Heat transfer engineering an international journal, tập 31, số 10, 2010

At low film Reynolds numbers,the top tubes of the array showed a large peak in the measuredheat transfer coefficients most probably, this is an impingementeffect due to the surface geome

Copyright
C Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903547461

Film Condensation of R-134a and

R-236fa, Part 1: Experimental

Results and Predictive Correlation

for Single-Row Condensation on

Enhanced Tubes

MARCEL CHRISTIANS, MATHIEU HABERT, and JOHN R THOME

Laboratory of Heat and Mass Transfer (LTCM), Faculty of Engineering Science Ecole Polytechnique F´ed´erale de Lausanne

(EPFL), Lausanne, Switzerland

New predictive methods for falling film condensation on vertical arrays of horizontal tubes using different refrigerants are

proposed, based on visual observations revealing that condensate is slung off the array of tubes sideways and significantly

affects condensate inundation and thus the heat transfer process For two types of three-dimensional enhanced tubes,

advanced versions of the Wolverine Turbo-C and Wieland Gewa-C tubes, the local heat flux is correlated as a function of

condensation temperature difference, the film Reynolds number, the tube spacing, and liquid slinging effect The proposed

methods work best when using R-134a, as these tubes were designed with this refrigerant in mind.

INTRODUCTION

Tubes in shell-and-tube condensers, widely used in

refrig-eration, heat pumps, and chemical process industries, are

sub-jected to condensate inundation from the neighboring upper

tubes In order to increase the efficiency of these systems, plain

tubes were replaced by all types of enhanced tubes, from finned

tubes to tubes with advanced two-dimensional (2D) and

three-dimensional (3D) enhancement geometries However, it is

nec-essary to characterize the performance of new tubes, so that

design engineers have a solid foundation on which to base their

designs Furthermore, it is of interest to test the performance

of these tubes with several refrigerants, such that the differing

behavior may be quantified and taken into account during the

design stage itself

The authors thank the laboratory’s industrial sponsors Johnson Controls,

Trane, Wieland Werke, and Wolverine Tube, Inc., for funding this study Special

thanks to the tube manufacturers, Wieland Werke and Wolverine Tube, Inc., for

supplying the tubes utilized.

Address correspondence to Prof John R Thome, Laboratory of Heat and

Mass Transfer (LTCM), Faculty of Engineering Science, Ecole Polytechnique

F´ed´erale de Lausanne (EPFL), Station 9, Lausanne CH-1015, Switzerland.

E-mail: john.thome@epfl.ch

PREVIOUS HEAT TRANSFER COEFFICIENT STUDIES

Jung et al [1–3] performed falling film condensation testsusing plain, low-fin and enhanced tubes and pure refriger-ants R-11, R-12, R-123, R-22, and R-134a, and zeotropicand azeotropic refrigerant mixtures R-407C, R-410A, R-32/R-134a, and R-134a/R-123 on a test section comprised of asingle tube at a saturation temperature of 39◦C The finnedtubes had 1,024 fins per meter, while the enhanced tube testedwas the Turbo-C Chang et al [4] performed tests on sin-gle tubes connected by a U-bend, at a saturation tempera-ture of 39◦C on low-fin and 3D enhanced tubes, using re-frigerant R-134a The finned tubes had 1,024 and 1,574 finsper meter, while the two 3D enhanced tubes had T- and Y-shape fins Kumar et al [5, 6] tested plain and finned tubeswith refrigerant R-134a on single tubes, at a saturation tem-perature of 39.3◦C The finned tubes had fin densities of 472(rectangular), 934, 1,250, 1,560, and 1,875 fins per meter.Sreepathi et al [7] tested several proprietary finned tubes,commercial finned tubes (748 and 1,574 fins per meter), andthe enhanced tubes Thermoexcel-C and Thermoexcel-CC1in asingle tube configuration, using R-11 and R-123, at saturationtemperatures of 23.5 and 27.4◦C Wen et al [8] studied the per-formance of four tubes (667 and 1000 fpm, with and without799

filled fin-roots) in a single tube test section using R-113 at a

saturation temperature of 47.6◦C

Kang et al [9] tested low-fin and 3D enhanced tubes in a test

section consisting of five horizontal tubes placed on a single

horizontal plane (i.e., side by side), at a saturation temperature

of 60◦C, using refrigerant R-134a The tested tubes included

one low-fin tube and three Turbo-C variants Gst¨ohl and Thome

[10, 11] performed tests on a single column of several tubes

(varying the pitch between tubes), as well as plain, low-fin, and

3D enhanced tubes at a saturation temperature of 31◦C In these

tests, it was possible to vary the overfeed onto the first tube

of the column to simulate flow deeper in a bundle The tubes

tested were a Turbo-Chil low-fin tube, and both Wolverine and

Wieland enhanced condensation tubes (Turbo-CSL and

Gewa-C) using only R-134a As a continuation of this work, Habert

et al [12] presented additional flow regime transition criteria

for Wieland and Wolverine enhanced tubes using an additional

refrigerant (R-236fa) However, in this study, no heat transfer

measurements were presented

As such, the aim of this article is to present and discuss the

results obtained in the LTCM’s falling film facility for advanced

versions of the Turbo C and Gewa C 3D enhanced tubes, using

both R-134a and R-236fa R-236fa was chosen as a second

test fluid because of its compatibility with the experimental test

stand In addition to the preceding, prediction methods based on

Gst¨ohl and Thome’s [11] original R-134a data-only predictive

model are developed and presented

EXPERIMENTAL FACILITY

The experimental setup is comprised of three circuits,

namely, the refrigerant, water–glycol, and water circuits The

refrigerant circuit is shown schematically in Figure 1 It

com-prises an electrically heated evaporator (Figure 1, (1) flooded

evaporator) to maintain the desired saturation condition, a

con-denser (Figure 1, (5) auxiliary overhead concon-denser) to condense

Figure 1 Schematic of the refrigerant circuit in the Falling Film Facility.

any vapor not condensed in the test section, and the test sectionitself (Figure 1, (4) test section)

In the refrigerant circuit, starting from the flooded rator (Figure 1, (1) flooded evaporator), the refrigerant flowsthrough the filter (not shown) and the subcooler (Figure 1, (6)liquid subcooler) to the gear pump (self-lubricating without oil:Figure 1, (7) overfeed pump) Parallel to the pump, bypass pip-ing is installed so that, together with a frequency controller onthe pump, the desired liquid flow rate can be accurately set ACoriolis mass flow meter (Figure 1, (8) Coriolis mass flow me-ter) follows, after which an electric heater (Figure 1, (9) liquidheater) is installed to bring the liquid close to saturation con-ditions at the test section inlet At this point, the liquid entersthe test section and is distributed uniformly on the top row ofthe heated tubes Special care has been taken in the distributordesign in order to achieve uniform liquid distribution on the toptube Once the liquid leaves the distributor, it falls onto the top

evapo-of the cooled tube array, on which the vapor in the test tion is partially condensed; the residual liquid leaves the testsection by gravity From the exit of the test section, the liquidflows back to the flooded evaporator by the effect of gravity.The vapor that runs through the test section is generated in theflooded evaporator, where by natural convection it rises to thetop of the test facility It flows in to the top of the test section,where the vapor flow is uniformly distributed over its length,and any remaining vapor is sucked out at the bottom of the testsection After exiting the test section from the bottom, it flowsback into the condenser, and the liquid drops by gravity back tothe flooded evaporator The amount of vapor flow can be con-trolled by increasing the heat input in the flooded evaporator,which in turn generates more vapor Consequently, to maintain

sec-a constsec-ant system pressure, the cooling losec-ad on the sec-auxilisec-arycondenser is greater In these tests, it was attempted to main-tain the vapor velocity as low as possible, such that vapor sheareffects were minimized

The water circuit (not shown) is responsible for the coolingeffect in the test section The water is driven through the testtubes by a centrifugal pump An electronic speed-controller,together with a bypass line, ensures good precision in any watermass flow adjustment The water flows through two liquid–liquid heat exchangers; the first is cooled with industrial watersourced from Lake Geneva at a constant temperature of 7◦C,while the second is heated with hot water from a closed-loopcircuit heated by a heat pump This water has its flow rate set by

a computer-controlled valve The water temperature at the testsection inlet is thus automatically maintained constant The totalwater mass flow rate is measured with a Coriolis flow meter (notshown) Before entering the test section, the test-line water flow

is split into three subcircuits, each supplying to two tubes in thetest section Each subcircuit has two tube passes; i.e., water goes

in a copper tube in one direction (left to right) and comes backthrough the copper tube just above in the opposite direction Awater–glycol mixture from a network installation is used as acold source for the auxiliary condenser

heat transfer engineering vol 31 no 10 2010

M CHRISTIANS ET AL 801The test section is a rectangular stainless-steel vessel with six

large windows situated at the front and rear in order to have full

visual access into the experimental setup, to observe the flow on

the tubes The copper test tubes had a nominal outer diameter of

18.38 mm and are arranged horizontally in a vertical array The

length of the tubes was 554 mm In total, six tubes (i.e., three

subcircuits) were installed, at a industry-standard pitch of 38.5

mm

Furthermore, a stainless-steel tube with an external diameter

of 8 mm was inserted inside each copper test tube Pairs of

thermocouples were located at three positions axially along the

tube, protruding out through holes to measure the temperature

of the water in the annulus between the stainless steel tube and

the copper tubes At every location, one thermocouple is facing

upward and one is facing downward A copper wire with a

rectangular cross section wound helically around the

stainless-steel tube promoted mixing, and further increased the water-side

heat transfer coefficient

Pressure transducers connected to the test section above and

below the array of tubes were used to measure the vapor pressure

in the test section The vapor temperature in the test section was

measured above and below the tube array using sheathed

ther-mocouples The temperatures of the liquid entering and leaving

the test section, as well as the vapor leaving the test section,

were measured

EXPERIMENTAL ERRORS AND PROCEDURES

The internally mounted thermocouples measuring the water

temperature within the tube annulus along the axial length of

the tubes provide the water temperature profile as a function of

the distance x along the tubes Assuming only heat flow in the

radial direction, the local heat flux on the outside of the tube,

q o, may thus be expressed as

q o= m˙water c p,water

π D o

dT water dx



(1)

where D o is the outside diameter The value (dT water /dx) is

ob-tained by differentiating a second-order polynomial fit of the

water temperature profile Nearly identical temperatures for the

pairs of thermocouples located at each location indicate good

mixing of the water (the temperatures were within

thermocou-ple uncertainty), which helps increase the accuracy of the data

reduction method

To determine the external local heat transfer coefficient, h o,

between the outside surface of the copper tubes and the

refrig-erant, a modified Wilson plot procedure using nucleate pool

boiling (as in Robinson and Thome [13]) on the outside of

the tubes was implemented The modified Wilson plot method

takes into account slight variations in the heat flux by

assum-ing a relation for the external heat transfer coefficient given by

h o = C o q o 0.7 The internal heat transfer coefficient is the one

given by the Gnielinski [14] correlation, h gni, multiplied by a

constant C ithat takes into account the increase in heat transfer

Table 1 Calculated values for the internal heat transfer multiplier C i

due to any internal enhancement, the reduced flow area, and creased turbulence due to the inserted helical tape The Wilsonplot expression for the tubes is thus

a plot of the values in the brackets on the left versus the values in

the brackets on the right gives the value of C i, while the inverse

of the abscissa intercept yields C o Thus, the heat transfer on theoutside of the tube at any location along its axis can be calculated

with the value of C i, along with the measured water temperatureprofile, the water mass flow rate, and the saturation temperature

of the refrigerant However, in this study the local coefficient

is only evaluated at the midpoint of every tube This calculatedvalue is a perimeter-averaged heat transfer coefficient based onthe external tube diameter The modified Wilson tests were con-ducted over a water-side Reynolds number range varying from

6,000 to 16,000 Table 1 shows the values of C iobtained by this

study It can be seen that the C ivalue obtained for the WolverineTurbo-C enhanced condensing tube of 7.38 is higher than forthe other tube, due to its 3D internal enhancement structure

To eliminate all traces of non-condensable gases that mighthave been introduced into the facility (i.e., during tube or re-frigerant changes), a vacuum pump (not shown in Figure 1) isconnected to the system and is run until the two low-pressurereference pressure transducers show no more than 100 Pa (ab-solute) Once the vacuum pump is stopped, the system pressure

is monitored to make sure that no leaks are present Only oncethese two steps have been accomplished is the system refilledwith refrigerant to proceed with testing Any remaining traces ofnon-condensable gases in the system will migrate to the over-head condenser, where they remain The measured saturationtemperature using thermocouples and that obtained from thepressure sensors and REFPROP v8 [15] differed by 0.1 K, avalue within the uncertainty of both measurements

For experiments involving overfeed, the film flow rate of theliquid arriving on the first tube was evaluated from the measuredmass flow rate and the tube length, assuming that the refrigerant

is at saturation conditions The mass flow of refrigerant densing on the first tube is calculated by an energy balance ondifferential elements and added to the film flow rate arriving onthe first tube to obtain the film flow rate at the top of the secondtube and so on This means an ideal one-dimensional down-ward flow is assumed on the tube rows and assumes that all thecondensate flows from one tube to the next without leaving theheat transfer engineering vol 31 no 10 2010

con-Table 2 Uncertainties of measured heat transfer coefficients at the three

heat flux conditions tested

δho/ ho Tube qo = 20 kW/m 2 qo = 40 kW/m 2 qo = 60 kW/m 2

(condensing)

(condensing)

tube row In case of no overfeed, a similar procedure was

ap-plied, with the initial flow rate onto the top tube set to 0 The

two-pass water design gives a nearly uniform axial condensate

distribution along the tube array after each pair of tubes The

saturation temperatures, as well as the transport and

thermody-namic properties, are calculated according to REFPROP v8 [15]

from the mean of the pressures measured by pressure

transduc-ers above and below the tube array

Tests were conducted by gradually decreasing the liquid film

flow rate on the top tube at a fixed heat flux The data were

logged only if steady-state conditions were attained An error

analysis was performed, and the mean relative errors in the local

heat transfer coefficient at a saturation temperature of 31◦C are

tabulated in Table 2 A more detailed description of the test

facility, data reduction methods, and measurements accuracies

can be found in Gst¨ohl and Thome [10, 11]

EXPERIMENTAL RESULTS WITH THE SINGLE-ROW

TEST SECTION

Tests were performed using the Wolverine Turbo and

Wieland Gewa condensing tubes (both of them have an

18.38-mm nominal outer diameter) provided by the manufacturing

companies Before installation into the test section, the tubes

were thoroughly cleaned In the column of six tubes (single

vertical row), the center-to-center tube pitch was 38.5 mm, and

tests were performed using refrigerants R-134a and R-236fa, at

a saturation temperature of 31◦C Furthermore, tests were

per-formed at constant tube array nominal heat fluxes of 20, 40, and

60 kW/m2

In Figures 2–7, it can be seen that the refrigerant in use has

a very large effect on the performance of each tube For these

tubes, and at all heat fluxes, the R-236fa results show lower

performance over the entire Reynolds number range

Further-more, when using R-134a, the heat transfer performance of the

first (top) tube is considerably higher than the rest of the array,

something especially true at lower Reynolds numbers This is

probably related in some manner to the overfeed from the liquid

distributor—the value at the lowest Reynolds number in each

diagram for tube 1 (that is, without overfeed) usually aligns well

with the trend of the rest of the data

For tests at a constant nominal array heat flux, it can be seen

that there is a very slight or almost no dependence on the tube

row number, a trend that was also evident in the testing presented

by Gst¨ohl and Thome [10]

0 500 1000 1500 2000 2500 3000 3500 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

R−134a

R−236fa

Figure 2 Heat transfer performance of the six Wolverine Turbo C condensing test tubes at a nominal array heat flux of 20 kW/m 2 using both R-134a and R-236fa.

With the Turbo condensing tube/R-134a combination(Figures 2–4), the behavior of the tube is similar to the three-dimensional enhanced tubes tested originally by Gst¨ohl andThome [10] This similarity is not in terms of the heat trans-fer coefficient values themselves, since the Turbo-CSL re-sults presented [11] had peaks of roughly 25 kW/m2-K whilethis tube’s peak is at 28 kW/m2-K, but rather in the gen-eral form of the evolution of the heat transfer with increasingReynolds number Also using R-134a, the Wieland Gewa data(Figures 5–7) show that the top two tubes have a large heattransfer peak at lower Reynolds numbers With both tubes, thedata at the highest Reynolds numbers fluctuate and still seem toform a plateau like that seen in the Turbo-CSL results [11] In

0 500 1000 1500 2000 2500 3000 3500 4000 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

R−236fa R−134a

Figure 3 Heat transfer performance of the six Wolverine Turbo C condensing test tubes at a nominal array heat flux of 40 kW/m 2 using both R-134a and R-236fa.

heat transfer engineering vol 31 no 10 2010

Figure 4 Heat transfer performance of the six Wolverine Turbo C condensing

test tubes at a nominal array heat flux of 60 kW/m 2 using both 134a and

R-236fa.

contrast to the results of Gst¨ohl and Thome [10, 11], the heat

transfer degradation with increasing Reynolds number is not

as severe; while it does occur at essentially the same Reynolds

number, and with the same slope, the heat transfer coefficient

stabilizes at ∼50 to 60% of the peak measured heat transfer

coefficient, while for the tubes tested by Gst¨ohl and Thome,

the plateau was found at around 20% of the peak heat transfer

coefficient value Evidently, this will have a beneficial effect on

condenser performance For the Wieland tube, tubes 2 through

6 are closely grouped The general trend for the tubes in the

array is an increase to a stable plateau Furthermore, tubes 1 and

2 are the only ones to show significant heat transfer degradation

as the film velocity increases This could be due to a type of

R−134a

R−236fa

Figure 5 Heat transfer performance of the six Wieland Gewa C condensing

test tubes at a nominal array heat flux of 20 kW/m 2 using both R-134a and

R-236fa.

0 500 1000 1500 2000 2500 3000 3500 4000 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

R−236fa

Figure 6 Heat transfer performance of the six Wieland Gewa C condensing test tubes at a nominal array heat flux of 40 kW/m 2 using both R-134a and R-236fa.

trance effect (impingement) only apparent due to the surface’sgeometry

Using R-236fa, Figures 2–4 show that for the Wolverinecondensing tube, the behavior of the heat transfer coefficient isvastly different In this case, the heat transfer coefficient slowlyincreases to a band within which the heat transfer fluctuates yetremains bound As neither the type of tube, nor the geometricdistribution, nor the measurement technique was changed, it can

be safely concluded that the difference in heat transfer evolutionand the degradation of performance with respect to the R-134atests is solely a function of the thermophysical properties of therefrigerant under consideration Looking at the Wieland tube

0 5000 10000 15000 20000 25000

Film Reynolds number, Rebottom [ − ]

2 K]

Array, Gewa C tube, tube spacing 38.5mm, heat flux: 60kW/m2

Tube 1 Tube 2 Tube 3 Tube 4 Tube 5 Tube 6

R−134a

R−236fa

Figure 7 Heat transfer performance of the six Wieland Gewa C condensing test tubes at a nominal array heat flux of 60 kW/m 2 using both R-134a and R-236fa.

heat transfer engineering vol 31 no 10 2010

Table 3 Comparison of the physical properties of the two refrigerants at

data (Figures 5–7), these results differ from the R-134a data

in that they are contained within a small band, and lower in

magnitude Furthermore, the maximum tube 1 peak using

R-134a was around 23 kW/m2-K, while with R-236fa this peak

was found at 12.5 kW/m2-K

The large difference in absolute performance (and with

re-spect to increasing Reynolds number) can be attributed to the

geometric design of the tubes themselves; both of these were

optimized for R-134a condensate drainage, and using a

dif-ferent refrigerant is going to have an impact on performance

Table 3 shows a comparison of the physical properties of the

two refrigerants at 31◦C In both falling film evaporation and

condensation, two thermodynamic properties that have large

in-fluence are the liquid viscosity and the surface tension It can

be seen that there is a 36% difference in viscosity and 25%

difference in surface tension between the two refrigerants at a

saturation temperature of 31◦C This will primarily affect the

thickness of the liquid film and its interaction with the tube

UPDATED PREDICTION METHOD

Background

Gst¨ohl and Thome [11] presented two heat transfer models

for 3D enhanced condensing tubes: the first for when there is

no slinging (of condensate off the side of the tube), while the

second one takes into account the reduction of the Reynolds

number due to the slinging They first correlated the heat flux to

the Reynolds number on top of the tube by

q o = (a + cRe top ) T b (3)

where the coefficients a, b, and c for the tubes that were tested

are given in Table 1 in [11] However, it was found that for 3D

enhanced tubes, as the Reynolds number increased, a fraction of

the liquid refrigerant left the tube array sideways [16] This was

due to the fact that the liquid film did not fall as a stable sheet,

but rather fell with an oscillatory motion Thus, they calculated

the critical angle (a function of the tube geometry and tube pitch)

for which the liquid film would begin to not reattach the tube

where r o is the tube radius and p is the tube pitch Then, the

slinging angle is defined as a linear function of the Reynoldsnumber

The portion of liquid that leaves the tube is assumed to beproportional to the ratio of (θ – θcrit )/θ This means that the film Reynolds number on the top of the nth tube in the array can

To apply, the calculation is started on the top tube of the array

As long as there is no slinging (i.e., θ≤ θcrit), Eq (3) is used

to determine the heat transferred by the tube, and the amount ofliquid leaving the bottom of the tube can be calculated In thiscase, all the liquid flowing off the bottom of the tube is assumed

to fall on top of the tube below (Retop,n = Rebottom,n−1) As

soon as the liquid starts to sling out (i.e., when θ > θ crit), Eq.(6) can be used to determine the amount of liquid that arrives onthe tube below Equation (7) is used to determine the heat fluxtransferred by the tube To determine the heat transfer coefficientfrom the preceding equation, it suffices to divide the heat flux

by the temperature difference, that is,

corre-Updated Model

The preceding method is fluid/enhanced tube specific, andhence, to update its validity for the new tubes, it is evident thatthe coefficients utilized should be modified to better fit the newdata This is also required, since no general model accountingfor the enhancement geometry and its dimensions is availablefor these fluid/enhanced tubes combinations in the literature

A nonlinear least-squares optimization method was utilized tominimize the difference between the prediction method andthe measured heat transfer data The optimization process wasstarted from multiple initial positions (spread from the upper toheat transfer engineering vol 31 no 10 2010

the lower bounds of the parameter constraints), and all arrived

either at the presented solution or very close, showing that the

minimum found is a global minimum rather a local minimum

The coefficients for use in Eqs (7) and (8) are shown in Table 4

There are four sets of coefficients, one for each tube/refrigerant

combination tested

Figure 8 shows a comparison of the prediction method found

using the nonlinear least-squares optimization and the measured

heat transfer coefficient data obtained using the Turbo enhanced

condensing tube and R-134a This method predicts 87% of the

results within an error range of±15%, while 100% of the data

are within a±30% error band Comparing the obtained

coeffi-cients to those found by Gst¨ohl and Thome (Table 2 of [11]), it

is found that the resulting coefficients are similar in magnitude

(a = ∼25,000, b = ∼0.8, c = ∼−6.5, d = ∼0.0004, e = 0).

Continuing the analysis of the results obtained with the Turbo

enhanced tube (now using R-236fa), the same optimization

al-gorithm was implemented (using Gst¨ohl and Thome’s model),

even though the data do not show a pronounced degradation

in heat transfer The prediction method (using the coefficients

shown in Table 4), plotted on the same figure as the results, is

shown in Figure 9 For R-236fa, this method only predicts 70%

of the data within±15%, and 95% of the data to within 30%

Film Reynolds number, Rebottom,n−1 [ − ]

2 K]

R−134a, Turbo, tube spacing 38.50mm

Tube 1 Tube 3 Tube 5 Model

Figure 8 Prediction method for the single-row Turbo condensing tube data

using R-134a.

However, the R-236fa data are, for most of the tube/refrigerantconfigurations, relatively constant, showing little influence withrespect to Reynolds number The optimization algorithm shiftedthe onset of the plateau region to a smaller Reynolds number

by first suppressing the slinging angle (θ) such that it has

al-most no effect It also flattened the prediction by setting a

y-intercept 50% lower than has been previously calculated (forR-134a and the different tubes tested), and slightly decreas-

ing the power of the exponent b that affects the temperature difference T Furthermore, for R-236fa, the multiplier c acts

to suppress the influence of both the slinging angle and theReynolds number, rather than to amplify it as seen in the R-134aresults

Applying the method to the Wieland Gewa C enhanced densing tube and test refrigerant R-134a results in the predictionshown in Figure 10 The method predicts 90% of the resultswithin an error range of ±15%, while 100% of the data arewithin a±30% error band Comparing the empirical coefficients

con-to those found by Gst¨ohl and Thome for the Gewa-C, it is found

that the resulting (a) y-intercept coefficient and (b) temperature difference exponent are similar in magnitude (a = ∼20,000,

b = ∼0.9, c = ∼−0.6, d = ∼0, e = ∼0) However, there is

a relatively large change for the Reynolds number multiplier c,

0 5000 10000 15000 20000 25000

Film Reynolds number, Rebottom,n−1 [ − ]

2 K]

R−236fa, Turbo, tube spacing 38.50mm

Tube 1 Tube 3 Tube 5 Model

Figure 9 Prediction method for the single-row Turbo condensing tube data using R-236fa.

heat transfer engineering vol 31 no 10 2010

Figure 10 Prediction method for the single-row Gewa condensing tube data

using R-134a.

which in this case acts to suppress the the influence of both the

slinging angle and the Reynolds number, rather than amplify it

(that is, this tube slings less) The optimization algorithm shifted

the onset of the plateau region to a much smaller Reynolds

number by suppressing the slinging angle (θ) such that it has

almost no effect

The prediction method for the Gewa condensing

tube/R-236fa configuration (using the coefficients shown in Table 4)

is plotted on the same figure as the results in Figure 11 For

R-236fa, this method predicts 70% of the data within±15%

and 90% of the data to within 30% As with the R-134a data,

the optimization algorithm shifted the onset of the plateau

re-gion to a smaller Reynolds number The method utilized to

Figure 11 Prediction method for the single-row Gewa condensing tube data

using R-236fa.

perform this is as explained for the R-134a results The c

mul-tiplier acts to slightly amplify the Reynolds number effect, aswas the case with the previous results obtained by Gst¨ohl andThome

Presently, it is not possible to present one set of constants a–

ethat works for all the fluid/enhanced tubes combinations Toachieve this, one needs first to develop a theory-based 3D con-densation model, and then a predictive-based slinging model;such a model requires local film flow measurements and is agood topic of research for the future

CONCLUSIONS

The heat transfer performance of the new versions of theWolverine Turbo C and Wieland Gewa C condensing tubes, us-ing refrigerants R-134a and R-236fa, has been measured UsingR-134a, the heat transfer coefficient of the two enhanced tubesvaried as a function of the film Reynolds number, and was char-acterized by two distinct zones At low film Reynolds numbers,the top tubes of the array showed a large peak in the measuredheat transfer coefficients (most probably, this is an impingementeffect due to the surface geometry), after which the heat trans-fer coefficient decreased almost linearly Above a certain filmReynolds number, the heat transfer coefficient decreases muchmore slowly and achieves an almost constant value (that is,reaches a plateau) Using R-236fa, this large degradation in heattransfer with increasing film Reynolds number was not seen;

in fact, there was almost no change in the heat transfer mance with increasing film Reynolds number (only fluctuationwithin a bound region) For both 3D enhanced tubes, as well

perfor-as both refrigerants, the local heat flux on a tube in the arraywas correlated as a function of the condensation temperaturedifference and the condensate inundation in the form of the filmReynolds number falling on the tube The coefficients in thecorrelation were found to be close for both tubes apart from the

coefficient c, which corresponds to the slope in the

deteriora-tion in heat transfer performance with increasing film Reynoldsnumber When using R-134a, the heat transfer coefficient of theGewa-C condensing tube decreases less rapidly with increasingfilm Reynolds number; however, the peak reached is not as large

as that found using the Turbo-C tube Using R-134a, the meanrelative error of the fluid/enhanced tube specific method wasless than 1%, with a standard deviation of less than 10% UsingR-236fa, the measurements were predicted by their respectivemethods with mean relative errors of less than 3% and standarddeviations of less than 18%

NOMENCLATURE

a prediction method constant, W/m2-K

b prediction method constant

C Wilson plot method constantheat transfer engineering vol 31 no 10 2010

M CHRISTIANS ET AL 807

c prediction method constant, W/m2-K

cp specific heat at constant pressure, J/(kg-K)

d prediction method constant

e prediction method constant

h local heat transfer coefficient, W/(m2-K)

hlv heat of vaporization (J/kg)

k thermal conductivity, W/(m-K)

M molar mass (kg/kmol)

˙

m mass flow rate, kg/s

p center to center tube pitch, m

pcrit critical pressure, kPa

q local heat flux relative to a surface, W/m2

R thermal resistance m2K/W

r tube radius, m

Re film Reynolds number, 4/µ

U overall thermal resistance, K/W

x coordinate in axial direction, m

Greek Symbols

 T condensation temperature difference, Tsat− Tw

ε mean relative error

 film mass flow rate on one side per unit length of

tube, kg/(m-s)

θ slinging angle, rad

θcrit critical deflection angle, defined by Eq (4), rad

ρ density, kg/m3

σ standard deviation

µ kinematic viscosity, Pa-s

Subscripts

bottom at the bottom of the tube

i internal side of tube

gni Gnielinski (heat transfer coefficient)

l saturated liquid

n number of rows measured from top row

o external side at fin tip

sat saturated conditions

top at the top of the tube

v saturated vapor

REFERENCES

[1] Jung, D., Chae, S., Bae, D., and Yoo, G., Condensation Heat

Transfer Coefficients of Binary HFC Mixtures on Low Fin and

Turbo-C Tubes, International Journal of Refrigeration, vol 28,

no 2, pp 212–217, 2005

[2] Jung, D., Kim, C.-B., Cho, S., and Song, K., Condensation HeatTransfer Coefficients of Enhanced Tubes With Alternative Refrig-

erants for CFC11 and CFC12, International Journal of

Refriger-ation, vol 22, no 7, pp 548–557, 1999.

[3] Jung, D., Kim, C.-B., Hwang, S.-M., and Kim K.-K., sation Heat Transfer Coefficients of R22, R407C, and R410a

Conden-on a HorizConden-ontal Plain, Low Fin, and Turbo-C Tubes,

Interna-tional Journal of Refrigeration, vol 26, no 4, pp 485–491,

Integral-Fin Tubes With Trapezoidal Fins, Heat Transfer

Engi-neering, vol 21, no 2, p 29, 2000.

[6] Kumar, R., Gupta, A., and Vishvakarma, S., Condensation of 134a Vapour Over Single Horizontal Integral-Fin Tubes: Effect of

R-Fin Height, International Journal of Refrigeration, vol 28, no 3,

Low-Fin and Turbo-C Tubes, International Journal of

Refrigera-tion, vol 30, no 5, pp 805–811, 2007.

[10] Gst¨ohl, D., and Thome, J R., Film Condensation of R-134a onTube Arrays With Plain and Enhanced Surfaces: Part I, Experi-

mental Heat Transfer Coefficients, Journal of Heat Transfer, vol.

[12] Habert, M., Ribatski, G., and Thome, J R., Experimental Study

on Falling Film Flow Pattern Map and Intercolumn Distance With

R-236fa, ECI International Conference on Boiling Heat Transfer,

Spoleto, Italy, 2006

[13] Robinson, D M., and Thome, J R., Local Bundle Boiling Heat

Transfer Coefficients on a Plain Tube Bundle (RP-1089), HVAC

and R Research, vol 10, no 1, pp 33–51, 2004.

[14] Gnielinski, V., New Equations for Heat and Mass Transfer in

Turbulent Flow Through Pipes and Ducts, Forschung Im

Inge-nieurwessen, vol 41, no 1, pp 359–368, 1975.

[15] NIST, NIST Thermodynamic Properties of Refrigerants and

Refrigerant Mixtures Database, ver 8.0, Gaithersburg, MD,

Marcel Christians is a Ph.D student at the

Labora-tory of Heat and Mass Transfer at the Swiss eral Institute of Technology in Lausanne (EPFL), Switzerland He received his B.Eng and M.Eng (me- chanical) degrees at the University of Pretoria, South Africa, where his thesis topic covered in-tube conden- sation of refrigerants in the intermittent flow regime.

Fed-His current research is on falling film flow tion, as well as falling film evaporation and conden- sation heat transfer on bundles of enhanced tubes.

visualiza-Mathieu Habert performed his Ph.D thesis on

falling film evaporation on single rows and bundles

of plain and enhanced tubes at the Laboratory of Heat and Mass Transfer at the Swiss Federal Institute of Technology in Lausanne (EPFL), Switzerland, com- pleting his degree in February 2009 Currently, he is chief technical officer of CHS in Gland, Switzerland.

John R Thome has been a professor of heat and

mass transfer at the Swiss Federal Institute of nology in Lausanne (EPFL), Switzerland, since 1998 His primary interests of research are two-phase flow and heat transfer, covering boiling and condensation

Tech-of internal and external flows, two-phase flow terns and maps, experimental techniques on flow vi- sualization and void fraction measurement, and more recently two-phase flow and boiling in microchan- nels He received his Ph.D at Oxford University, England, in 1978, and was formerly an assistant and associate professor

pat-at Michigan Stpat-ate University He left in 1984 to set up his own

interna-tional engineering consulting company He is the author of four books,

En-hanced Boiling Heat Transfer (Taylor & Francis, 1990), Convective Boiling and Condensation (Oxford University Press, 1994, 3rd ed., with J G Col-

lier), Wolverine Engineering Databook III (2004), and Nucleate Boiling on

Micro-Structured Surfaces (with M E Poniewski, 2008), which are now

avail-able free at http://www.wlv.com/products/databook/db3/DataBookIII.pdf and http://www.htri-net.com/ePubs/NucleateBoiling.pdf He received the ASME Heat Transfer Division’s Best Paper Award in 1998 for a three-part paper on

flow boiling heat transfer published in the Journal of Heat Transfer He also authored the chapter “Boiling” in the new Heat Transfer Handbook (2003) He

is an associate editor of Heat Transfer Engineering and is chair of ALEPMA

(the Aluminum Plate Fin Heat Exchanger Manufacturers Association).

heat transfer engineering vol 31 no 10 2010

Copyright
C Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903547487

Film Condensation of R-134a and

R-236fa, Part 2: Experimental

Results and Predictive Correlation

for Bundle Condensation on

Enhanced Tubes

MARCEL CHRISTIANS, MATHIEU HABERT, and JOHN R THOME

Laboratory of Heat and Mass Transfer (LTCM), Faculty of Engineering Science Ecole Polytechnique F´ed´erale de Lausanne

(EPFL), Lausanne, Switzerland

Local test results for two enhanced condensing tubes (next-generation versions of the Wieland Gewa and Wolverine Turbo

enhanced condensing tubes) are obtained for refrigerants R-134a and R-236fa on the center row of a three row-wide tube

bundle The “bundle effect” on the heat transfer performance of the test section is observed and described New predictive

methods for falling film condensation on bundles are proposed, based on a modification of the previous vertical

single-row method with condensate slinging The modifications performed to the experimental setup to allow for bundle tests are

described For two types of enhanced tubes and two refrigerants, the local heat flux is correlated as a function of condensation

temperature difference, the film Reynolds number, the tube spacing, and liquid slinging effect.

INTRODUCTION

The heat transfer performance of tubes in shell-and-tube

con-densers is a function of a large amount of variables Not only is

it dependent on the condensate inundation from the tubes above,

but the geometric distribution of the tubes can also affect the

performance In order to increase the efficiency of falling film

condensers, it is necessary to accurately characterize the

perfor-mance of new enhanced tubes in a test section that attempts to

approximate actual conditions in a bundle This is not to say that

single-row condensing tests are not necessary; on the contrary,

they are necessary to understand the fundamental flow around

these tubes before trying to understand the more complex flow

that occurs in a bundle Furthermore, it is of interest to test the

The authors thank the laboratory’s industrial sponsors Johnson Controls,

Trane, Wieland Werke, and Wolverine Tube, Inc., for funding this study Special

thanks to the tube manufacturers, Wieland Werke and Wolverine Tube, Inc., for

supplying the tubes utilized.

Address correspondence to Prof John R Thome, Laboratory of Heat and

Mass Transfer (LTCM), Faculty of Engineering Science, Ecole Polytechnique

F´ed´erale de Lausanne (EPFL), Station 9, Lausanne CH-1015, Switzerland.

E-mail: john.thome@epfl.ch

performance of these tubes with several refrigerants, such thatthe effect of the thermophysical properties of each fluid may bequantified and taken into account during the design stage

PREVIOUS HEAT TRANSFER COEFFICIENT STUDIES

In 1994, Huber et al [1–3] tested a 5× 5 bundle using finnedand three-dimensional (3D) enhanced tubes using R-134a and R-

12 as test refrigerants The bundle was arranged using horizontaltubes with a vertical pitch of 19.1 mm and a horizontal pitch

of 22.2 mm The finned tubes tested had 1,024 fins per meterand 1,574 fins per meter, while the two 3D enhanced tubestested were the Gewa-SC and the Turbo-Cii The saturationtemperature for these tests was 35◦C In their test section, alltubes were cooled with water, only the middle tubes of each rowwere instrumented, and the tube length-averaged heat transfercoefficients were measured Concurrently with [1–3], Cheng andWang [4] tested plain, finned, and 3D enhanced tubes in a threerows wide by two columns deep bundle Adjacent tubes wereconnected by U-bends The horizontal pitch used was 30 mmwith a vertical pitch of 50 mm R-134a was tested at a saturation809

temperature of 38◦C They used finned tubes with 1,024, 1,260,

and 1,614 fins per meter, and three types of 3D enhanced tubes

The measured heat transfer coefficients were the average values

of the two-pass axial length Rewerts et al [5] used the same test

section and the same tubes as in their previous research [1–3]

They also tested at the same saturation temperature with R-134a

Again, only the middle tubes of each row were instrumented

and the heat transfer coefficients measured were tube

length-averaged All the tubes were cooled with water

Belghazi et al [6–8] used plain, commercially available

finned tubes and enhanced tubes in a staggered bundle

com-prising of 33 tubes distributed in 13 horizontal rows The

hori-zontal spacing was 24 mm and the vertical spacing was 20 mm

There was one cooled tube on each odd row, and there were two

cooled tubes on even rows They tested several refrigerants and

refrigerant mixtures, namely, R-134a and mixtures of R-23 and

R-134a, with the concentration of R-23 varying from 0 up to

11% The finned tubes had 433, 748, 1,024, 1,260, and 1,574 fins

per meter The enhanced tube tested as the Gewa-C+ The

sat-uration temperature for these tests was 40◦C The researchers

measured average tube heat transfer coefficients using a

Wil-son plot method In [6], the authors presented a tube-specific

(Gewa-C+) method that took into account the drainage around

the enhancement structure of the tube

Honda et al [9–13] tested finned tubes with several

refriger-ants for both in-line and staggered tube bundles The bundle was

comprised of 38 tubes distributed in 15 rows and three columns,

with a horizontal spacing of 22 mm and a vertical spacing of 22

mm The tubes tested had 1,040, 1,923, and 2,000 fins per meter

They tested R-123, and a mixture of R-134a (14%) and R-123

The heat transfer coefficients were calculated according to the

average row heat flux based on a water-side energy balance

As such, the aim of this article is to first detail the

modifi-cations performed on the LTCM installation allowing for

con-densation bundle tests to be performed, and second, to present

and discuss the results obtained in the LTCM’s bundle falling

film facility for advanced versions of the Turbo C and Gewa C

3D enhanced tubes, using both R-134a and R-236fa The heat

transfer coefficients measured here are local values rather than

1

3 2

4 5 6

6 First test tube

Figure 1 Original test section as used by Gst¨ohl and Thome [14, 15].

1 2

Figure 2 Top part of the test section as modified to run bundle tests.

tube length-averaged values and hence are more useful for thedevelopment of prediction methods Finally, a bundle predictionmethod based on the single-row method presented in Part 1 ofthis article (this issue) and in [14] is proposed

EXPERIMENTAL FACILITY

The basics of the experimental setup utilized in this studyare unchanged from those described in Part 1 of this article(this issue) and [15] The modifications that were required toconvert the single-row test section to a bundle are discussed inthis section

A part of the original test section is shown in Figure 1.Instead of using only the single distributor on top of the centertube array (single equi-spaced column of tubes), two sets of twohigh-performance condensing tubes were installed on both sides

of the distributor, through which cold glycol is run regardless ofwhether condensation or evaporation is being studied (Figure 2).These four tubes provide the condensate overfeed for the siderows The glycol mass flow rate can be controlled to regulatethe amount of condensate being generated To further fine tunethe amount of heat exchanged, a three-way valve is installed.Figure 3 shows the additional circuit installed that feeds thetwo side-row overfeed circuits The glycol flow rate through

Figure 3 Side overfeed circuits (glycol).

heat transfer engineering vol 31 no 10 2010

M CHRISTIANS ET AL 811

Figure 4 Auxiliary side-array circuits (water).

each side can be manually controlled such that there will be no

imbalance in the heat extraction between the two sides

To better simulate conditions in real condensers, tubes were

also installed around the center column of tubes arranged in a

staggered equilateral triangle layout In total, there are 22

periph-eral tubes around the six center column instrumented tubes, with

a vertical pitch of 38.5 mm and a horizontal pitch of 22.3 mm (as

recommended by our industrial sponsors) The side-array tubes

are also partially shown in Figure 2 Water is pumped through

these tubes; however, unlike the center column, the water goes

through all the side tubes in 11 passes (rather than the two passes

in the center column for each pair of tubes) The water flow rate

through each side-array can be controlled in a similar fashion to

the glycol overfeed circuit, such that the water-cooled side-array

condensate flow remains balanced The auxiliary water circuit

installed for the bundle tests is shown in Figure 4

EXPERIMENTAL ERRORS AND PROCEDURES

As the central row of tubes was not changed with the

instal-lation of the additional tubes, the internal enhancement

coeffi-cients calculated using the Wilson plot method also remained

unchanged This means that both the values calculated and the

uncertainties tabulated in Table 1 of Part 1 remain valid and

are not presented in this section Furthermore, since the

mea-surement method utilized in the center column of tubes has not

changed either, and neither were the tubes, the uncertainties

tab-ulated in Table 2 of Part 1 are also valid and not repeated here

The saturation temperature was kept unchanged at 31◦C

For experiments involving condensate overfeed, the film flow

rate of the liquid arriving on the first tube was evaluated from

the measured mass flow rate and the tube length, assuming that

the refrigerant is at saturation conditions The mass flow of

re-frigerant condensing on the first tube is calculated by an energy

balance on differential elements and added to the film flow rate

arriving on the first tube to obtain the film flow rate at the top

of the second tube and so on This means, however, that it isassumed that there is an ideal flow on the central row, withoutslinging onto or from the side rows In the case of no over-feed, a similar procedure was applied, with the initial flow rateonto the top tube set to 0 Any slinging from the side-overfeed

or top side-array tubes onto the first tube are ignored Again,

as in Part 1, the saturation temperatures, as well as the port and thermodynamic properties, are calculated according toREFPROP v8 [16] from the mean of the pressures measured bypressure transducers above and below the tube array The pres-sure drops from top to bottom of the bundle were in fact quitesmall, and the subsequent change in thermal properties thereforenegligible

trans-Due to the fact that there are now two additional controllableoperating conditions or states (i.e., the heat transferred fromthe glycol-cooled side-overfeed circuits and the water-cooledside-row circuits) in the test section, and the fact that we areonly acquiring data for the center column, it was required thatthe influence of these conditions on the heat transfer behavior

be quantified A further objective was to identify the resultsfor which a true “bundle effect” could be distinguished andthus use those to establish our “bundle database” for use inbuilding such a prediction method Thus, the presentation ofthe bundle heat transfer results is presented first by refrigerant,and then the state A test state (condition) is defined as anycombination of the glycol side-overfeed condensate flow andwater side-array condensate flow that will create a distinctlydifferent environment for the center tube row There are fourdistinct states achievable in the bundle, namely, both glycol andwater flow rates at maximum values, zero for both circuits (i.e.,

at 0 kg/s), the glycol flow rate at its maximum with no waterflow rate, and vice versa These four conditions are illustrated

in an approximate schematic of each in Figure 5 The notionthat specific inundation rates on the central vertical row are stillachieved by using the overfeed pump (which goes on to themulti-part distributor and the half-tube) should be clear Refer

to Part 1 of this article for a schematic of the test facility.These four states will be mentioned often; for brevity’s

sake, in all graphs they will be indicated as mOmSA mum side-overfeed, maximum side-array), mOnSA (maximum

(maxi-No side-overfeed flow Maximum side-array flow (nOmSA) Maximum Side-overfeed flow

Maximum Side-array flow (mOmSA)

Maximum Side-overfeed flow

No Side-array flow (mOnSA)

No side-overfeed flow

No side-array flow (nOnSA)

Figure 5 Flow variable states for the glycol and water side-array auxiliary loops.

heat transfer engineering vol 31 no 10 2010

side-overfeed, no side-array), nOmSA (no side-overfeed,

maxi-mum side-array), and nOnSA (no side-overfeed, no side-array).

Tests were conducted by gradually decreasing the liquid

over-feed flow rate (from the pump) on the center top tube at a fixed

heat flux This liquid overfeed is completely independent of

the previously mentioned bundle states This liquid flow rate

is measured using a Coriolis flow meter and can be very

accu-rately controlled The data were logged only once steady-state

conditions were attained A more detailed description of the test

facility, data reduction methods and measurements accuracies

can be found in Gst¨ohl and Thome [14, 15], as well as in Part 1

of this two-part article

EXPERIMENTAL RESULTS WITH THE BUNDLE TEST

SECTION

Search for the “Bundle Effect”

Tests were performed using the Wolverine Turbo and

Wieland Gewa condensing tubes (both of them have 18.38 mm

nominal outer diameter) provided by the tube manufacturers

Before installation into the test section, the tubes were

thor-oughly cleaned The film flow rate (per unit length) was varied

from 0.25 kg/m-s to 0 kg/m-s Tests were performed at constant

tube array nominal heat fluxes of 20, 40, and 60 kW/m2 The

tests were repeated for each variable state condition, such that a

comparison could be made

Due to the large amount of data obtained (two tubes, two

refrigerants, and four state combinations), representative figures

are presented to advance the discussion of the appearance of a

“bundle effect.” Unless stated otherwise, the trends presented are

indicative of the behavior at all heat fluxes, for both refrigerants,

and for both tubes Once the comparison of the four states is

finished, the results independent of the variable states will be

shown

mOmSA

Figure 6 presents the heat transfer performance of the six

center tubes in the bundle with the side-overfeed and side-row

circuits exchanging the largest amount of heat possible It is

immediately clear that there is a large difference for the bundle

heat transfer coefficients measured as opposed to the single-row

results of Figures 2–7 of Part 1 The first significant aspect is that

the bundle seems to have completely flattened out the peak that

was present in the single-row results This is advantageous for

condenser performance since tubes 4 through 6 have flattened

out at quite high values Furthermore, it is also clear that the

results are no longer grouped together on more or less one

“curve.”

In addition, starting from the first (top) tube in the center row

of the bundle, the measured heat transfer coefficient is very low

and then rises monotonically from tube to tube up to the fifth

0 500 1000 1500 2000 2500 3000 3500 4000 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

Figure 6 Turbo performance using R-134a in the tube bundle at a nominal

bundle heat flux of 40 kW (mOmSA), representative of all heat fluxes and tubes.

tube, after which there is a decrease in the measured heat fer values of the sixth tube This trend seems to be attributable

trans-to “entrance” and “exit” effects on tubes 1–3 and tube 6, tively Instead, comparing the bundle performance of tubes 4, 5,and 6 to those of the single row at a film Reynolds number of1,000 shows that the performances are comparable to the valuespresented in Part 1 of this article A reason that might explainthe decrease in the performance of the sixth tube with respect tothe fifth tube is the fact that due to the geometric constraints im-posed by the test section vessel itself, it is the lowermost tube inthe bundle, and is not completely surrounded by the side-arraytubes, as shown in Figure 7 Thus, not having other tubes for thecondensate to fall onto would also affect the heat transfer due

respec-to the change in the liquid film flow characteristics (i.e., more

of the condensate will flow onto tube 6 than tube 5)

In this configuration, the heat transfer performance of thebundle is essentially constant as a function of the heat flux.There is some variation on tubes 1, 5, and 6 at low Reynoldsnumbers, but as this increases, the results quickly collapse into

a similar range For the top tube rows, the overfeed condensateneeds to get distributed and may create a “flooding” effect thatreduces their heat transfer coefficients

M CHRISTIANS ET AL 813

mOnSA

In essence, the general trends spotted in the results for

the mOmSA results are also found in this subsection’s results.

Figure 8 shows that there is a large separation between the heat

transfer results of the first tube and the rest of the tubes of the

center column The entrance effects on the first tube are

eas-ily visible, as there is a constant increase in the heat transfer

coefficient with increasing mass flow This was not seen in the

previous section’s results

Apart from the first tube, which has an increasing trend, the

rest of the tubes behave in a very constant fashion, and only the

fifth shows a small decrease in heat transfer performance at the

highest Reynolds numbers The sixth tube again has an “exit”

effect

In comparison with the previous results, it can be seen that

the first tube has lower heat transfer coefficients throughout the

entire Reynolds number range, as well as at the different heat

fluxes It is proposed that there are two main reasons for this

phenomenon; first, this may be due to the variability in the heat

flux when testing at a nominal bundle heat flux—i.e., each tube

has its own midpoint so reporting here the nominal heat flux

does not show this effect (the actual heat fluxes are used later

for the prediction method) However, this effect alone will not

decrease the heat transfer coefficient of the first tube so

drasti-cally Second, as noted earlier, it is possible that in the bundle

test section, liquid is retained both above and around the first

tube, in a flooding effect, which would lead to decreased

perfor-mance throughout the Reynolds number range This would also

depend on the enhanced tube utilized and its particular drainage

characteristics This phenomenon is found for both Wieland and

Figure 8 Gewa performance using R-236fa in the tube bundle at a nominal

bundle heat flux of 20 kW (mOnSA).

0 500 1000 1500 2000 2500 3000 3500 4000 4500 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

Figure 9 Gewa performance using R-134a in the tube bundle at a nominal

bundle heat flux of 6 0kW (nOmSA).

nOmSA

As shown in Figure 9, due to the absence of the side-overfeed,the first three tubes show a tendency to decrease the heat transfercoefficient with an increase in Reynolds number, much likethe single-array tests shown in Part 1 There is a noticeabledifference between the top three tubes and the bottom threetubes of the bundle The top three tubes show a very slightdecrease in measured heat transfer coefficient, while the bottomthree show an equally slight increase There is a very largedifference (between 60 and 100%) between the top and bottomtubes In all probability, this is a “bundle effect” acting on thethree bottom tubes, which receive the redistributed film flow

on them, as opposed to the top three, which again appear tohave a flooding effect created by the configuration of the testsection

nOnSA

The results found when neither the side-overfeed nor theside-array circuits are active show that the heat transfer of thetop three tubes degrades as a function of Reynolds number;however, as can be seen in Figure 10, this degradation onlyhappens at higher Reynolds numbers (tubes 2 and 3), while tube

1 degrades almost immediately However, unlike the single-row

or the nOmSA bundle tests, there is a heat transfer recovery

that essentially increases the heat transfer coefficient back topre-degradation levels

Of further interest is that over the three heat fluxes, the heattransfer performance of the center column in the bundle in-creases as the tube number increases (except tubes 2 and 3; theyhave essentially the same performance), which was not the casewhen the side-overfeed circuits were active What is more, the

fifth tube (which was the better performing tube in the mOmSA

heat transfer engineering vol 31 no 10 2010

Figure 10 Turbo performance using R-236fa in the tube bundle at a nominal

bundle heat flux of 20 kW (nOnSA).

and mOnSA cases) is not the best performing tube any longer,

which serves to show that in the nOnSA case, the heat transfer

degradation effect on the sixth tube caused by having installed

the auxiliary tubes is diminished (yet not negligible)

R-134a Variable State Comparison

As has been shown in the preceding subsection, the first three

tubes appear to be heavily influenced by the variation in the side

flow conditions, while for the sixth tube results are influenced

by its position as the bottom-most tube in the bundle Plotting

the results obtained for the fourth and fifth tubes for all four

states tested together (separated by heat flux), it may thus be

possible to satisfactorily conclude whether these tubes show

true “deep-in-the-bundle” heat transfer behavior

Turning to Figure 11 (which is representative of both tubes

at all heat fluxes for R-134a), the effect of the side conditions

is evident, especially as the results from the nOnSA case are

compared to the results found when one or both of the side

circuits are active Over the entire range of Reynolds number,

the heat transfer performances of the fourth and fifth tubes when

in nOnSA conditions are either the lowest or among the lowest

measured This is not only seen in the 20-kW/m2 results, but

also in the 40- and 60-kW/m2data

Of particular interest is that one of the definitive “bundle

ef-fects” is that the degradation found in the single-row results is

not present any longer This is very probably due to the

redis-tribution of the flow normally incoming to tubes 4 and 5 to the

side-array tubes This effect is seen repeated in the results taken

at 40 and 60 kW/m2(not shown for brevity)

Finally, over the three heat fluxes tested, almost constant heat

transfer coefficients were measured on the fifth tube, with any

or both of the side circuits active (i.e., mOmSA, mOnSA, or

nOmSA cases) The only measurements from the fifth tube that

500 1000 1500 2000 2500 3000 3500 4000 4500 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

Figure 11 Comparison of the four variable states on tubes 4 and 5 for the Turbo condensing tube and R-134a at a heat flux of 40 kW/m 2 , representative

of both tubes and all heat fluxes.

were considerably lower than the rest were those found when

testing the nOnSA case.

R-236fa Variable State Comparison

For a heat flux of 20 kW/m2, the comparison is plotted inFigure 12 At this heat flux, it can be seen that the major-ity of the results are relatively constant For tube 5, the re-

sults found when using mOnSA, nOmSA, and nOnSA are almost

identical, and only deviate at larger Reynolds numbers

How-ever, the mOmSA data are found to be consistently±15% lowerthan the rest This is contrary to the expected result in whichthe conditions with any of the side circuits active would per-

form similarly, with the nOnSA data being the odd ones out.

The reason why this occurred is not clear, except that thesetubes’ enhancement was not optimized by the manufacturers forR-236fa condensate drainage The results that were obtainedfrom tube 4 also vary much more from case to case than whenR-134a was utilized

At 40 kW/m2 in Figure 13, the results presented are sentative for the two tubes and the test heat fluxes of 40 and

repre-60 kW/m2(the results shown in Figure 12 were the exception).The results on the fifth tube show the same trends as when usingR-134a, namely, that regardless of the side-state status, all thedata are grouped The results for the fourth tube show that the

data from mOmSA and the nOmSA cases exhibit very similar trends, while the results from the mOnSA and nOnSA cases are

similar

The bundle configuration showed much larger gains in formance for R-236fa, and particularly the Wieland condens-ing tube/R-236fa results achieved almost the same performance

per-as when using R-134a, which wper-as an unexpected result Thisphenomenon is seen repeated in the results taken at 40 andheat transfer engineering vol 31 no 10 2010

Figure 12 Comparison of the four variable states on tubes 4 and 5 for the

Gewa condensing tube and R-236fa at a heat flux of 20 kW/m 2

60 kW/m2 This was seen for the Wolverine Turbo condensing

tube as well

Results Independent From Variable State Conditions

The preceding results showed that to obtain results

indepen-dent of the variation in the side states, the top three tubes and

the sixth tube could not be used as it is evident that they were

influenced by entrance and exit effects, respectively Isolating

the fourth and fifth tubes, it was found that the results were

independent of the side states as long as one or both of the side

circuits were active

Film Reynolds number, Rebottom [ − ]

2 K]

R−236fa, Bundle, Gewa Condensing Tube, tube spacing 38.5mm, heat flux: 40kW/m2

mOmSA T4 mOnSA T4 mOnSA T5 nOmSA T5 nOnSA T4

Figure 13 Comparison of the four variable states on tubes 4 and 5 for the

Gewa condensing tube and R-236fa at a heat flux of 40 kW/m 2 , representative

of both tubes and all heat fluxes.

Using the Turbo enhanced condensing tube and R-134a, at aheat flux of 20 kW/m2(all Reynolds numbers) the average heattransfer coefficient measured is roughly 17–19 kW/m2-K (tubes

4 and 5) This can be compared to the results found in the row results (Figure 2 of Part 1), in which the average heat transfer

single-coefficient before degradation (Re <∼1,000) is an average of

22 kW/m2-K Of particular interest is that one of the definitive

“bundle effects” is that the progressive degradation found in thesingle-row results with increasing film Reynolds number is nolonger present However, while there is no degradation, there is

no sharp peak either The absence of the peak is probably due

to the redistribution of the flow between tubes 4 and 5 and theside-array tubes This effect is seen repeated in the results taken

at 40 and 60 kW/m2 The results obtained for all heat fluxes,using the Turbo condensing tube and R-134a, are shown inFigure 14

Utilizing the Wieland Gewa condensing tube and R-134a,the average heat transfer coefficient measured is roughly15–15.5 kW/m2-K (tubes 4 and 5) In the single-row results(Figure 5 of Part 1), the average heat transfer coefficient before

degradation (Re <∼1,000), and excluding the first tube results,was an average of 17 kW/m2-K The same bundle effect de-scribed earlier is thus also present in Wieland’s tube data, andagain the degradation in heat transfer coefficient with increasingReynolds number is not present anymore and neither is the peak.The heat transfer data (all heat fluxes) for the Gewa condensingtube are presented in Figure 15

Going back to the Turbo condensing tube, using R-236fa, theaverage heat transfer coefficient measured is roughly 13 kW/m2-

K (tubes 4 and 5, all heat fluxes) In the single-row results (again,Figures 2–4 of Part 1), the average heat transfer coefficient(excluding tube 1) was 7–7.5 kW/m2-K Unlike R-134a, therewas no worsening of the heat transfer peak performance—infact, the bundle testing resulted in higher measured heat transfer

500 1000 1500 2000 2500 3000 3500 4000 4500 0

5000 10000 15000 20000 25000

Film Reynolds number, Re

Figure 14 Variable state independent results found on tubes 4 and 5 of the test section: Turbo condensing tube, all heat fluxes for R-134a.

heat transfer engineering vol 31 no 10 2010

Figure 15 Variable state independent results found on tubes 4 and 5 of the

test section: Gewa condensing tube, all heat fluxes for R-134a.

coefficients throughout the Reynolds number range, especially

on tubes 4 and 5, which can be seen in Figure 16 The fact

that this tube was designed for use with R-134a affects the

condensate drainage from the enhancement; this effect coupled

to the redistribution of the flow around tubes 4 and 5 is the

reason why there is an increase in performance Again, since

the results are relatively heat flux independent, the results for all

heat fluxes have been presented on the same figure

Finally, for the Gewa tube with R-236fa, the average heat

transfer coefficient measured for tubes 4 and 5 is roughly 13–

13.5 kW/m2 In the single-row R-236fa results, the average

(without the first [top] tube) was 9 kW/m2-K (Figures 5–7 of Part

1) Again, the bundle configuration showed a large performance

gain when using R-236fa Figure 17 shows the results for all

Figure 16 Variable state independent results found on tubes 4 and 5 of the

test section: Turbo condensing tube, all heat fluxes for R-236fa.

heat fluxes since the trends and magnitudes are the same It ispossible that, due to the increased surface tension of R-236fa(with respect to R-134a) and to the relatively tight spacing ofthe tubes, the redistribution of refrigerant was more pronounced.Redistribution, in our opinion, has an effect of thinning out theaverage thickness of the film over the tubes, thus increasing theheat transfer coefficient

BUNDLE PREDICTION METHOD

The method formulated in [14] and used in Part 1 of thisarticle, for single-row falling film condensation on plain andenhanced tubes, correlated the heat flux as a function of the filmReynolds number and the wall temperature difference, whenthere was no slinging, as

q o = (a + cRe top )
T b (1)

To include the effect of the fraction of condensate that would

be slung off the tubes defined with respect to the critical slingingangle, the heat flux when there was slinging and the maximumslinging angle are determined as follows:

5000 10000 15000 20000 25000

Film Reynolds number, Re

Figure 17 Variable state independent results found on tubes 4 and 5 of the test section: Gewa condensing tube, all heat fluxes for R-236fa.

heat transfer engineering vol 31 no 10 2010

M CHRISTIANS ET AL 817

Table 1 Coefficients in Eqs (1)–(5) and relative errors of the prediction

method (using the first set of parameter constraints) for the bundle data

a b c d e ε σ Tube Refrigerant [W/m 2 -K] [ — ] [W/m 2 Kb] [ — ] [ — ] [%] [%]

The preceding method was shown to correctly predict the

results obtained by Gst¨ohl and Thome [14] with the previous

tubes and the same test section and refrigerant; it also performed

well with the new tubes tested, R-134a and R-236fa Due to this

previous empirical success, it was decided to attempt to fit the

data found with the bundle test section to the same form of

equation

Using the same nonlinear least-squares algorithm developed

for the single-row results on the data obtained on the fourth and

fifth tubes of the test facility, the resulting empirical coefficients

are presented in Tables 1 and 2 In each table, a set of empirical

coefficients for each tube/refrigerant combination is presented,

with each showing essentially the same goodness of fit

The main difference between the two methods stems from

the assumptions utilized to run the optimization For the first

(method 1), it was assumed that the slinging angle should be

equal to 0 for a Reynolds number of 0 (i.e., e= 0), while the

second (method 2) allowed it to find a minimum for a non-null

value by penalizing a zero value of e For the methods in Table 1

for which both d and e are equal to 0, the prediction method

utilized collapses to the form shown in Eq (1)

For the Turbo condensing tube/R-134a combination, the fit

predicted 75% of the data within a±15% error band and 95%

of the data within±30% When using R-236fa, the Turbo

con-densing method predicted only 60% of the data within a±15%

error band but 90% of the data within±30% A comparison of

the first method against the R-134a and R-236fa data is shown

in Figures 18 and 19, respectively

In the case of the Turbo condensing tube (R-134a data), the

multiplier in front of the Reynolds number c (in the first method)

Table 2 Coefficients in Eqs (1)–(5) and relative errors of the prediction

method (using the second set of parameter constraints) for the bundle data

a b c d e ε σ Tube Ref [W/m 2 -K] [ — ] [W/m 2 K b ] [ — ] [ — ] [%] [%]

Film Reynolds number, Rebottom,n−1 [ − ]

2 K]

R−134a, Turbo condensing tube, tube spacing 38.50mm

Tube 4 Model

Figure 18 Comparison of the prediction (method 1) with the bundle T4 and T5 Turbo condensing tube data using R-134a.

acts to suppress the influence of an increase in the Reynoldsnumber For R-236fa (first method) and for the second method(both R-134a and R-236fa), it is used to magnify the effect of

an increase in the film Reynolds number The constant a gives

the “height” of the performance plateau, while the temperature

difference exponent b is relatively close to 1 and decreases the effect of the temperature drop (the resultant b−1 leads to

an exponent both small and negative), by shallowing-out theprediction For the same tube (and both refrigerants), the second

method utilizes a nonzero slinging angle constant e (in radians,

equivalent to an angle of –9◦off the y-axis for R-134a, –35◦for

R-236fa) to offset the onset of slinging The small multiplier d in

front of the Reynolds number decreases the effect of the increasewith Reynolds number, and is also used to effectively retard

the onset of slinging The temperature difference exponent b

0 5000 10000 15000 20000 25000

Film Reynolds number, Rebottom,n−1 [ − ]

2 K]

R−236fa, Turbo condensing tube, tube spacing 38.50mm

Tube 4 Model

Figure 19 Comparison of the prediction (method 1) with the bundle T4 and T5 Turbo condensing tube data using R-236fa.

heat transfer engineering vol 31 no 10 2010

Film Reynolds number, Rebottom,n−1 [ − ]

2 K]

R−134a, Gewa condensing tube, tube spacing 38.50mm

Tube 4 Model

Figure 20 Comparison of the prediction (method 1) with the bundle T4 and

T5 Gewa condensing tube data using R-134a.

essentially stays unchanged, as in the first method Furthermore,

although a is larger in the second method than in the first, the

much larger c increases the effect of the Reynolds number,

producing a difference between the two terms in the parentheses

(a + cRe) that is on the order of the first method.

Turning to the Wieland Gewa condensing tube, the R-134a

fits (shown in Tables 1 and 2) predict 86% of the data within

a±15% error band and 100% of the data within ±30% Using

R-236fa, both of the fits presented predict 80% within a±15%

error band and 95% within±30% The assumptions utilized to

run the optimization were the same as those delineated earlier

A comparison of the first method against the use of R-134a and

R-236fa is shown in Figures 20 and 21, respectively

The multiplier in front of the Reynolds number, c, in the first

method, again acts to suppress the influence of an increase in the

Film Reynolds number, Rebottom,n−1 [ − ]

2 K]

R−236fa, Gewa condensing tube, tube spacing 38.50mm

Tube 4 Model

Figure 21 Comparison of the prediction (method 1) with the bundle T4 and

T5 Gewa condensing tube data using R-236fa.

Reynolds number for R-134a However, for R-236fa (method

1), or both refrigerants in method 2, c is used to magnify the

effect of the Reynolds number, since it is also being ously decreased by the slinging angle multiplier The constant

simultane-aagain gives the “height” of the performance plateau, while the

temperature difference exponent b is close to 1 and decreases

the effect of the temperature drop by shallowing-out the diction In fact, since the exponent is positive (for R-134a, bothmethods), this shows that there is a very slight increase in heattransfer performance with increase temperature difference (theexponent corresponds to more or less taking the ninth root ofthe temperature difference)

pre-The second method utilizes a nonzero slinging angle constant

e(in radians, equivalent to an angle of 2◦ off the y-axis for

R-134a, and –15◦for R-236fa) to offset the onset of slinging The

small multiplier d in front of the Reynolds number decreases

the effect of any increase in Reynolds number, and is also used

to effectively retard the onset of slinging The temperature

dif-ference exponent b essentially stays unchanged as in the first method Furthermore, although a is larger in the second method than in the first, the much larger c increases the effect of the

Reynolds number

In summary, for the two tubes test, both methods developedresulted in essentially the same mean relative error and standarddeviation of the prediction However, method 1 is recommended

for use, as the fact that e is set to 0 has physical meaning.

CONCLUSIONS

Modifications were made to the falling film facility such that

a bundle configuration of tubes could be tested under tion conditions A large database of results was gathered fromthe ensuing experimental campaign It was found that for thisparticular tubes and bundle configuration, R-134a performs bet-ter than R-236fa, since these tubes have been optimized for usewith R-134a However, for the Turbo condensing tube, the dif-ference in performance between the two refrigerants was onlyaround±2.5 kW/m2-K on average for tubes 4 and 5; the largeincrease in performance compared to the single-row data of Part

condensa-1, when using R-236fa, was not expected The difference in formance when using the Gewa enhanced condensing tube alsowas remarkably lower than expected Furthermore, it was foundthat when using R-134a as the test refrigerant, the largest bundleeffect was experienced when both the auxiliary water side-arraycircuits and the glycol side-overfeed circuits were active, al-though tubes 4 and 5 were shown to be essentially independent

per-of the side flow states The first tubes per-of the bundle showed aperformance decrease when using R-134a, while with R-236fa

it was found that the results measured were of the same order

as the original single-row data In the case of R-134a, it seemsplausible that there is condensate holdup (flooding) around thefirst tube

Visual comparison of these bundle data with the results found

in the single-row test section experimentation shows that theheat transfer engineering vol 31 no 10 2010

M CHRISTIANS ET AL 819method proposed by [14] involving slinging of the liquid film

off the tube could be modified to predict the R-134a data best

In particular, the trends of the bundle results (when viewed as

an ensemble) more closely resemble those found for the

Turbo-Chil low-fin tube, which had a flatter performance over a large

Reynolds number range, but where results were dependent on

the position in the column itself [15]

The heat transfer method developed by Gst¨ohl and Thome

[14] was modified to fit the data gathered for the bundle

con-figuration The measured results from all six tubes were used

when developing the single-row prediction in Part 1, while only

the fourth and fifth tube data were utilized in the bundle

config-uration, due to “entrance” and “exit” effects on the other tube

rows

The complexity of the trends in tubes from 1 to 6 in the

bundle suggests that local observation of the flows between the

tubes would be valuable to gain a physical insight into the liquid

distribution process

NOMENCLATURE

a prediction method constant, W/m2-K

b prediction method constant

c prediction method constant, W/m2-K

d prediction method constant

e prediction method constant

p center to center tube pitch, m

q local heat flux relative to a surface, W/m2

r tube radius, m

Re film Reynolds number, 4/µ

T temperature, K

Greek Symbols

T condensation temperature difference, T sat − T w

ε mean relative error

 film mass flow rate on one side per unit length of tube,

kg/(m s)

θ slinging angle, rad

θcrit critical slinging angle, rad

σ standard deviation

µ kinematic viscosity, Pa-s

Subscripts

bottom at the bottom of the tube

n number of rows measured from top row

o external side at fin tip

sat saturated conditions

top at the top of the tube

REFERENCES

[1] Huber, J B., Rewerts, L E., and Patee, M B., Shell-Side sation Heat Transfer of R-134a—Part II: Enhanced Tube Perfor-

Conden-mance, Proceedings of the ASHRAE Annual Meeting, Jun 25–29

1994, Atlanta, GA, vol 100, pp 248–256, 1994.

[2] Huber, J B., Rewerts, L E., and Patee, M B., Shell-Side densation Heat Transfer of R-134a—Part III: Comparison With

Con-R-12, Proceedings of the ASHRAE Annual Meeting, June 25–29

1994, Atlanta, GA, vol 100, pp 257–264, 1994.

[3] Huber, J B., Rewerts, L E., and Patee, M B., Shell-Side densation Heat Transfer of R-134a—Part I: Finned-Tube Perfor-

Con-mance, Proceedings of the ASHRAE Annual Meeting, June 25–29

1994, Atlanta, GA, vol 100, pp 239–247, 1994.

[4] Cheng, W.-Y., and Wange, C.-C., Condensation of R-134a on

Enhanced Tubes, Proceedings of the ASHRAE Annual Meeting,

June 25–29 1994, Orlando, FL, USA, vol 100, pp 809–817, 1994.

[5] Huber, J B., Rewerts, L E., and Patee, M B., Effect of R-134a

In-undation on Enhanced Tube Geometries, ASHRAE Transactions,

Tubes, International Journal of Refrigeration, vol 26, no 2, pp.

214–223, 2003

[8] Belghazi, M., Bontemps, A., Signe, J C., and Marvillet, C., densation Heat Transfer of a Pure Fluid and Binary Mixture Out-side a Bundle of Smooth Horizontal Tubes Comparison of Ex-

Con-perimental Results and a Classical Model, International Journal

of Refrigeration, vol 24, no 8, pp 841–855, 2001.

[9] Honda, H., Fujii, T., Uchima, B., Nozu, S., and Nakata, H., densation of Downward Flowing R-114 Vapor on Bundles of

Con-Horizontal Smooth Tubes, Heat Transfer Japanese Research, vol.

18, pp 31–52, 1989

[10] Honda, H., Takamatsu, H., Takada, N., and Makishi, O., densation of HCFC123 in Bundles of Horizontal Finned Tubes:

Con-Effects of Fin Geometry and Tube Arrangement, International

Journal of Refrigeration, vol 19, no 1, pp 1–9, 1996.

[11] Honda, H., Takamatsu, H., and Takada, N., Experimental surements for Condensation of Downward-Flowing R123/R134a

Mea-in a Staggered Bundle of Horizontal Low-FMea-inned Tubes With Four

Fin Geometries, International Journal of Refrigeration, vol 22,

Bundle of Horizontal Finned Tubes: Effect of Fin Geometry,

In-ternational Journal of Refrigeration, vol 25, no 1, pp 3–10,

[15] Gst¨ohl, D., and Thome, J R., Film Condensation of R-134a on

Tube Arrays With Plain and Enhanced Surfaces: Part I,

Experi-mental Heat Transfer Coefficients, Journal of Heat Transfer, vol.

128, pp 21–32, 2006

[16] NIST, NIST Thermodynamic Properties of Refrigerants and

Re-frigerant Mixtures Database, ver 8.0, Gaithersburg, MD, 2007.

Marcel Christians is a Ph.D student at the

Labora-tory of Heat and Mass Transfer at the Swiss eral Institute of Technology in Lausanne (EPFL), Switzerland He received his B.Eng and M.Eng (me- chanical) degrees at the University of Pretoria, South Africa, where his thesis topic covered in-tube conden- sation of refrigerants in the intermittent flow regime.

Fed-His current research is on falling film flow tion, as well as falling film evaporation and conden- sation heat transfer on bundles of enhanced tubes.

visualiza-Mathieu Habert performed his Ph.D thesis on

falling film evaporation on single rows and bundles

of plain and enhanced tubes at the Laboratory of Heat and Mass Transfer at the Swiss Federal Institute of Technology in Lausanne (EPFL), Switzerland, com- pleting his degree in February 2009 Currently, he is chief technical officer of CHS in Gland, Switzerland.

John R Thome has been a professor of heat and

mass transfer at the Swiss Federal Institute of nology in Lausanne (EPFL), Switzerland, since 1998 His primary interests of research are two-phase flow and heat transfer, covering boiling and condensa- tion of internal and external flows, two-phase flow patterns and maps, experimental techniques on flow visualization and void fraction measurement, and more recently two-phase flow and boiling in mi- crochannels He received his Ph.D at Oxford Uni- versity, England, in 1978 and was formerly an assistant and associate pro- fessor at Michigan State University He left in 1984 to set up his own in- ternational engineering consulting company He is the author of four books,

Tech-Enhanced Boiling Heat Transfer (Taylor & Francis, 1990), Convective Boiling and Condensation (Oxford University Press, 1994, 3rd ed., with J G Col-

lier), Wolverine Engineering Databook III (2004), and Nucleate Boiling on

Micro-Structured Surfaces (with M E Poniewski, 2008), which are now

avail-able free at http://www.wlv.com/products/databook/db3/DataBookIII.pdf and http://www.htri-net.com/ePubs/NucleateBoiling.pdf He received the ASME Heat Transfer Division’s Best Paper Award in 1998 for a three-part pa-

per on flow boiling heat transfer published in the Journal of Heat

Trans-fer He also authored the chapter “Boiling” in the new Heat Transfer book (2003) He is an associate editor of Heat Transfer Engineering and is

Hand-chair of ALEPMA (the Aluminum Plate Fin Heat Exchanger Manufacturers Association).

heat transfer engineering vol 31 no 10 2010

Copyright
C Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903547545

Dropwise Condensation Heat

Transfer on Plasma-Ion-Implanted

Small Horizontal Tube Bundles

ALI BANI KANANEH, MICHAEL HEINRICH RAUSCH, ALFRED LEIPERTZ,

Lehrstuhl f¨ur Technische Thermodynamik, Universit¨at Erlangen-N¨urnberg, Erlangen, Germany

Stable dropwise condensation of saturated steam was achieved on stainless-steel tube bundles implanted with nitrogen ions

by plasma ion implantation For the investigation of the condensation heat transfer enhancement by plasma ion implantation,

a condenser was constructed in order to measure the heat flow and the overall heat transfer coefficient for the condensation

of steam on the outside surface of tube bundles For a horizontal tube bundle of nine tubes implanted with a nitrogen ion dose

of 10 16 cm −2 , the enhancement ratio, which represents the ratio of the overall heat transfer coefficient of the implanted tube

bundle to that of the unimplanted one, was found to be 1.12 for a cooling-water Reynolds number of about 21,000 The heat

flow and the overall heat transfer coefficient were increased by increasing the steam pressure The maximum overall heat

transfer coefficient of 2.22 kW ·m −2 ·K −1 was measured at a steam pressure of 2 bar and a cooling-water Reynolds number

of about 2,000 At these conditions, more dropwise condensation was formed on the upper tube rows, while the lowest row

received more condensate, which converted the condensation form to filmwise condensation.

INTRODUCTION

Dropwise condensation (DWC) can be described as a

phe-nomenon of incomplete wettability of a surface The wettability

of a surface is mostly responsible for the formation of a

cer-tain type of condensation and has a very strong effect on the

performance of the respective heat transfer process As firstly

discovered by Schmidt et al [1], the heat transfer coefficient for

DWC of steam can be up to one order of magnitude larger in

comparison with filmwise condensation (FWC) This can result

in a reduction of the condenser size and thus in a decrease of

capital costs Furthermore, the operating costs are also lower

due to the reduction of the pressure losses on both the cooling

and condensation side Although the conditions necessary for

promoting DWC have been well known for several decades and

experiments with coatings as promoters have been carried out

successfully, at least in part, the application of DWC is currently

The authors gratefully acknowledge the financial support for parts of this

work by the German National Science Foundation (DFG, Deutsche

Forschungs-gemeinschaft).

Address correspondence to Prof Dr.-Ing Andreas Paul Fr¨oba, Lehrstuhl f¨ur

Technische Thermodynamik, Universit¨at Erlangen-N¨urnberg, Am

Weichselgar-ten 8, D-91058 Erlangen, Germany E-mail: apf@ltt.uni-erlangen.de

still in a testing phase The reason for this is, on the one hand,that the implementation of condensers with DWC surfaces isconnected with a large financial expenditure, and on the otherhand, that long time stability of DWC has not been achievedwith most of the tested methods so far

Different methods were already examined to reduce the tability of the condenser surface by applying fatty acids or oils[2, 3] or coatings with low surface free energy materials likeorganics [4, 5] and polymers [6] At LTT-Erlangen, diamond-like carbon (DLC) coatings and direct modification of the metalsurfaces by ion implantation are studied [7] The latter method

wet-is considered to reduce the surface free energy of the metal andwas applied for the first time by Zhao et al [8, 9] using ionimplantation of N, Ar, He, H, and Cr in copper tubes In anotherwork by Choi [10], stable DWC could be generated on metallicsurfaces by an appropriately implemented ion beam implanta-tion process with ion doses of 1015 up to 1017 cm−2, using ni-trogen ions In the same work, measurements on ion-implantedcondenser plates resulted in condensation heat transfer coeffi-cients up to 17 times larger than those predicted by Nußelt’stheory for FWC Recent work points out that an enhancementfactor between 2 and 6 depending on the surface subcoolingand the condenser material seems to be a more realistic value[11, 12]

821

DWC on tube bundles was also examined by different authors

using the method of ion plating technology and ion beam

im-plantation A horizontal tube bundle condenser was constructed

in the year 1987 in Dalian Power Station by Zhao et al [9] to

maintain DWC using a combined method consisting of ion

plat-ing and ion beam treatment with Cr and N on the outer surface

of the tubes The condenser was operated for about 2 years,

after which some of the tubes at the steam entrance were

dam-aged due to a mechanical constructional defect [13] Zhao et al

[14] studied FWC and DWC of steam in vertical and

horizon-tal U-type condensers with a steam pressure of 1.2 to 1.8 bar

Each condenser contained 7 U-type 70Cu-Ni30 white copper

tubes with an outside diameter of 16 mm and a length of 1 m

treated with activated reactive-magnetron sputtering ion plating

of Cr+, N+, and C2H6 A small bundle of magnetron sputtering

ion plating-treated tubes was investigated by Burnside and Zhao

[15] The overall heat transfer coefficients for the treated tubes

were 62 to 81% larger in comparison with the untreated ones

The implementation of the method of plasma ion

implanta-tion represents a further development of previously investigated

surface treatment methods, particularly aiming at technical

ap-plications Plasma ion implantation is a pulsed process and

in-duced by surrounding the sample with plasma, which has been

created using a high voltage, and accelerating the cations in

the plasma onto the substrate by charging it negatively The

advantage of this technique is the possibility of achieving

si-multaneous implantation of the tube surface from all directions

more easily than by directed ion beam implantation In an earlier

work [16], stable DWC was achieved on plasma-ion-implanted

single horizontal stainless-steel tubes implanted with nitrogen

ion doses of 1015 and 1016cm−2 In this work, plasma ion

im-plantation is used for achieving stable DWC on condenser tube

bundles The increase in the heat flow and in the overall heat

transfer coefficient for the condensation of saturated steam on

these tubes is determined experimentally

WETTABILITY OF THE SURFACE AND ION

IMPLANTATION

In general, the wettability of a solid surface depends on the

interfacial tensions of the phase boundaries between solid and

liquid, solid and gas, and liquid and gas A liquid droplet on

a plane horizontal solid surface attains an equilibrium shape

characterized by the equilibrium contact angle
as shown in

Figure 1

From the surface free energy of a solid, which is equivalent to

the interfacial tension of the phase boundary between solid and

vacuum, it can be estimated whether a liquid wets a surface or

not When the surface free energy of a solid is below 40 mN.m−1,

it is relatively unwettable by water If its surface free energy is

larger than 60 mN.m−1, the surface will be wettable and a water

drop will spread with a low contact angle [17] The surface free

Figure 1 Equilibrium droplet and contact angle on a horizontal surface.energy γsurfis given by

γsurf = Usurf− T Ssurf (1)which can be decreased both by reducing the internal energy of

the surface Usurf and by increasing its entropy Ssurf A mental approach for explaining the effects of ion implantationconcerning the adjustment of DWC by Zhao and Burnside [13]suggests that implantation of foreign elements into the surface

funda-can increase Ssurf Furthermore, if the implanted elements havehigher energy, the bonding energy in the surface layers will

be decreased and hence Usurf will be reduced The higher theenergy of the elements implanted, the larger is the decrease inthe surface free energy of the implanted surface Several otherapproaches are provided by the same authors, all of them result-ing in a reduced surface free energy of the metal In contrast,recent experimental results with titanium surfaces show that thesurface free energy criterion often fails in predicting DWC In-stead, nucleation processes on places with locally altered surfacechemistry and induced microscopic surface roughness seem to

be another possible reason for the appearance of DWC [12, 18]

EXPERIMENTAL

Experiments have been executed to quantitatively describethe heat transfer enhancement caused by the adjustment of stableDWC of steam on ion implanted horizontal tube bundles Forthis, an experimental condenser has been constructed to measurethe heat flow and the overall heat transfer coefficient on bundleswith different numbers of tubes and tube arrangements Theexperimental apparatus consists of four main parts, namely, anelectric evaporator with automatic water supply, a condensertest cell, a cooling-water cycle, and a condensate collectionand recycling system A schematic diagram of the experimentalapparatus is shown in Figure 2

The tubular condenser used in this work can contain a bundle

of nine horizontal tubes (3 rows× 3 columns) The condensertubes have a length of 500 mm, an outer diameter of 20 mm, and

a wall thickness of 2 mm For multiple tube experiments, the tical and horizontal tube pitches are 60 mm Because of its highstability against corrosion, stainless steel X10CrNiMo18-9 (ma-terial no 1.4571, thermal conductivity 16.3 W.m−1.K−1) washeat transfer engineering vol 31 no 10 2010

ver-A B KANANEH ET AL 823

Figure 2 Experimental apparatus.

used as substrate material in the experiments The tubes were

mechanically polished with a 100-µm glass paper to remove

any impurities from their surfaces Before ion implantation and

before installing the tubes in the condenser, their surfaces were

cleaned in an ultrasonic bath with a special cleaner to remove

fat, oil, or any other impurities Afterward, the tubes were rinsed

with tap water, acetone, and finally with distilled water

DWC was adjusted on the same material by surface

modifica-tion with plasma ion implantamodifica-tion Nitrogen ions were selected

as doping elements with ion doses of 1015 and 1016 cm−2 at

an implantation energy of 20 keV This usually implies a

pene-tration depth in the range of 10 to 100 nm [17] Nitrogen ions

were chosen due to economical reasons They are generated

from gaseous nitrogen, which is cheap, easy to handle, and

non-toxic Furthermore, long time stability of DWC for more than a

year could be found by Choi [10] on stainless-steel plates

im-planted with N+ by ion-selective ion beam technology [17]

In the present work, experiments with different numbers of

stainless-steel tubes (three and nine tubes) and different tube

arrangements (horizontal row and vertical row) were carried out

with unimplanted and implanted tubes The unimplanted

bun-dle served as a check of the cleanliness of the condenser and

for comparative purposes A film covering the whole surface of

the unimplanted tubes indicates that no impurities are present

In this way it can be guaranteed that DWC on implanted tubes

is caused only by the applied surface modification Stability of

DWC on these tubes was studied, resulting in no observable

changes during the operation time of about 20 days Heat

trans-fer measurements at diftrans-ferent subcoolings at steam pressures

of 1,050, 1,500, and 2,000 mbar were accomplished and pared for DWC and FWC The variation of the subcooling wasachieved by varying the overall cooling-water flow rate Thecooling-water inlet temperature was about 20◦C The cooling-water outlet temperature was measured after mixing the singletube outlet flows For an approximately even distribution of theoverall cooling-water flow into the single tubes, an appropriatetubing system was installed

com-The heat flow on the cooling-water side ˙Qcan be determinedby

˙

Q = ˙mcwc p,cw (T cw,out − T cw,in) (2)where ˙mcwis the mass flow rate of the cooling water, c p,cwis its

mean specific heat capacity, and T cw,in and T cw,outare the inlet

and outlet cooling-water temperatures The heat flux ˙q is the heat flow per unit surface area Aoof the tubes The overall heat

transfer coefficient Hois determined by

Ho= ˙q

where TLMTD is the log mean temperature difference in the

condenser, which can be calculated by the steam temperature Tsand the cooling-water temperatures according to

A more detailed description of the apparatus, the experiment,

and the data evaluation procedure can be found elsewhere [19]

RESULTS AND DISCUSSION

DWC for Arrangements With Three Tubes

Two different arrangements of three tubes in the form of

a horizontal row and a vertical row were used to study the

phenomenon of DWC at a steam pressure of 1,050 mbar The

condenser was at first tested for FWC using unimplanted tubes

All of the unimplanted tubes showed stable FWC on the

com-plete surface DWC with tubes implanted with an ion dose of

1016 cm−2was obtained for both arrangements, the horizontal

and vertical row The results of the heat transfer measurements

for DWC are shown in Figure 3 in comparison with FWC The

heat flow ˙Q and the overall heat transfer coefficient Hoincrease

with increasing Reynolds number Re of the cooling-water flow

inside the tubes for both the implanted and unimplanted tubes

For the vertical row, ˙Q and Ho for the implanted tubes were

about 16% and 20% larger in comparison with the unimplanted

tubes For the horizontal row, ˙Qfor the implanted tubes was

Figure 3 Measured heat flow ˙Q and overall heat transfer coefficient Ho on

three tubes implanted with a nitrogen ion dose of 10 16 cm −2as a function of

Re for two different tube arrangements at a steam pressure of 1,050 mbar in

comparison with unimplanted tubes.

about 10% larger in comparison with unimplanted ones, while

Ho for the implanted tubes was about 11% larger The heattransfer values for horizontal row arrangement are larger thanfor the vertical row, as can be seen in Figure 3 The heat flowfor the implanted horizontal row was about 5 to 8% larger incomparison with the vertical row arrangement, while the overallheat transfer coefficient for the implanted horizontal row wasabout 6 to 9% larger The difference increases by increasing

Re for the cooling-water flow inside the tubes The horizontal

row behaves like single horizontal tubes because no condensatecomes from upper tubes and affects the condensation process,

in contrast to the arrangement in vertical row In the latter, thelower tubes receive condensate from the upper ones, which al-ters the heat transfer process because more condensate is present

on the lower second tube and even more on the third The

con-densation rate increases with Re due to the increased surface subcooling Thus at larger Re, the lowest tube is loaded with

more condensate and the condensation form is converted fromDWC into mixed condensation and into FWC on some parts ofthe tube The condensate film acts as a thermal resistance andhence reduces the heat flow and the overall heat transfer coeffi-

cient At small Re, the effect of the condensate coming from the

upper tubes on the heat transfer at the lowest tube is lower andthe difference in ˙Q and Hobetween the horizontal and verticalrow arrangements is decreased

DWC With Nine Tubes

Stable FWC was obtained on all of the nine unimplantedtubes installed in the 3 × 3 arrangement Before starting theheat transfer measurements, FWC was maintained over 2 days.Afterward, two different tube bundles implanted with nitrogenion doses of 1015and 1016cm−2were installed in the condenserfor studying DWC At low cooling-water flow rates, DWC wasformed on all of the tubes A photo taken from simultaneouslyvisible parts of the upper two tube rows with DWC on tubesimplanted with a nitrogen ion dose of 1016 cm−2 at a smallcooling-water Reynolds number of 2,244 is shown in Figure 4.The heat flow and the overall heat transfer coefficient as afunction of the cooling-water Reynolds number for different iondoses at a steam pressure of 1,050 mbar are presented in Fig-

ure 5 Both heat transfer values increase with increasing Re The

increase of ˙Q and Hois more pronounced for low values of Re.

Visual observations showed that for the tubes implanted with

1016 cm−2, on the upper row DWC, on the middle row mixedcondensation, and at the lowest row mixed and film condensa-

tion was formed at larger Re of 17,347 to 21,231 The formation

of FWC on some parts of the lowest tubes at larger Re was

induced by the condensate downward flow from the tubes ofthe upper two rows and the surface wettability not low enough

to maintain DWC at higher condensate loads By the formation

of mixed and film condensation, the thermal resistance of thecondensate increases and hence the heat flow and the overall

heat transfer coefficient are reduced As Re decreases, larger

heat transfer engineering vol 31 no 10 2010

A B KANANEH ET AL 825

Figure 4 DWC on the first two tube rows of a bundle of tubes implanted with

a nitrogen ion dose of 10 16 cm −2.

DWC zones were observed In this case, the reduced heat

trans-fer and hence smaller condensate flow rate ensure more stable

DWC, as flooding effects are less pronounced than for larger Re

numbers As a consequence of this behavior, the lower two rows

of the nine tubes were also completely covered with DWC at

small Re of 2,244 The effect of increased wetting of the lower

Figure 5 Measured ˙Q and Ho on a bundle of nine implanted tubes as a

function of Re for two different ion doses at a steam pressure of 1,050 mbar in

comparison with unimplanted tubes.

tubes was stronger for the tubes implanted with an ion dose of

1015cm−2 This can be attributed to a reduced efficiency of theion implantation for inducing DWC at smaller ion doses As aresult, the heat transfer values for these tubes are smaller thanfor those with an ion dose of 1016 cm−2at the same Re num-

bers The enhancement ratio εHo, which is the ratio between theoverall heat transfer coefficient for implanted tubes and that of

unimplanted ones at constant Re, was found to be 1.12 for tubes

implanted with a nitrogen ion dose of 1016cm−2 at a Re value

of 21,231 This means that Hofor tubes implanted with an iondose of 1016cm−2was increased by 12%; i.e., the heat transferarea can be decreased by 12% for achieving the same ˙Qas withunimplanted tubes

The influence of the steam pressure on the heat flow andthe overall heat transfer coefficient for DWC on the bundle ofnine tubes implanted with a nitrogen ion dose of 1016 cm−2 isshown in Figure 6 The heat flow and the overall heat transfercoefficient increase with increasing steam pressure for constant

Re This behavior is caused by the decrease of the interfacial

resistance to mass transfer at the liquid–vapor interface with

increasing pressure [20] At larger Re and higher pressures,

more condensate was formed on the upper tubes Consequently,the lowest row receives more condensate, which graduallyconverted the condensation form to approximately FWC at

Figure 6 Measured ˙Q and Ho on a bundle of nine tubes implanted with a nitrogen ion dose of 10 16 cm −2as a function of Re for different steam pressures.

heat transfer engineering vol 31 no 10 2010

Figure 7 Measured ˙Q and Ho for different numbers and arrangements of

tubes implanted with a nitrogen ion dose of 10 16 cm −2as a function of Re at a

steam pressure of 1,050 mbar.

a steam pressure of 2,000 mbar and a Re value of 22,656.

As a result, the positive effect of increasing steam pressure

on Ho is reduced, especially at larger Re The maximum Ho

of 2.22 kW·m−2·K−1 was achieved at a steam pressure of

2,000 mbar and Re number of 22,656.

Comparison Between the Different Tube Arrangements

In the following, the heat transfer measurements for the

dif-ferent numbers of tubes and arrangements installed inside the

condenser are compared For different numbers of tubes

im-planted with a nitrogen ion dose of 1016 cm−2, the measured

heat flow ˙Q and the overall heat transfer coefficient Ho are

shown in Figure 7 as a function of Re at a steam pressure of

1,050 mbar

The heat flow increases as the number of tubes increases for

constant Re, because the heat transfer area Ao increases, and

hence the condensate mass flow increases The increase in the

heat flow is in a nonlinear way proportional to the number of

tubes In the case of multiple tubes, the cooling-water flow rate

is distributed among the tubes Furthermore, the effect can be

attributed to the influence of the condensate on the lower tubes,

especially for three tubes in vertical row and for the nine tubes.The effect of the number of tubes on the overall heat transfer

coefficient Hois contrary to the effect on ˙Q, as can also be seen

in Figure 7 For constant Re, Hodecreases as the number of tubesand consequently the condensation surface area increases; see

Eq 3

CONCLUSIONS

Heat transfer measurements were performed on implanted horizontal tube bundles with different arrangements

plasma-ion-of three and nine stainless-steel tubes Ideal FWC was formed

on all of the well-cleaned unimplanted tubes For three tubes

in a horizontal row arrangement, the heat flow and the overallheat transfer coefficient were larger than for three tubes in avertical row because the lower tubes in a vertical row receivedcondensate from the upper ones The increased amount of con-densate on the lower tubes results in an increase of the thermalresistance, which decreases the heat flow and the overall heattransfer coefficient

For a bundle of nine tubes implanted with nitrogen ion doses

of 1015 and 1016 cm−2, stable DWC was achieved at a smallReynolds number of about 2,200 At larger Reynolds numbersbetween 17,000 and 22,000, DWC was formed on the uppertubes, mixed condensation on the middle row, and mixed andfilm condensation on the lowest row The formation of mixedand filmwise condensation increased the thermal resistance onthe condensation side, reducing the overall heat transfer coeffi-cient The enhancement ratio εHowas found to be 1.12 for tubesimplanted with a nitrogen ion dose of 1016cm−2 at a Reynoldsnumber of about 21,000 The heat flow and the overall heat trans-fer coefficient were increased by increasing the steam pressure

NOMENCLATURE

Ao outside heat transfer area, m2

cp specific heat capacity, J.kg−1.K−1

Ho overall heat transfer coefficient, W.m−2.K−1

TLMTD log mean temperature difference, K

Usurf surface internal energy, N.m−1

[1] Schmidt, E., Schurig, W., and Sellschopp, W., Versuche ¨uber die

Kondensation von Wasserdampf in Film- und Tropfenform, Tech.

Mech Thermodyn., vol 1, no 2, pp 53–63, 1930.

[2] Blackman, L C F., Dewar, M J S., and Hampson, H., An

Inves-tigation of Compounds Promoting the Dropwise Condensation of

Steam, Applied Chemistry, vol 7, pp 160–171, 1957.

[3] Watson, R G H., Birt, D C P., Honour, C W., and Ash, B

W., The Promotion of Dropwise Condensation by Montan Wax

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539–546, 1962

[4] Das, A K., Kilty, H P., Marto, P J., Andeen, G B., and Kumar,

A., The Use of an Organic Self-Assembled Monolayer Coating to

Promote Dropwise Condensation of Steam on Horizontal Tubes,

Journal of Heat Transfer, vol 122, pp 278–286, 2000.

[5] Vemuri, S., and Kim, K J., An Experimental and Theoretical

Study on the Concept of Dropwise Condensation, International

Journal of Heat and Mass Transfer, vol 49, no 3–4, pp 649–657,

2006

[6] Marto, P J., Looney, D J., Rose, J W., and Wanniarachchi, A S.,

Evaluation of Organic Coatings for the Promotion of Dropwise

Condensation of Steam, International Journal of Heat and Mass

Transfer, vol 29, no 8, pp 1109–1117, 1986.

[7] Leipertz, A., and Fr¨oba, A P., Improvement of Condensation Heat

Transfer by Surface Modification, Heat Transfer Engineering, vol.

29, no 4, pp 343–356, 2008

[8] Zhao, Q., Zhang, D., and Lin, J., Surface Materials With

Drop-wise Condensation Made by Ion Implantation Technology,

Inter-national Journal of Heat and Mass Transfer, vol 34, no 11, pp.

2833–2835, 1991

[9] Zhao, Q., Zhang, D., Zhu, X., Xu, D., Lin, Z., and Lin, J.,

Indus-trial Application of Dropwise Condensation, Proc 9th

Interna-tional Heat Transfer Conference, Jerusalem, vol 4, pp 391–394,

1990

[10] Choi, K.-H., Gezielte Einstellung und w¨armetechnische

Charak-terisierung der Tropfenkondensation auf ionenimplantierten

Oberfl¨achen, Dr.-Ing Thesis, Friedrich-Alexander-Universit¨at

Erlangen-N¨urnberg, Germany, 2001

[11] Rausch, M H., Fr¨oba, A P., and Leipertz, A., Dropwise

Con-densation Heat Transfer on Ion Implanted Aluminum Surfaces,

International Journal of Heat and Mass Transfer, vol 51, no.

5–6, pp 1061–1070, 2008

[12] Rausch, M H., Leipertz, A., and Fr¨oba, A P., Dropwise

Conden-sation of Steam on Ion Implanted Titanium Surfaces, International

Journal of Heat and Mass Transfer, vol 53, no 1–3, pp 423–430,

2010

[13] Zhao, Q., and Burnside, B M., Dropwise Condensation of Steam

on Ion Implanted Condenser Surfaces, Heat Recovery Systems &

CHP, vol 14, no 5, pp 525–534, 1994.

[14] Zhao, Q., Liu, J J., Bai, T., Lin, J., Cui, B Y., Shen, J L., andFang, N T., Dropwise Condensation of Steam on Vertical and

Horizontal U-Type Tube Condensers, Proc 10th International

Heat Transfer Conference, Brighton, pp 117–121, 1994.

[15] Burnside, B M., and Zhao, Q., Dropwise Condensation of Steam

at High Velocity and Vacuum Pressures Over a Small Tube

Bank, Proc Eurotherm Seminar, Paris, vol 27, pp 196–204,

1995

[16] Bani Kananeh, A., Rausch, M H., Fr¨oba, A P., and Leipertz,A., Experimental Study of Dropwise Condensation on Plasma-

Ion Implanted Stainless Steel Tubes, International Journal of

Heat and Mass Transfer, vol 49, no 25–26, pp 5018–5026,

2006

[17] Roth, J R., Industrial Plasma Engineering Volume 2: Applications

to Nonthermal Plasma Processing, IOP Publishing Ltd., London,

2001

[18] Rausch, M H., Leipertz, A., and Fr¨oba, A P., On the Origin

of Dropwise Condensation of Steam on Ion Implanted

Metal-lic Surfaces, Proc 20th International Symposium on Transport

Properties, Victoria, BC, paper 70, 2009.

[19] Bani Kananeh, A., Experimental Study of Dropwise Condensation

on Ion Implanted Horizontal Single Tubes and Tube Bundles, Ing Thesis, Friedrich-Alexander-Universit¨at Erlangen-N¨urnberg,Germany, 2005

Dr.-[20] Rose, J W., Dropwise Condensation, in Heat Exchanger Design

Handbook 1998, ed G F Hewitt, pp 2.6.5-1–2.6.5-11, Begell

House, New York, 1998

Ali Bani Kananeh is a product manager at GEA

Ecoflex GmbH, which is part of the GEA Process Equipment Division, and where he is working on condensation, evaporation, and fouling inside plate heat exchangers He was awarded an M.Sc in chem- ical engineering at the Jordan University of Science and Technology (JUST) in Irbid in 1997 and worked afterward as a sales engineer for the Ideal Group for Water Treatment Company, Amman, Jordan From

1999 until 2005, he was a Ph.D student at the partment of Engineering Thermodynamics at the Institute of Chemical and Bio- engineering of the University of Erlangen-Nuremberg, where he was awarded

De-a Dr.-Ing in 2005 His reseDe-arch interests include modificDe-ation of heDe-at trDe-ans- fer surfaces for condensation and evaporation applications, fouling and cor- rosion inside plate heat exchangers, and crystallization of cretin salts from solutions.

trans-Michael Heinrich Rausch is a Ph.D student at

the Department of Engineering Thermodynamics at the Institute of Chemical and Bioengineering of the University of Erlangen-Nuremberg He received his diploma in chemical engineering in Erlangen in 2003 and is currently studying dropwise condensation heat transfer on modified metallic surfaces, as well as the changes in metal surface characteristics induced by ion implantation and their effects on the condensation form for various working fluids.

heat transfer engineering vol 31 no 10 2010

Alfred Leipertz is head of the Department of

Engi-neering Thermodynamics at the Institute of Chemical and Bioengineering of the University of Erlangen- Nuremberg and coordinator of the Erlangen Graduate School in Advanced Optical Technologies (SAOT).

He was awarded a diploma in physics from the versity of Gießen in 1974, and a Dr.-Ing and a Dr.-

Uni-Sc habil in heat and mass transfer at the School of Mechanical Engineering of the University of Bochum

in 1979 and 1984, respectively His research work covers a wide range of topics related to thermodynamics, fluid dynamics, heat

and mass transfer, and particle and combustion technology In several of these

topics he has contributed significantly by the development and application of

new laser diagnostic techniques He is the author of more than 600 publications,

more than 200 of which have been published in international peer-reviewed

journals In 2000 he received the Arch T Colwell Merit Award of the Society

of Automotive Engineers (SAE) He is an elected member of the Subcommittee

on Transport Properties of the Commission I.2 (Thermodynamics) of the

Inter-national Union of Pure and Applied Chemistry (IUPAC), of the InterInter-national

Association for Transport Properties (IATP), of the Scientific Working Group

for Technical Thermodynamics (WATT e.V.), and a Fellow of IUPAC, of the

Optical Society of America, and of the SAE He is also a member of the

ed-itorial board of the Internet journal “diffusion-fundamentals”

(www.diffusion-fundamentals.org) and of the peer review boards of more than 30 scientific journals.

Andreas Paul Fr¨oba occupies a tenure-track

posi-tion at the junior professor level established in the framework of the Erlangen Graduate School in Ad- vanced Optical Technologies (SAOT) at the Univer- sity of Erlangen-Nuremberg, since the beginning of

2008 Until then, he was an assistant professor and head of the group Heat and Energy Engineering & Thermophysical Property Research at the Depart- ment of Engineering Thermodynamics at the Institute

of Chemical and Bioengineering of the University of Erlangen-Nuremberg, where he was awarded a Dr.-Ing in 2002 and a Dr.-

Sc habil in engineering thermodynamics in 2009 He received his diploma in physics at the Department of Technical Physics at the Institute of Physics of Condensed Matter of the University of Erlangen-Nuremberg in 1997 He has

a strong expertise in the determination of thermophysical properties of fluids

by dynamic light scattering (DLS) Besides thermophysical property research for working fluids in chemical and energy engineering, his current interests in- volve condensation heat transfer and seawater desalination by mechanical vapor compression.

heat transfer engineering vol 31 no 10 2010

Copyright
C Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903547560

Experimental Investigation of the

Effect of Surface Inclination Angle

on Saturated Pool Film Boiling Heat Transfer in Transient Regime

ABDURRAHIM BOLUKBASI and DOGAN CILOGLU

Department of Mechanical Engineering, Ataturk University, Erzurum, Turkey

In order to examine the effect of surface inclination angle on saturated pool film boiling heat transfer in transient regime,

an experimental study was carried out The experiments were performed through a cylindrical rod, made up of brass

20 mm in diameter and 75 mm in length, placed at six inclination angles about the vertical (from 0 to 50◦) under atmospheric

pressure The test specimen heated at high temperatures was immersed in a distilled water pool at saturated condition.

Temperature of the specimen during the cooling process was recorded using a K-type thermocouple embedded at the

center of the specimen In the experiments, the pool film boiling was observed for each inclination angle In the film

boiling region, the heat transfer coefficients were calculated by means of a lumped parameter method The experimental

results showed that the heat transfer coefficient increased as the inclination angle increased In addition, to predict the

Nusselt number, an empirical formula including the inclination angle as well as the Grashof, Prandtl, and Jakob numbers

was developed, and good agreement between the predicted and experimental data for the vapor Nusselt numbers was

observed.

INTRODUCTION

The boiling heat transfer mechanism is an important problem

for the heat exchangers and the heat removal systems

Engi-neering approaches such as heat treatments of steel, cooling

applications of cryogenic systems, cooling of the

supercon-ducting magnets, core safety of light water reactors, and the

rapid solidification processing are some application fields of

the pool boiling heat transfer In the applications, more

ef-fective and enhanced heat transfer with high efficiency and

without unexpected accidents due to high heat flux from the

superheated solid surface to the coolant liquid is desired by

the engineers One of the most effective parameters for the

enhanced heat transfer rate is the inclination angle (θ) of the

heated surface The effects of several design parameters on

This research was supported by the BAP-2002/38 project of the Research

Fund of Ataturk University and performed in the laboratories of the Mechanical

Engineering Department, Engineering Faculty, Ataturk University.

Address correspondence to Dogan Ciloglu, Department of Mechanical

En-gineering, Ataturk University, 25240, Erzurum, Turkey E-mail: dciloglu@

m−2 with surface orientations varying from horizontally ing upward (0◦), to vertical (90◦), to horizontally facing down-ward (180◦) They reported that the resulting nucleate boilingcurves revealed considerable boiling hysteresis and enhancedsurfaces showed two to three times better heat transfer than

fac-a plfac-ain surffac-ace In film boiling, enhfac-anced surffac-aces fac-also ited better heat transfer characteristics and the role of surface829

exhib-orientation on the motion and stability of the vapor film was

declared

The effects of tube inclination angles on nucleate pool boiling

heat transfer of water at atmospheric pressure were

experimen-tally investigated by Kang [3] The experiments were performed

for seven angles (0, 15, 30, 45, 60, 75, and 90◦) with two tube

diameters (12.7 and 19.1 mm) of 540 mm in length It was

reported that the inclination angle produced much change in

heat transfer coefficients Also, when a tube was placed near

the horizontal and the vertical position, the maximum and the

minimum heat transfer coefficients were obtained, respectively,

and the maximum values were found about five to seven times

greater than the minimum values depending on the tube

diam-eter and the wall superheat According to these experimental

results, the heat transfer coefficient decreased as the tube

diam-eter increased, except for some inclination angles (15 and 30◦),

which were affected by the strong liquid agitation To

deter-mine effects of the tube inclination angle on pool boiling heat

transfer, an experimental study of a tubular heat exchanger was

carried out by Kang [4] It was described that tube inclination

caused much change on pool boiling heat transfer, and the

ef-fect of the inclination angle was more strongly observed in the

smooth tube In addition, if a tube was properly inclined,

en-hanced heat transfer was expected due to the decrease in bubble

slug formation on the tube surface and liquid easily accessed the

surface

The effects of the diameter and the orientation of electrically

heated wires on their critical heat flux for both in saturated

pool boiling and in surface boiling were investigated by Stralen

and Sluyter [5] It was defined that the horizontal type was

more effective than the vertical one in both natural convection

and boiling regions The peak heat flux for 0◦was 45% higher

than that for 90◦ It was also reported that the major reason for

the reduction in heat transfer for the vertical position was the

formation of large vapor slugs

The pool film boiling heat transfer for the superheated

ver-tical cylindrical specimens with different diameters and

dif-ferent lengths in case of cooling in the water pool with ity conditions was experimentally investigated by Bolukbasiand Ciloglu [6] They found that the important parameter interms of the boiling heat transfer was the characteristic length.Although the specimen’s diameters and lengths differ consid-erably, it was declared that the specimens having the samecharacteristic lengths exhibited the same heat transfers perfor-mance

grav-Although there are a few studies related to the effect of face inclination angle on the nucleate pool boiling heat transfer

sur-in the open literature, the effects on pool film boilsur-ing heat fer have not been investigated for inclined cylinders in saturatedfilm boiling conditions Therefore, the present study aims to de-termine the effects of surface inclination angle on saturated poolfilm boiling heat transfer in a transient regime using a cylindricalrod placed at different inclination angles

trans-POOL BOILING EXPERIMENTS Experimental Setup

The experimental system is shown schematically in Figure 1

It consists of a liquid tank, a furnace, a pneumatic piston, and

a measuring and control unit The liquid tank, made of nized steel (8 mm in thickness), is cylindrical in shape and has

galva-a digalva-ameter of 400 mm galva-and galva-a height of 700 mm The furngalva-ace(diameter 400 mm × 200 mm) is capable of raising the tem-perature up to 1,000◦C, and the outer surface of the furnace isinsulated to prevent the heat loss The test specimen inserted

on the support rod is movable up and down by using matic piston To measure the variation of temperature at thecenter region of the test specimen, a K-type thermocouple wasused Additionally, two K-type thermocouples (located 150 and

pneu-300 mm above of the liquid tank base) were used for suring the boiling liquid temperature The thermocouples wereinterfaced into a data acquisition unit integrated into a personal

mea-Figure 1 The experimental system.

heat transfer engineering vol 31 no 10 2010

A BOLUKBASI AND D CILOGLU 831

Figure 2 The test specimen.

computer (PC) via an Advantech ISA PCL-818HG board A

PCLD-8115 32-channel thermocouple amplifier was used for

temperature measurements during the experiments, and Genie

software was used for data acquisition and system

configura-tion The boiling liquid was heated by a thermostat-controlled

electrical heater with a maximum power of 2 kW In order to

check the water level and to observe boiling phenomena, two

observation windows (60× 150 mm) were installed facing one

another 140 mm above the tank base To ensure gas and liquid

inlet–outlet, the valves were used at the surface and the bottom

of the tank as seen in Figure 1

The selected test specimen (diameter 20× 75 mm) is shown

in Figure 2 In order to prevent film collapse, the bottom of the

cylinder was shaped as semi-spherical A hole was drilled to

the center of cylindrical rod to measure the temperature on the

center of the specimen as seen in Figure 2

Experimental Procedure

The test specimen was heated up to 600◦C in the furnace

under the atmospheric pressure In order to prevent oxidation,

the nitrogen gas was injected into the furnace during the heating

process When the specimen temperature reached 600◦C, it was

suddenly immersed into distilled water pool in saturated

con-dition The saturated condition was 92◦C corresponding to the

atmospheric pressure conditions in Erzurum city at an altitude

of 1,850 m Then the center temperature and the cooling time

were recorded during the cooling process For this

investiga-tion, the inclination angles were selected as 0, 10, 20, 30, 40,

and 50◦ The maximum angle value was limited to 50odue to

the geometrical limitations of the experimental system To

min-imize the probability of causing the aging process, the specimen

surface was polished with emery papers (1200 mesh) by

remov-ing the contaminations formremov-ing on the test surface after each

testing, and the surface roughness values were in the range of

0 150 300 450 600

Film boiling

Nucleate boiling

Natural convection

Figure 3 The variations of the center temperature with the cooling time for various surface inclination angles.

0.05 and 0.10 µm For these measurements, a surface roughnessprofilometer with TR 200 trademark was used The experimentswere carried out three times for each inclination angle Whenthe variations of the center temperature with the cooling timeobtained from ternary experiments were investigated, they havethe same slope in film boiling regions If this state was provided,one of the ternary tests was taken into consideration Thereafter,the variations of the measured center temperature with the cool-ing time for selected surface inclination angles were drawn asseen in Figure 3

Calculation Procedure and Numerical Analysis

In a heat transfer analysis, the temperature of the heated face (Ts), the heat transfer coefficient (h), and the heat flux (q)from the specimen surface to the water are important param-eters In this investigation, the heat transfer from the surface

sur-to water occurs by convection and radiation Thereby, the heattransfer coefficient obtained from the experiments includes bothconvective and radiative heat transfer In order to develop em-pirical equations from the experimental results with the aim

of calculating the heat transfer coefficient, the time-dependentheat transfer analysis was performed at two stages At the firststage, the heat transfer coefficient was estimated by the Lumpedmethod The equation for Lumped analysis can be shown asfollows:

−hA(T − Tsat)= ρVcdT

The important parameter for validation of this analysis is theBiot number (Bi), which compares the contribution of internalconductive resistance to the overall heat transfer in the systemrelative to that of the convection on the specimen surface A

small Biot number (<0.1) represents negligible internal

resis-tance, such that there is a larger amount of heat transfer takingplace by conduction than that by convection This means thatinternal temperature gradients can be negligible Therefore, thecenter temperature and surface temperature of the specimen canheat transfer engineering vol 31 no 10 2010

Table 1 Regression statistics for Eq (1)

be approximately equal The Biot number of the test specimen

used in the experiments was 0.011 Since this value is much

smaller than 0.1, the heat transfer analysis can be performed

by the Lumped method The analysis was therefore expected to

yield an error less than 5% [7]

In Eq (1), the heat transfer coefficient varies with

tempera-ture Since how the heat transfer coefficient changes with

tem-perature is not known, Eq (1) cannot be solved analytically

In order to determine the relationship between them, first, the

variations of the center temperature with the cooling time

ob-tained from the experiments for various inclination angles were

drawn Then, to determine the relationship between the center

temperature and the cooling time in pool film boiling region, an

empirical model in Eq (2) was used:

The regression constants B and C and coefficient of

determina-tion values for Eq (2) were numerically calculated by

Statis-tica Software [8] using quasi-Newton curve fitting method [9]

These regression statistics and the initial temperatures can be

seen in Table 1 This table shows that empirical models for all

inclination angles represent the experimental results very well

Substituting the derivative of Eq (2) into Eq (1), with the

as-sumption T= Tc, the approximate solution for the heat transfer

coefficient can be found At the end of the first stage, the

em-pirical relationship in Eq (3) between the temperature and heat

transfer coefficient was developed by Statistica Software:

The h value is a function of the surface temperature Therefore,

the h values calculated from Eq (3) include some error due

to the assumption of T= Tc In additon, there is some

differ-ence between the center and surface temperature In the boiling

heat transfer mechanism, this temperature difference has a large

effect on the heat transfer coefficient Thus, the heat transfer

coefficient must be determined according to the surface

temper-ature To determine the surface temperature of the specimen (Ts),

the second stage of the analysis was performed At this stage,

the time-dependent and spatially dependent temperature

distri-butions in the specimen were calculated by the finite-element

method using FEMLAB software [10] In this calculation, the

initial and boundary conditions were required The initial

tem-perature of specimen was taken as the initial condition The

Table 2 Regression constants for Eq (3)

as the convergence criterion The convergence criterion and timeinterval were selected as 0.01◦C and 1 s, respectively At theend of the second stage, the h values were calculated usingthe surface temperatures where the convergence criterion wasreached Regression constants D and E, obtained from Eq (3),are given in Table 2 As seen in Table 2, only the constant

D varies with the surface inclination angle The increasing Dvalues indicate an enhancement in heat transfer Consideringthe h values calculated by using the surface temperatures, theheat flux was then calculated by the following equation:

Engineers and scientists prefer analyzing heat transfer problems

as a function of certain dimensionless numbers The most portant dimensionless number is the Nusselt number (Nu) for aconvection problem For the test specimen, the Nusselt numbercan be calculated by the following equation:

where Lsis the characteristic length and is defined as the ratio ofthe specimen volume to the specimen surface area [11] The Lsvalue calculated for the test specimen was 4.8 mm The Nusseltnumber is related to the Grashof (Grv), Prandtl (Prv), and Jakob(Jav) numbers, which are important dimensionless numbers forthe pool boiling In order to predict the Nusselt number, thefollowing empirical formula including the inclination angle aswell as the Grashof, Prandtl, and Jakob numbers was developed:

Nuv= C1



GrvPrv

Jav(1+ θ)

C2

(7)

where θ is the inclination angle about the vertical in radian form,and C1and C2are regression constants calculated by StatisticaSoftware In Eq (7), the regression constants are 0.0018 and0.6, respectively

heat transfer engineering vol 31 no 10 2010

A BOLUKBASI AND D CILOGLU 833

Table 3 Uncertainties in the measured and calculated parameters

The uncertainties of the measurement parameters were

ana-lyzed by applying the general theory of error propagation [12]

The relative errors of the parameters used in Eq (1) were listed

in Table 3 The errors of other standard parameters accepted in

the literature are ignored [13] As a result, the maximum relative

errors of the heat transfer coefficient, the heat flux, and the

Nus-selt number were found to be±3.23%, ±3.24%, and ±4.47%,

respectively

RESULTS AND DISCUSSION

The cooling curves obtained from the experiments are

il-lustrated in Figure 3 As seen in this figure, the saturated film

boiling preventing the heat transfer between the specimen

sur-face and the water was observed for all inclination angles As can

be seen in Figure 3, the variations of the center temperature with

the cooling time in film boiling region were almost linear In

nu-cleate boiling, the variations of the temperature with the cooling

time were more than that of film boiling and natural convection

As was expected, the increasing θ values increased the cooling

capability of the specimen For engineering applications, the

assessment of the time-dependent heat transfer coefficient (h),

the heat flux (q), and the Nusselt number (Nuv) values with the

surface superheat known as the difference between the surface

temperature (Ts) and the saturated water temperature (Tsat) is

Figure 4 The variations of the heat transfer coefficient with the surface

su-perheat for various surface inclination angles in pool film boiling.

25 35 45 55 65 75

0° 10° 20° 30° 40° 50°

Figure 5 The variations of the heat flux with the surface superheat for various surface inclination angles in pool film boiling.

useful For the saturated water temperature of 92◦C, h and q

variations with the surface superheat (Ts– Tsat) during pool filmboiling are illustrated in Figures 4 and 5, respectively

In Figure 4, it can be clearly seen that the heat transfer ficient increased with increasing the surface inclination angle,since the curves shifted to an upper level However, the shift ten-dencies of curves to an upper level were not proportional to theinclination angle A regular shift of the curve was observed onlyfrom 0 to 20◦ inclination angles When Figure 4 was carefullyinvestigated, two considerable shifts of the curve were observed

coef-at inclincoef-ation angles of 30 and 50◦, respectively It was thoughtthat the heat transfer coefficient would be affected by the liq-uid agitation in these inclination angles Also, the differencesamong heat transfer coefficients for various inclination angleswere found to be increasing with decreasing surface superheat.For example, with 50◦ inclination angle, the enhancement ob-served in the heat transfer coefficient was about 27% and 25%for low and high surface superheat, respectively The compo-nents of the buoyant force perpendicular to the upper surface ofthe test specimen increase with the increase of the inclinationangle The increase of this force increases the rate of detach-

Predicted Nuv 10

15 20 25 30 35 40 45

0° 10° 20° 30° 40° 50°

Table 4 The absolute and relative errors in the predicted Nu v values

ments of the vapor layer from the upper surface of the specimen

This leads to smaller average vapor film thickness around the

specimen and higher heat transfer coefficient and higher Nusselt

number As a result, one of the possible reasons for enhanced

heat transfer can be explained by the decrease of the average

vapor film thickness around the specimen with the increase

of the inclination angle It was also concluded that the vapor

film thickness increased with the increase of the surface

super-heat during pool film boiling Therefore, the h values decreased

while the surface superheat increased, as seen in Figure 4

In addition, the heat transfer coefficient obtained from the

ex-periments includes both convective and radiative heat transfer

The radiative heat transfer portion of the total heat transfer was

about 7% and 34% for low and high surface superheat,

respec-tively Also, the effect of radiative heat transfer increased with

the increase of the surface temperature Figure 5 shows that q

values increased with increasing the surface superheat

Increas-ing of q values against decreasing of h values occurred due

to the increase of the surface superheat Figure 6 presents the

comparison of the Nuvnumbers predicted from Eq (7) and

ob-tained from experimental studies for different inclination angles

from 0◦ to 50◦ in the film boiling region The error band for

this subset of data is found to be+2.78 to –11.05% As seen

in Table 4, the absolute and relative errors of the Nuv values

predicted from Eq (7) are less at small angles than those at high

angles

CONCLUSIONS

An experimental study was performed to investigate the

ef-fect of inclination angles on the saturated pool film boiling

heat transfer in transient regime The experiments were carried

out under the atmospheric pressure by using the surface of a

cylindrical rod placed at different inclinations (from 0 to 50◦)

When the test specimen was heated to high temperatures, it was

immersed in a distilled water pool at saturated condition The

temperature of the specimen during the cooling process was

recorded using a K-type thermocouple The heat transfer

coeffi-cients were calculated by means of a lumped parameter method

Based on the experiments performed in the study, the following

conclusions can be drawn:

1 The film boiling was observed at each inclination angle from

0 (vertical) to 50◦

2 The heat transfer coefficient, heat flux, and Nusselt numberincreased with the increase of the surface inclination angle.However, this enhancement was not proportional to the in-clination angle For example, two considerable shifts of thecurve were observed at inclination angles of 30 and 50◦,respectively

3 The enhancement in the heat transfer coefficient was about27% and 25% for low and high surface superheat, respec-tively

4 In order to predict the Nusselt number in saturated pool filmboiling, an empirical formula including the inclination anglewas developed It was concluded that there was a consider-ably good agreement between the predicted and experimentaldata for the Nuvnumbers

Further studies are needed in order to determine the combinedeffect of both the characteristic length and the inclination angle

in cylindrical geometries on pool film boiling heat transfer

h heat transfer coefficient, W m−2K−1

hfg latent heat of vaporization, J kg−1

Jav Jakob number, cp(Ts−Tsat)/hfg

A BOLUKBASI AND D CILOGLU 835

[1] Chun, M H., and Kang, M G., Effects of Heat Exchanger Tube

Parameters on Nucleate Pool Boiling Heat Transfer, ASME

Jour-nal of Heat Transfer, vol 120, pp 468–476, 1998.

[2] Jung, D S., Venant, J E S., and Sousa, A C M., Effects of

Enhanced Surfaces and Surface Orientations on Nucleate and Film

Boiling Heat Transfer in R-11, International Journal of Heat and

Mass Transfer, vol 30, pp 2627–2639, 1987.

[3] Kang, M G., Effect of Tube Inclination On Pool Boiling Heat

Transfer, Nuclear Engineering and Design, vol 220, pp 67–81,

2003

[4] Kang, M G., Effect of Tube Inclination on Pool Boiling Heat

Transfer, ASME Journal of Heat Transfer, vol 122, pp 188–192,

2000

[5] Stralen, S J D., and Sluyter, W M., Investigations on the Critical

Heat Flux of Pure Liquids and Mixtures Under Various

Condi-tions, International Journal of Heat and Mass Transfer, vol 12,

pp 1353–1384, 1969

[6] Bolukbasi, A., and Ciloglu, D., Investigation of Heat Transfer by

Means of Pool Film Boiling on Vertical Cylinders in Gravity, Heat

and Mass Transfer, vol 44, pp 141–148, 2007.

[7] Incropera, F P., and DeWitt, D P., Fundamentals of Heat and

Mass Transfer, 4th ed., John Wiley and Sons, New York, 1996.

[8] Statistica for Windows, Release 6.0, StatSoft, Inc., Tulsa, OK,USA, 2003

[9] Fletcher, R., Practical Methods of Optimization, Wiley, New York,

1987

[10] COMSOL AB, FEMLAB Version 3.1 pre, Reference Manual,January 2005

[11] Bayazitoglu, Y., and Ozisik, M N., Elements of Heat Transfer,

McGraw-Hill, New York, 1988

[12] Buchanan, J L., and Turner, P R., Numerical Methods and

Anal-ysis, McGraw-Hill, New York, 1992.

[13] Yesilata, B., A Simple Experimental Method for Determining

Natural Convection Heat Transfer Coefficient in Liquids,

Ter-modinamik, vol 146, pp 94–102, 2004.

Abdurrahim Bolukbasi is a faculty member of the

Mechanical Engineering Department in Ataturk versity, Erzurum, Turkey He received his Ph.D in

Uni-1997 on film boiling heat transfer His remain search interests are boiling heat transfer, nanofluid heat transfer, computational fluid dynamics (CFD), and numerical heat transfer He has authored 23 con- ference and journal publications.

re-Dogan Ciloglu is a Ph.D student in the

Depart-ment of Mechanical Engineering, Ataturk University, Erzurum, Turkey He received his master’s degree from Ataturk University in 2004, and his bachelor’s degree in mechanical engineering from the same uni- versity in 2000 Currently, he is working on pool boiling heat transfer in nanofluids.

heat transfer engineering vol 31 no 10 2010

Copyright
Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903547602

Review of Improvements on

Shell-and-Tube Heat Exchangers With Helical Baffles

QIUWANG WANG, GUIDONG CHEN, QIUYANG CHEN, and MIN ZENG

School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an, China

Helical baffles are employed increasingly in shell-and-tube heat exchangers (helixchangers) for their significant advantages

in reducing pressure drop, vibration, and fouling while maintaining a higher heat transfer performance In order to make

good use of helical baffles, serial improvements have been made by many researchers In this paper, a general review is

provided of developments and improvements on helixchangers, which includes the discontinuous helical baffles, continuous

or combined helical baffles, and the combined multiple shell-pass helixchangers Extensive results from experiments and

numerical simulations indicate that these helixchangers have better flow and heat transfer performance than the conventional

segmental baffled heat exchangers Based on these new improvements, the conventional heat exchangers with segmental baffles

might be replaced by helixchangers in industrial applications to save energy, reduce cost, and prolong the service life and

operation time.

INTRODUCTION

Heat exchangers play an important role in many engineering

processes such as oil refining, chemical industry, environmental

protection, electric power generation, refrigeration, and so on

Among different types of heat exchangers, the shell-and-tube

heat exchangers (STHXs) have been commonly used in

indus-tries [1] It was reported that more than 35–40% of the heat

ex-changers are of the shell-and-tube type, because of their robust

construction geometry as well as easy maintenance and

possi-ble upgrades [2, 3] In order to meet the special requirements

of modern industries, various ways are adopted to enhance the

heat transfer performance while maintaining a reasonable

pres-sure drop for the STHXs [4] One useful method is using baffles

to change the direction of flow in the shell side to enhance

turbulence and mixing

This work is supported by the National Nature Science Foundation of China

(grant no 50776068) and Program for New Century Excellent Talents in

Uni-versity of China (grant no NCET-04-0938) The authors also acknowledge

Guangdong Jirong Air-Conditioning Equipment Corporation and postdoctorate

Qiang Gao for providing the experimental results for the continuous helical

baffled shell-and-tube evaporator of an air-conditioning system.

Address correspondence to Professor Qiuwang Wang, State Key

Lab-oratory of Multiphase Flow in Power Engineering, School of Energy and

Power Engineering, Xi’an Jiaotong University, Xi’an, 710049, China E-mail:

wangqw@mail.xjtu.edu.cn

For many years, various types of baffles have been designed,for example, the conventional segmental baffles with differentarrangements, the deflecting baffles, the overlap helical baffles,the rod baffles, and others [5–10] The most commonly usedsegmental baffles make the fluid flow in a tortuous, zigzag man-ner across the tube bundle in the shell side This improves theheat transfer by enhancing turbulence and local mixing on theshell side of heat exchangers However, the traditional STHXswith segmental baffles have many disadvantages: (1) high pres-sure drop on the shell side due to the sudden contraction andexpansion of flow, and fluid impinging on the shell wall caused

by segmental baffles; (2) low heat transfer efficiency due tothe flow stagnation in the so-called “stagnation regions,” whichare located at the corners between baffles and shell wall; (3)low shell-side mass velocity across the tube bundle due to theleakage between baffles and shell wall caused by inaccuracy inmanufacturing tolerance and installation; and (4) short opera-tion time due to the vibration caused by shell-side flow normal

to tube bundle When the traditional segmental baffles are used

in STHXs, higher pumping power is often needed to offsethigher pressure drop for the same heat load During the pastdecades, deflecting baffles, rod baffles, and disk-and-doughnutbaffles have been developed to solve these shortcomings ofthe traditional segmental baffles However, none of these bafflearrangements can solve all the principal problems mentionedearlier New designs are still needed to direct the flow in plug836

Q W WANG ET AL 837

Figure 1 STHXs with helical baffles [16].

flow manner, to provide adequate support to the tubes, and to

have a better thermodynamic performance

The shell-and-tube heat exchanger with helical baffles is

usu-ally called a helixchanger [11–15] It was invented in Czech

Republic and commercially produced by ABB Lummus Heat

Transfer [16] (Figure 1) Helical baffles offer a possible

alterna-tive to segmental baffles by circumventing the aforementioned

problems of conventional segmental baffles; they are accepted

for their outstanding advantages, including: (1) improved

shell-side heat transfer rates/pressure drop ratio; (2) reduced bypass

effects; (3) reduced shell-side fouling; (4) prevention of

flow-induced vibration; and (5) reduced maintenance In the past

decades, the helixchangers have been continuously developed

and improved and have been widely accepted by engineers

The aim of this paper is to present a critical review of the

developments and improvements conducted on helixchangers,

which is of importance for further improvements research in the

future

PRINCIPLE OF HEAT TRANSFER ENHANCEMENT

OF HELIXCHANGERS

As mentioned earlier, one useful method to enhance heat

transfer performance of STHXs is using baffles to change the

flow direction to enhance turbulence and mixing, as do the

helixchangers In the shell side of the helical baffled STHXs,

the helical baffles are located at a certain angle to the tube

bundle, creating a helix flow path for the working fluid The

helix flow provides some characteristics to enhance heat transfer

and low pressure drop Different arrangements of helical baffles

form different constructions of helixchangers (Figure 2)—that

is, baffles touching at the perimeter, overlapping baffles, double

helical baffles, and so on [17]

Lutcha and Nemcansky [17] explained that large differences

in heat exchanger effectiveness are a result of different flow

patterns—that is, perfect mixing flow and perfect plug flow—

the situation depicted in Figure 3 They indicated that the

per-fect plug flow has significant advantages in heat transfer

ver-sus perfect mixing flow, because mixing flow has a substantial

reduction of local driving force for heat transfer, i.e., the

tem-perature difference between the two fluids Therefore, a proper

baffle arrangement should result in a flow pattern that is close

to plug flow A comparison of the helical baffle arrangement

and segmental baffle arrangement approaching plug flow

condi-Figure 2 Different layouts of discontinuous helical baffles [17].

tions was made, and the results suggested that the helical bafflearrangement induced a flow pattern closer to plug flow patternthan segmental baffle arrangement (Figure 4)

Kral et al [18] carried out a series of flow tests to determinethe residence time distribution, in which a standard stimulusresponse technique was used to determine the relative volume

of the dead zones and the amount of back-mixing in the heatexchangers The test results indicated that helical baffle arrange-ments have less back-mixing occurring in the heat exchangercompared with the segmental baffle arrangement (Figure 5) Ithad also been reported that the convective heat transfer across

Figure 3 Comparison of heat exchanger effectiveness for perfect mixing flow and plug flow [17].

heat transfer engineering vol 31 no 10 2010