7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30June 4, 2010
Performance Evaluation of a Horizontal Helically Coiled Tube in Boiling Heat Transfer
Enhancement and its Effect on Pressure Drop
M.A. AkhavanBehabadi, F.M. Shemirani and H. Aria
School of Mechanical Engineering, College of Engineering, University of Tehran, Tehran, IRAN
akharan~ut. ac. ir, fshemirani~ut. ac. ir, razi. pooyan agmail. com
Keywords: Boiling, Coiled Tube, Heat transfer, Helically, Pressure Drop
Abstract
An experimental study has been carried out to investigate flow boiling heat transfer and pressure drop
characteristics of R134a inside horizontal straight and helically coiled tubes. Both testevaporators are
tubeintube heat exchangers, with an inner copper tube of 8.3 mm inside diameter and 9.52 mm outside
diameter and an outer tube with 29 mm inner diameter. The length of straight and helically coiled
evaporators is 1200 mm and 5870 mm, respectively. Refrigerant R134a flows in the inner tube and warm
water flows in the annulus in opposite direction. The experiments were performed with three refrigerant
mass velocities of 112, 132, and 152 kgm2S1. It has been found that the use of helically coiled tubes
enhance the heat transfer coefficient on relatively higher pressure drop penalty, in comparison to that for
the straight tube.
Introduction
Curved tubes have been introduced as one of the passive
heat transfer enhancement techniques and are widely used
in various industrial applications due to their compact
structure and high heat transfer coefficient. Helical coils
are well known types of curved tubes which have been
used in a wide variety of applications such as heat recover
processes, airconditioning and refrigeration systems
chemical reactors, food and dairy processes.
Heat transfer and pressure drop characteristics of
refrigerants during condensation have been studied by a
large number of researchers. The study of the boiling heat
transfer and pressure drop inside helically coiled tubes has
received comparatively little attention in the literature.
Garimella et al.[1] studied the forced convection heat
transfer in coiled annular ducts. They found that the heat
transfer coefficients obtained from the coiled annular ducts
were higher than those obtained from a straight annulus
especially in the laminar region.
Xin et al.[2] investigated the singlephase and twophase
airwater flow pressure drop in annular helicoidal pipes
with horizontal and vertical orientations. Experiments were
performed for the superficial water Reynolds number from
210 to 23,000 and superficial air Reynolds number from 30
to 30,000. A friction factor correlation for singlephase
flow in laminar, transition and turbulent flow regime was
proposed. The twophase flow pressure drop multipliers in
annular helicoidal pipe were found to be dependent on the
LockhartMartinelli parameter and the flow rate of air or
water. The effect of flow rate tended to decrease as the
pipe diameter decreased.
An experimental study on condensation heat transfer of
R134a flowing inside a helical pipe while cooling water
flowing in annulus, was reported by Yu et al. [3]. The
experiments were performed for mass velocities of 100
400 kg/nr's and the Reynolds number of cooling water in
the ranges of 150010000. They found that the orientation
of the helical pipe had significant effect on the heat
transfer coefficient.
Wongwises and Polsongkram [4] investigated both
condensation and evaporation of R134a experimentally to
study the effect of heat flux, mass velocity, and evaporation
temperature on heat transfer coefficients and pressure drop.
They conducted their experiments for high mass velocities
ranging from 400800 kgm s', and provided correlations
for refrigerant Nusselt number and pressure drop.
Al Hajeri et al. [5] reported an experimental investigation
of condensation heat transfer and pressure drop of
refrigerant R134a inside a helical tube. Their study
concerns the condensation of R134a flowing through
annular helical tubes with different operating refrigerant
saturated temperatures including 360C, 420C, and 480C.
The average pressure drop is measured and compared. The
measurements of R134a were performed on mass
velocities ranges from 50 to 680 kgm s '. Their
experimental results shows that the refrigerant side heat
transfer coefficient and overall heat transfer coefficients
FIGURE 1
(A) STRAIGHT AND HELICALLY COILED TUBES USED AS TEST
SECTION (B) EXPERIMENTAL SETUP
TABLE 1
DIFFERENT ELEMENTS OF FIGURE 1
1Compressor 6 Expansion valve
2 Condenser 7 Pre evaporator
3 Flow meter 8 Test evaporator
4 IReceiver 9 After evaporator
5 Filter drier 10 Accumulator
Thermocouple P Pressure gauge
DP Differential pressure transducer
Data Reduction
Figure 2 shows the refrigerant flow diagram from the
entrance of expansion valve to the exit of after evaporator.
The following procedure is used to determine the quality
of refrigerant at the inlet and exit of the test evaporator, the
heat transfer coefficient and pressure gradient, from the
data recorded during each test run at steady state
7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30June 4, 2010
Location of thennoccuple
increase with increase of the mass flux of flowing R134a
and decrease of vapor saturated temperatures of R134a.
The pressure drops in annular helical tubes increase as the
mass velocity of refrigerant increases.
Yang et al. [6] carried out a numericalexperimental study
on flow boiling of R141B in a horizontal coiled tube.
They conducted their tests in two flow rates of 10 and 15
1/hr and three different heat fluxes, and observed that flow
velocity predicted in the simulations showed a strongly
dependence on the phase distribution. Moreover, they
depicted the evolution of flow pattern inside the tube.
Existing Ozone in the Stratosphere absorbs deadly
radiation of ultraviolet from sun. This is why the depletion
of Ozone layer is becoming a great concern to most
societies. A major cause of this phenomenon is releasing
large amounts of CFCs which converts Ozone to Oxygen.
Since the thermo physical properties of R134a are very
similar to those of the conventional CFCs, it is the best
choice to replace them. Therefore, the properties of R134a
should be studied in detail before it is applied.
In the present study, an experimental investigation on flow
boiling heat transfer and pressure drop of R134a inside
horizontal straight and helically coiled tube evaporators
have been carried out.
Experimental Facility
The experimental setup which was used in this
investigation is a well instrumented vapor compression
refrigeration system. There are three different evaporators;
namely preevaporator, test evaporator, and after
evaporator. Both preevaporator and afterevaporator were
electrically heated. The purpose of using preevaporator
was to achieve any desired vapor quality at the inlet of test
evaporator. Afterevaporator was used to ensure that all the
fluid in the test tube is superheated before entering the
compressor.
Both the testevaporators were tubeintube heat
exchangers, with an inner copper tube and an outer
stainless steel tube. The length of straight and helically
coiled test evaporators was 1200 mm and 5870 mm,
respectively. Refrigerant R134a flowed in the inner tube
and warm water flowed in the annulus in opposite
direction.
The straight and helically coiled tubes used as test
evaporators and also the schematic diagram of
experimental setup have been shown in Figure 1.
Different elements of this setup are introduced in Table 1.
The inside diameter of the inner copper tube is D,=9.52
mm and its thickness is 0.62 mm. The outside diameter of
the outer tube is De=34 mm. The pitch of the coil is b=45
mm and the outer diameter of the coil is D=305 mm. To
measure the outside surface temperatures of inner tube,
Ttype thermocouples were used at six different locations
on both straight and helically coiled tubes (see Figure 1). At
each section, two thermocouples were employed on
opposite sides of outer surface of the inner tube. The
pressure drop across the test section was also measured by
an accurate differential pressure transducer apparatus. The
provisions were also made to measure all the other
necessary parameters.
n
~V~~Vb
where, vilwr cywr Twa aand Twru are themass
flow rate, specific heat, inlet temperature and outlet
temperature of hot water flowing in test evaporator,
respectively.
The average vapor quality in test evaporator, x:
The average vapor quality in test evaporator is defined as:
7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30June 4, 2010
X, + 3
x= (6)
The average heat transfer coefficient, h:
h = 27
A w~T)
Where, Tand Al are the average inside wall
temperature, and inside surface area of test evaporator
inner tube, respectively. T, is the average temperature of the
refrigerant at the inlet and outlet of test evaporator.
Al = yr DI L (8)
conditions. The thermodynamic and transport properties of
R134a are evaluated using ASHRAE Handbook of
Fundamentals [7].
T1 PI Qa P2 p Pa
(I 1 2
use .vawm cnawam mwat" t
FIGURE 2
FLOW DIAGRAM IN EXPANSION VALVE AND EVAPORATORS
The inlet vapor quality of the test evaporator, x2
72
where a2 iS the refrigerant enthalpy at the test evapo
inlet, if is the enthalpy of saturated liquid and ifg i
enthalpy of vaporization.
(1) The fricionalpressue graienPF)
The twophase flow pressure gradient is the sum of three
rator components: the frictional pressure gradient, the
s the momentum pressure gradient and the hydrostatic pressure
gradient as follows:
12 = 11 1 (2)
where i, is the enthalpy of the liquid refrigerant at the inlet
of preheater, rizref is the mass flow rate of the refrigerant,
Qz is the heat transfer rate from electrical coil to the
refrigerant in the preevaporator.
The outlet vapor quality of the test section, x3 :
i3 f
x3 = (3)
where i3 iS the refrigerant enthalpy at the test evaporator
outlet. As a consequence, the outlet enthalpy of the
refrigerant flow is calculated from
13 = 12 + (4)
where the heat transfer rate from the hot water to the
refrigerant in the test evaporator, Q2, iS obtained from
The hydrostatic pressure gradient is defined as:
(g = gsini9 ap +(1a)pl,
where p, and pf are the density of vapor and liquid
TOSpectively and a is the void fraction.
The momentum pressure gradient can be calculated by the
results of one dimensional two phase separated flow model
analySIS aS follows:
dP 2d (1 x)2 2
dz dz p P(1a) pga
in which, Gt is mass velocity of refrigerant. The void
fraction values in the above equations can be determined
from [8] as follows:
a=Pg [(~~(iPg
1.18(1 x) [go(pf pg )]o 5 P
G2 05
e2 = iwat p,wat (Twat~,,n Twat,out
Finally, the frictional pressure gradient is computed by
subtracting the calculated momentum and hydrostatic
pressure gradients from measured two phase total pressure
gradient by Equation (13).
dP)= dPF dP I dP
dz dz dz dz
dP dP dP dP
F a Ig
dzdz dz dz
7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30June 4, 2010
It is found from Figure 4 that by increasing the vapor
quality, the pressure drop increases. At a same average
quality, it can be seen that the pressure drop increases with
the increasing of mass velocity. This can be explained by
the fact that, any increasing in mass velocity accelerates
the twophase flow velocity, which leads to a greater tube
wall shear stress.
Helically coiled tube
Heat transfer coefficients of horizontal helically coiled
tube versus average vapor quality for different mass
velocities are depicted in Figure 5. It can be seen that same
as straight tube, the heat transfer coefficient increases by
increasing of vapor quality and mass velocity.
Besides the reasons stated about this phenomenon in
straight tube section, the effect of strengthening the
Secondary flow should be noted. By increasing the strength
Of secondary flow, the redopisition and entrainment rates
increase. This enhancement induces the increased number
and larger size of waves on the liquid film surface, which
leads to increasing of heat transfer area. In addition, the
higher velocity, due to the increasing mass velocity,
increases the degree of turbulence of the fluid, which
results in a higher heat transfer coefficient.
s000
Results and Discussion
Straight Tube
The comparison of heat transfer coefficient of evaporative
R134a flow versus average vapor quality for three
different mass velocities is illustrated in Figure 3.
4500 ,,
2 4000
S3500
3000
S2000
S1500
S1000
1500
O "
a .
.
*G=112 Kg/m2s
o G=132 Kg/m2s
A G=152 Kg/m2s
0 8 1
0 0 2 0 4 0 6
Vapor Quality
FIGURE 3
HEAT TRANSFER Ci)EFFICIENT <)F Hi)RIZi)NTAL STRAIGHT TUBE
It is obvious from Figure 3 that the refrigerant side heat
transfer coefficient continuously increases by increasing
vapor quality. This trend has two major reasons. Firstly
because of the boiling, the mass of liquid refrigerant which
has a low specific volume decrease, which leads to
increase the velocity of fluid and therefore a higher heat
transfer coefficient is achieved. Secondly, as the average
vapor quality increases, the liquid film thickness decreases,
which reduces the thermal resistance in the liquid film and
a higher heat transfer coefficient is obtained. It is also
observed that by increasing the mass velocity the heat
transfer coefficient increases. However, the mass velocity
is found to have less effect on heat transfer coefficient at
low average vapor qualities.
The frictional pressure gradient of R134a flowing in
straight horizontal tube versus average vapor quality for
different mass velocities is depicted in Figure 4.
8000
S7000
S6000
~5000
S4000
'3000
S2000
1000
a =
a
As a
I
4 *
* G=112 Kg/m2s
SG=132 Kg/m2s
A G=152 Kg/m2s
O 0 2 0 4 0 6 0 8 1
Vapor Q~uality
FIGURE 5
HEAT TRANSFER CO EFFICIENT <)F HO RIZO NTAL HELICALLY CO ILED TUBE
The variation of frictional pressure gradient of horizontal
helically coiled heat exchanger versus average vapor
quality is depicted in Figure 6 for different mass velocities.
The trend of changes of pressure drop is just same as that
of straight tube. Except the reasons stated for these trends
in straight tube section, there are two more reasons: the
longer distances that fluid particles should move and the
turbulence causes by secondary flows.
Comparison of heat transfer coefficient and frictional
preSsure gradient of helically coiled tube and straight tube
veTSus average vapor quality for different mass velocities
iS depicted in Figures 78. From these figures it is found
that at a same mass velocity and vapor quality, both heat
transfer coefficient and frictional pressure gradient of
horizontal helically coiled tubes are higher than those of
Straight tubes due to the effect of secondary flows.
D *
*
*G=112k/2
aG= 132k/2
G=152k/2
0 0 1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 1
Vapor Qualty
FIGURE 4
FRICTI<)NAL PRESSURE GRADIENT <)F Hi)RIZi)NTAL STRAIGHT TUBE
E` r .
2 *G=112 Kg/m2s
1I a G=132 Kg/m2s
rG=152 Kg/m2s
0 0 1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 1
Vapor Q~ualitr
FIGURE 6
FRICTIONAL PRESSURE GRADIENT OF HORIZONTAL HELICALLY COILED
TUBE
9000
8000
S7000
6000
O 4000
"3000
2000
loo
helical G=112Kg/m2s
a helical G=132Kg/m2s
helical G=152Kg/m2s
xstraight G=112Kg/m2s
straight G=132Kg/m2s
+straight G=152Kg/m2s
.
A
r a.
xxx x x x x
x
0 01 02 03 04 05 06 07 08 09 1
Vapor Quality
FIGURE 7
COMPARISON OF HORIZONTAL STRAIGHT AND HELICALLY COILED TUBE
HEAT TRANSFER COEFFICIENT
7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30June 4, 2010
them and can produce the conditions for the comparison of
these two parameters. This third parameter can be the
compressor power consumption for testevaporator length
to heat transfer coefficient ratio, W/h, or compressor power
consumption to heat transfer in it, W/Q. The reason is that
the pressure drop increasing rate directly depends upon the
COmpressor power consumption rate.
IH perfOrmance evaluation of twisted tape insert on boiling
heat transfer of R12 inside horizontal tubes, Agrawal and
Varma [9] also used the ratio of pumping power to heat
transfer enhancement rate as the evaluation criterion. They
calculated the increased pumping power due to the
preSence of twisted tape in testevaporator by multiplying
the volumetric flow rate and produced pressure drop in test
evaporator:
W = v AP,, (14)
Finally, based on this criterion, they prepared a figure for
different tubes. In the present work, firstly, the ratio of the
pOWeT COHSumption of compressor to heat transfer
COefficient of straight tube (Winl is calculated; then
the same ratio for helically coiled tubes (Winl is
COmputed. The compressor power is calculated from (14).
Eventually, the ratio for straight tube is divided by that for
the flow in helically coiled tube.
MP, Tes it hflelical 1 R
h Stagt hStraight R
(15)
(Test Helical
~M~s(MrTesragt1 Stah
If the ratio of Rh p, iS greater than one, to apply the coiled
tube is useful.
The variation of Rh/R, with mass velocity has been shown
in Figure 9. From this figure it is concluded that among the
different vapor qualities and different mass velocities, at
G=112 kgm2S1 and x=0.82, the horizontal helically coiled
tube has the best performance with Rh/R,=1.15. At a
constant vapor quality, by reducing mass velocity, Rh/R,
decreases.
14
*helical G=112Kg/m2s
xstrlaigh G=112Kg/m2s
straight G=132Kg/m2s a. .
+straight G=152Kg/m2s A a **
t a
e
a a
0 02 04 06 08 1
Vapor Quality
FIGURE 8
COMPARISON OF HORIZONTAL STRAIGHT AND HELICALLY COILED TUBE
FRICTIONAL PRESSURE GRADIENT
Performance evaluation of helically coiled tubes
The present experimental data analysis shows that the use
of horizontal helically coiled tube increases the heat
transfer coefficient with a pressure drop penalty. The
pressure drop and heat transfer coefficient are two
independent parameters which do not relate by any
equation, so comparison is difficult. Therefore, a third
parameter should be considered which relates to both of
O 6
*G= 112 Kg/m2s
a G= 132 Kg/m2s
AG= 152 Kg/m2s
0 4 0 6 08 8
Vapor Quality
FIGURE 9
VERSUS AVERAGE VAPOR QUALITY AT DIFFERENT
MASS VELOCITIES
0 2
EVALUATION OF Ro/Rn
7th International Conference on Multiphase Flow
ICMF 2010, Tampa, FL USA, May 30June 4, 2010
References
[1] Garimella, S., Richards, D.E. and Christensen, R.N.,
Experimental investigation of heat transfer in coiled
annular duct, Transactions of ASME, Vol. 110, 329
336 (1998).
[2] Xin, R.C., Awwad, A., Dong, Z.F. and Ebadian, M.A.,
An experimental study of singlephase and twophase
flow pressure drop in annular helicoidal pipes,
International Joumnal of Heat and Mass Transfer, Vol.
18, 482488 (1997).
[3] Yu, B., Han, J.R., Kang, H.J., Lin, C.X., Awwad, A.
and Ebadian, M.A., Condensation heat transfer of R
134a flow inside helical pipes at different orientations,
Int. Communications in Heat and Mass Transfer, Vol.
30, 745754 (2003).
[4] Wongwises, S. and Polsongkram, M., Evaporation
heat transfer and pressure drop of HFC134a in a
helically coiled concentric tubeintube heat
exchanger, Intemnational Joumnal of Heat and Mass
Transfer, Vol. 49, 658670 (2006).
[5] AlHajeri, M.H., Koluib, A.M., Mosaad, M. and Al
Kulaib, S., Heat transfer performance during
condensation of R134a inside helically coiled tubes,
Energy Conversion and Management, Vol. 48, 2309
2315 (2007).
[6] Yang, Z., Peng, X.F. and Ye, P., Numerical and
experimental investigation of two phase flow during
boiling in a coiled tube, Intemnational Joumnal of Heat
and Mass Transfer, Vol. 51, 10031016 (2008).
[7] ASHRAE, Handbook of Fundamentals, American
Society of Heating, Refrigerating, and Air
Conditioning Engineers, Atlanta, GA, 1997.
[8] Ould Didi, M.B., Kattan, N. and Thome, J.R.,
Prediction of TwoPhase Pressure Gradients of
Refrigerants in Horizontal Tubes, Intemnational
Journal of Refrigeration, Vol. 25, 935944 (2002).
[9] Agrawal, K.N. and Varma, H.K., Heat transfer during
forced convection boiling of R12 under swirl flow,
Journal of Heat Transfer, Vol. 108, 567572 (1986).
Also from Figure 9 It can be seen that in some points
specially at high vapor quality region, this ratio is greater
than one; in these points the use of helically coiled tube is
recommended.
If the ratio of Rh p, iS less than one, only under the certain
conditions and for specific applications considering the
appropriate consistence between the heat transfer
performance and the rate of energy penalty paid for
increased pressure drop, the horizontal helically coiled
tube can be used. For example, in the places where due to
the limitation of spaces, more compact and smaller heat
exchangers are needed and low cost energy resources are
available to supply the excess power consumed for the
increase in pressure drop.
Finally, it is concluded that, the penalty which should be
paid indirectly for compensation of pressure drop
increasing due to the use of horizontal coiled tube instead
of straight tube, generally, is more than the related heat
transfer enhancement obtained. In the other word, the main
purpose of using the coiled tube instead of straight tube is
to have a compact heat exchanger with a lower occupied
space, not saving in consumption of energy.
Conclusions
1 The heat transfer coefficient of both straight and
helically coiled tube continuously increased by
increasing of vapor quality and mass velocity.
2 The pressure drop for both straight tube and
helically coiled tube increased with the increase
of mass velocity and vapor quality.
3 At a constant mass velocity and vapor quality,
both the heat transfer coefficient and frictional
pressure gradient of horizontal helically coiled
tubes were higher than those of straight tubes.
4 It was found that at mass velocity of 112 kgm2S1
and vapor quality of 0.82, the horizontal helically
coiled tube has the best performance with
Rh p,=1.15.
5 Finally, it was concluded that, the main purpose of
using the coiled tube instead of straight tube is to
have a compact heat exchanger with a lower
occupied space, not saving in consumption of
energy.
