat

exchangers, Chemical Engineering Progress Symposium 65

(1969) 34e35.

[26] A. Cavallini, Experimental data, Private communication,

2004.

[27] J. Wang, S. Lan, G. Chen, Experimental study on the turbu-

lent boundary layer flow over riblets surface, Fluid Dynamics

Research 27 (4) (2000) 217e229.

[28] J.A. Olivier, L. Liebenberg, J.P. Meyer. Experimental pressure

drops during condensation inside smooth, helical micro-fin

and herringbone tubes using R-22, R-134a, and R-407C, in:

Proceedings of the Sixth World Conference on Experimental

Heat Transfer, Fluid Mechanics, and Thermodynamics,

Matsushima, Miyagi, Japan, April 17e21.

[29] A.L. Souza, J.C. Chato, J.P. Wattelet, B.R. Christoffersen,

Pressure drop during two-phase flow of pure refrigerants

and refrigerant-oil mixtures in horizontal smooth tubes,

ASME Heat Transfer With Alternate Refrigerants, HTD,

Atlanta, Georgia 243 (1993) 35e41.

[30] L. Friedel, Improved friction pressure drop correlation for hor-

izontal and vertical two-phase two-component flow in pipes,

in: European Two-Phase Flow Group Meeting, Ispra, 1979.

[31] N.Z. Azer, S.A. Said, Augmentation of condensation heat

transfer by internally finned tubes and twisted tape inserts,

in: Proceedings of the Seventh International Heat Transfer

Conference, Munchen, West-Germany.

[32] F.J. Smit, J.P. Meyer, R-22 and zeotropic R-22/R-142b mix-

ture condensation in microfin, high-fin and twisted tape insert

tubes, Journal of Heat Transfer 124 (5) (2002) 912e921.

[33] F.J. Smit, J.R. Thome, J.P. Meyer, Heat transfer coefficients

during condensation of the zeotropic refrigerant mixture

HCFC-22/HCFC-142b, Journal of Heat Transfer 124 (6)

(2002) 1137e1146.

[34] M.M. Shah, A general correlation for heat transfer during film

condensation inside pipes, International Journal of Heat and

Mass Transfer 22 (1979) 446e547.

[35] L. Tang, Empirical study of new refrigerant flow condensation

inside horizontal smooth and micro-fin tubes, PhD thesis,

University of Maryland at College Park, 1999.

Heat transfer, pressure drop, and flow pattern recognition during condensation inside smooth,...

Documents

Transcript of Heat transfer, pressure drop, and flow pattern recognition during condensation inside smooth,...