Heat transfer enhancement using self-rewetting fluids

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Download by: [Mohammadreza Bohloul] Date: 20 September 2016, At: 01:36
Experimental Heat Transfer
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ISSN: 0891-6152 (Print) 1521-0480 (Online) Journal homepage: http://www.tandfonline.com/loi/ueht20
Heat transfer and Marangoni flow in a circular
heat pipe using self-rewetting fluids
S.M. Peyghambarzadeh, M.R. Bohloul & N. Aslanzadeh
To cite this article: S.M. Peyghambarzadeh, M.R. Bohloul & N. Aslanzadeh (2016): Heat transfer
and Marangoni flow in a circular heat pipe using self-rewetting fluids, Experimental Heat
Transfer
To link to this article: http://dx.doi.org/10.1080/08916152.2016.1233148
Accepted author version posted online: 20
Sep 2016.
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Heat transfer and Marangoni flow in a circular heat pipe using self-
rewetting fluids
S.M. Peyghambarzadeha
, M.R. Bohloulb
*, N. Aslanzadeha
a
Department of Chemical Engineering, Mahshahr branch, Islamic Azad University,
Mahshahr, Iran
b
Young Researchers and Elite Club, Mahshahr branch, Islamic Azad University, Mahshahr,
Iran
E-mail: mohammadreza.bohlool@gmail.com
Abstract
In this study, Marangoni flow and heat transfer enhancement in a heat pipe have been
investigated. The experiments were carried out at different heat inputs. Constant temperature
water bath was used at the condenser section at three temperature levels. Heat transfer
coefficients and thermal resistances of the heat pipe were measured for pure water and
water/butanol solutions. The experimental results confirmed that the heat pipe filled with
butanol solutions showed better thermal performance than the water filled-heat pipe. At
maximum heat flux, 25% heat transfer improvement was obtained when 7 %wt butanol
solution was used instead of pure water
Keywords: Heat pipe; Thermal resistance; Heat transfer coefficient; Marangoni effect;
Butanol; Self-rewetting fluids
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1. Introduction:
Heat pipes are simple and effective devices of very high thermal conductivity with no moving
parts. They have many benefits such as: low weight, maintenance-free, reliability, and
increased heat dissipation. Due to these reasons, they are widely used in electronic cooling,
air conditioning, power generation, chemical engineering, and spacecraft cooling [1, 2]. The
scientific background and the previous fundamental results on the thermal performance of the
heat pipes with different geometries and different working fluids are summarized in the next
part of this paper.
Generally, heat pipe performance is strongly dependent on the geometry, working fluid, wick
structure, surface tension, wetting angle of the fluid, and Marangoni flow in mixtures. One of
the methods for the heat transfer enhancement in the heat pipe is the application of additives
to the working fluids to change the fluid transport properties and flow features (such as
interfacial tension changes with temperature).
In general, for all the working fluids (pure liquids) that were used in conventional heat pipes,
the surface tension is a decreasing function of the temperature, and it has a detrimental effect
on the heat pipe performance. Since fluid motions due to a surface temperature gradient are
directed toward the cold regions of the surface, it may be unfavorable for the return of the
liquid to the evaporator. Researches on interfacial phenomena between liquid and vapor
phases have shown that the surface tension of alcohol (with carbon atoms higher than 4)
/water mixture, so called self-rewetting fluids, goes through a minimum as a function of
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temperature. For example, in Fig. 1 (a) [3, 4], after 65 ̊C, the surface tension of Heptanol
solution increases with increasing temperature. There is in a range of temperature in which
the surface tension of heptanol solution increases.
Justification can be attributed to temperature and composition variations, which can create
non-uniform variations of interfacial tension. A direct result of non-uniform variations of
interfacial tension is a surface flow that is directed from regions of lower surface tension to
regions with higher surface tension.
As non-uniformities continue to exist, liquid motion due to steady surface tension gradient
becomes established, it is also known as “Marangoni effect”. This effect creates circulation
flow, as shown in Fig .1 (b). As indicated, Marangoni effect will increase the pumping effect
and therefore, the liquid mass flow rate inside the heat pipe. Table 1 summarizes the previous
studies on heat pipes systems displaying the working fluids, range of operating conditions,
and effect of Marangoni flow on the heat transfer enhancement for binary mixtures.
Furthermore, Hu et al. [12] experimentally verified the heat transfer enhancement of micro
oscillating heat pipes using self-rewetting fluid (heptanol aqueous solutions). In their
experiments, the input powers to the evaporator section were 10 - 80 W and condenser wall
temperatures were 30 - 50 °C. Their results showed that at the horizontal orientation, the
micro oscillating heat pipes using the self-rewetting working fluids exhibited much better
thermal performance compared with water as the working fluid. On the other hand, at vertical
orientation, the enhancement due to the self-rewetting working fluids was not obvious.
Savino et al. [13] conducted an experimental study to evaluate thermal performance of heat
pipe filled with water and aqueous solutions of long chain alcohols. Their experiments have
been performed with axially grooved copper heat pipes with length of 250 mm, and diameters
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of 4 and 8 mm, and glass heat pipe with inner diameter of 10 mm, wall thickness of 1 mm,
and length of 165 mm. Their results showed that both heat pipes filled with long chain
alcohol solutions at suitable concentrations performed better than heat pipes filled with pure
water.
In this study, thermal performance of copper made heat pipe with porous wick and self-
rewetting working fluids have been analyzed. 4 %wt and 7 %wt of butanol was added to the
pure water and these solutions were used as working fluids in the heat pipe. It must be
emphasized that the heat pipe has been constructed in two different diameters and this
configuration has not been studied previously. In this configuration, the evaporation section
has larger diameter than the remaining sections of the heat pipe. The benefit of this
configuration is that the evaporated liquid passes through the nozzle shaped entrance of the
adiabatic section with higher velocity.
2. Experimental
2.1. Material
Butanol (C4H9OH) with the purity higher than 99.9% mol was purchased from Merck
Company and it was utilized without future purification. Also, its purity was obtained from
supplier. The deionized water with the purity higher than 99.5% mol was used in this study.
The butanol aqueous solutions (4 %wt and 7 %wt) were employed as the measurement
samples. Table 2 presents some of the physical properties of pure water and pure butanol.
2.2. Surface tension measurement
A number of different techniques have been developed to measure the surface tension of pure
liquids and aqueous solutions. Details of these different techniques were given in [15]. In this
study, to show the anomalous behavior of the self-rewetting fluids, the surface tension of the
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working fluids were measured using two different methods including ring method and glass
capillary tube method at different temperatures. In both methods, the liquid was warmed and
stirred using heater stirrer. When the required temperature nearly reached, the heater was
switched off and allowed the temperature to be stabilized.
The capillary method is the oldest method used for the surface tension measurement. A
consequence of the surface tension appearance at the liquid / gas interface is moving up of the
liquid into a thin tube which is usually made up of glass. This phenomenon was applied for
determination of the liquid surface tension. For this purpose, we used the glass capillary tube
with inner diameter of 1.5 mm, outer diameter of 1.8 mm, and length of 16 cm. The glass
capillary tubes were filled with pure water and butanol aqueous solutions. Also, the contact
angles have been measured for working fluids at different temperatures with the same
apparatus used for the surface tension measurement.
In the ring method, a thin plate is used to measure equilibrium surface or interfacial tension at
the air / liquid or liquid / liquid interfaces. The measuring ring was carefully degreased with
alcohol, rinsed in deionized water and dried. The ring was attached to the left arm of the
torsion dynamometer using a silk thread. The indicator of the torsion dynamometer was set to
zero and the weight of the ring compensated using the rear adjusting knob.
2.3. Experimental apparatus and procedure
The experimental setup used in this research was similar to our prior work [17]. Fig.2 shows
a schematic presentation of the experimental setup. The apparatus consisted of a circular heat
pipe, condenser, vacuum pump, DC power supply (DCPS), electrical heater, pressure and
temperature measuring instruments, water circulator, and data acquisition system. The heat
pipe was made up of smooth copper tube. Porous wick was attached to the inner surface of
the heat pipe wall. The wick consisted of three layers of aluminum mesh (mesh number 100,
sheet thickness 1.5 mm) which was flexible to be deformed. A close contact between the
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mesh and the inner wall could be guaranteed due to the internal tension of the mesh. The test
section was positioned horizontally and it consisted of evaporator, adiabatic, and condenser
sections. The dimensions of these sections are presented in Table 3.
As can be seen in Table 3, the evaporator section has larger diameter than the remaining
sections of the heat pipe. Almost, this kind of dual diameter circular heat pipe is new and this
configuration has a main benefit. In this kind of heat pipe in comparison with ordinary heat
pipes, the evaporated liquid enter the adiabatic section with higher velocity due to nozzle
shaped configuration. The evaporator section was heated by an electrical heater (Watlow Co.)
wrapped around the pipe. The condenser section was cooled continuously by the cooling
water circulating in a cube with the dimensions of 20×20×20 cm. The temperature and the
flow rate of the cooling water were accurately controlled to keep the operating pressure at a
constant value for different heat fluxes. To prevent heat loss through surfaces, the evaporator
and the adiabatic sections were carefully insulated by glass wool. Also, the flexible insulation
material allows the heat pipe to expand after its temperature rises.
Three E-type thermocouples were installed to measure the outside surface temperatures of the
heat pipe and three others to measure the working fluid temperatures. Each group includes
one thermocouple at the evaporator section, one at the condenser section, and one at the
adiabatic section. Very tiny grooves were machined in the heat pipe walls and a high
conductivity cement was utilized to embed the thermocouples within the heat pipe wall.
Distribution of the thermocouples along the axial direction is indicated in Fig. 3. The wall
temperature distribution along the circumference direction was quite uniform because the
mesh structure could make the liquid film uniformly filled into the mesh layers of the heat
pipe. A pressure transmission was placed at the central location of the adiabatic section which
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was used to measure the operating pressure. A DC power supply (MEGATEK, Model: MP-
3003D) was used as the source of power at the evaporator section.
Different working fluids including pure water, water/butanol 4 %wt, and water/butanol 7
%wt were filled into the heat pipe using a syringe. According to the previous researches on
the liquid filling ratio [8, 11 and 12], the liquid filling ratio was selected 50±5% of the total
volume of the heat pipe in all the experiments performed. Before the experiment, the vacuum
pumping and liquid preheating processes were performed to remove the dissolved gases in
the heat pipe and working fluid. Eventually, the experiments were performed with the heat
pipe in the horizontal orientation. Also, the heat input to the evaporator was varied from 1-27
W and the condenser wall temperature was maintained constant at the temperatures of 15, 25,
and 35 °C. For error reduction, the heat pipe was carefully cleaned with deionized water,
before every experiment.
2.4. Data reduction
The heat pipe performance was shown by different parameters such as evaporator thermal
resistance, condenser thermal resistance, and total thermal resistance. The evaporator thermal
resistance is defined as the temperature difference between the evaporator wall temperature
(Te,w) and the vapor temperature at the evaporator section (Te,v) divided by the input power
(Qe) [18, 19]:
, , , ,
,
1e w e v e w e v
e
e e e
e e e
T T T T
R
VIQ h A
D L
 
  
(1)
where V is voltage, I is amperage, De,e is external diameter of the evaporator section, Le is
length of the evaporator section, and he is heat transfer coefficient at the evaporator section.
The condenser thermal resistance is defined as the temperature difference between the
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condenser wall temperature (Tc,w) and the vapor temperature at the condenser section (Tc,v)
divided by the input power (Qe) [20]:
, , 1c v c w
c
e c c
T T
R
Q h A

  (2)
where hc is the heat transfer coefficient at the condenser section. The total thermal resistance
of the heat pipe is computed as [21]:
, ,e w c w
e
T T
R
Q

 (3)
2.4. Uncertainty analysis
An uncertainty analysis has been carried out according to the method proposed by Moffat
[22]. The uncertainty of the thermal resistances comes from the errors in the measurement of
wall and vapor temperatures at different sections, diameter, length, amperage, and voltage as
follows:
2 22 22 2
, ,v ,e
, ,v , ,v ,e ,e
( )
( ) ( )
e w e ee e e e e e e
e e w e e w e e e e e
T T DR R R R R R LV I
R T T T T V V I I D D L L
                
                                    
(4)
2 22 22 2
c,v c, ,e
c,v c, c,v c, ,e ,e
( )
( ) ( )
w ec c c c c c e
c w w e e e e
T T DR R R R R R LV I
                
                                    
(5)
2 2 22 2 2
, c, ,e
, c, , c, ,e ,e
( )
( ) ( )
e w w e e
e w w e w w e e e e
T T D LR R R V R I R R
                  
                                      
(6)
The contributions of the main parameters involving in the evaluation of thermal resistances
according to Eq. (4)-(6) were summarized in Table 4. The analysis showed that the maximum
values of the uncertainty of the evaporator thermal resistance, the condenser thermal
resistance, and the heat pipe thermal resistance were ±6.2 %, ± 5.5 %, and ±5.8%,
respectively.
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3. Result and discussion
3.1. Surface tension measurements
In both methods, to check the reproducibility of the experiments, all runs were repeated twice
and even trice. The repetition of the experiments showed that the maximum deviation was
less than ± 2.35 %. The averaged results of both methods are presented separately in Fig. 4 at
different temperatures. In this figure, the surface tension measured in the present work were
also compared with those published by Pachghare et al. [7] and Savino et al. [16] for the
similar solutions.
The measurements showed that, in general, the surface tension is a function of temperature
and concentration for alcohol solutions. It increases with temperature above some point, but it
usually increases with increasing concentration, which is beneficial for heat transfer
enhancement of the heat pipes. On the contrary, for pure water, the surface tension decreased
with temperature in the selected temperature range as shown in Fig. 1 (a).
Also, in the capillary method, contact angles were measured for different butanol solution at
different temperatures. The average measured contact angles for butanol solutions at different
temperatures are summarized in Table 5.
3.1. Temperature variations
Fig. 5(a-c) illustrates the vapor core temperatures along the heat pipe at constant input heat
flux (2400 W/m2
) and at different condenser temperatures. The results were reported for
different working fluids including water, water / 4 %wt butanol, and water / 7%wt butanol.
Fig. 6(a-c) demonstrates the variation of heat pipe wall temperature at constant input heat
flux (2400 W/m2
) and at different condenser temperatures when heat pipe was filled with
different working fluids. The results of both figures show that the higher vapor core and wall
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temperatures are obtained at the evaporator section, but these temperatures are different for
the working fluids. Also, higher temperature of the condenser causes the returning liquid to
the evaporator to be warmer. Due to this reason, the wall temperature and the vapor core
temperature of the evaporator section increase when the condenser temperature increases.
In order to have a better insight on the temperature variation inside the heat pipe, Fig. 7
shows that at the evaporator section, the vapor core temperature decreases with increasing the
concentration of the solution while at the adiabatic and the condenser sections, the vapor core
temperature increases with increasing the concentration of the solution. It means that heat
transferred to the cold sections with better performance when the heat pipe filled with butanol
solutions. This observation can be explained using two different phenomena influencing the
heat pipe thermal performance:
a) It is shown in Fig. 7 that lower vapor core temperatures at the evaporator section (Tev)
were observed when self-rewetting fluids were used in the heat pipe. Furthermore, Fig. 6
shows that lower wall temperatures at the evaporator section (Tew) were also recorded for
self-rewetting fluids. Totally, using self-rewetting fluids caused the colder evaporator in
comparison with water at similar heat input. This may refer to some of related physical
properties of the working fluids. Table 2 shows that the butanol solutions have lower heat
of vaporization (or higher vapor pressure) in comparison with water. Consequently, when
using self-rewetting fluids as working fluid, the amount of vapor generation in the
evaporator section will increase. The generated vapor with the accompanying heat goes
toward the colder region. This point simultaneously decreases the temperature of the
evaporator section and increases the temperature of the condenser section. We called this
effect as “vapor departing mechanism”.
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b) According to Fig. 2, for all pure working fluids used in conventional heat pipe, the
surface tension is a decreasing function of the temperature, conversely in binary mixtures
of the so called self-rewetting fluids, the surface tension increases with increasing
temperature. The existence of a temperature gradient along the heat pipe induces a
surface tension gradient at vapor-liquid interface that moves the liquid toward region of
higher surface tension. This effect, known as Marangoni flow, provides an additional
mechanism for liquid to return from the condenser to the evaporator, other than capillary
and gravitational forces. Thus, the liquid has less residence time at the condenser and
consequently, the vapor core temperature at the condenser section (Tcv) and at the
adiabatic section (Tav) did not decrease for self-rewetting fluids as for water. We called
this effect as “liquid arrival mechanism”.
Therefore, the existence of a suitable “vapor departing mechanism” and also an appropriate
“liquid arrival mechanism” increases the rate of heat transfer between different sections of
the heat pipe, or equivalently, decreases the temperature gradient along the heat pipe.
Different “vapor departing mechanism” have been investigated previously: the addition of
nanoparticles to the base fluids [23] and, the use of fluids with lower heat of vaporization are
two examples of this mechanism. Also, Marangoni flow, gravitational force, and capillary
force [24, 25] are some examples of “liquid arrival mechanism”.
3.2. Heat transfer coefficient
The change in the heat transfer coefficient with heat flux for different working fluids at
different condenser temperatures is shown in Fig. 8(a-c). As can be seen, the heat transfer
coefficient increases with increasing heat flux at the evaporator section. At a constant heat
flux, the heat transfer coefficient increases with increasing the condenser temperature. Also,
at similar conditions (heat flux and condenser temperature), heat transfer coefficient of
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water / 7 %wt butanol solution is higher than that in other working fluids. The temperature
difference between the evaporator wall (Te,w) and the vapor core (Te,v) is an indicative
criterion for good heat transfer performance in the evaporator section. This quantity decreases
with increasing the concentration of solution.
Fig. 8(a-c) clearly shows that at the lower operating temperature or heat flux, conduction is
the governing heat transfer mechanism. With increasing heat flux, evaporation heat transfer
occurs in the heat pipe and consequently, the heat transfer coefficient dramatically increases.
3.3. Thermal resistance
The variation of the evaporator thermal resistance as a function of heat flux and condenser
temperature was presented in Fig. 9 for different working fluids. Based on Eq. (1), the
thermal resistance decreases with increasing heat transfer coefficient at the evaporator
section. The results of the heat transfer coefficient showed that, at a constant heat flux, the
heat transfer coefficient increases with increasing condenser temperature. On the contrary,
the thermal resistance decreases with increasing condenser temperature (can see in Fig. 8).
Also, at similar conditions (heat flux and condenser temperature), the heat transfer coefficient
of water / 7 %wt butanol solution is higher than that in other working fluids, then, the thermal
resistance decreases with increasing the concentration of the solution. Both of these
observations is due to the existence of suitable “vapor departing mechanism” and “liquid
arrival mechanism” in butanol solutions.
Fig. 10 (a-c) indicates the variation of condenser thermal resistance as a function of heat flux
and condenser temperature. As can be seen, the condenser thermal resistance decreases with
increasing heat flux. The values of condenser thermal resistance for different working fluids
are very low and consequently, the heat transfer coefficient is very high. This is due to the
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condensation occurred in the condenser section of the heat pipe which is the most effective
mechanism of heat transfer. Figs. 5 and 6 show that the temperature difference between the
wall and the vapor core at the condenser section decreases with increasing the concentration
of solution. It is due to the surface tension variation with temperature and induced Marangoni
flow in self-rewetting fluid. Thus, according to Eq. (2), the thermal resistance at the
condenser section decreases with increasing the concentration of solution.
Fig. 11(a-c) shows the variation of total thermal resistance of the heat pipe as a function of
heat flux and condenser temperature. From these figures, it is found that the total thermal
resistance decreases with increasing heat flux at condenser temperatures of 15 °C and 25 °C.
But, at Tc=35 °C different trend was observed for all the working fluids. It is due to the fact
that the high condenser temperature and low heat flux lead to the lower temperature
difference between the condenser wall and the evaporator wall (Te,w – Tc,w). In addition, the
experiments pointed out that heat pipe filled with butanol solutions exhibit better heat transfer
performance than ones filled with water, and total thermal resistance of water / 7 %wt butanol
solution is less than that in the water / 4 %wt butanol solution. It is due to the surface tension
variation with temperature and Marangoni flow in the self-rewetting fluids. As a result, this
effect could improve the heat transfer performance of self-rewetting fluids in the heat pipe.
The importance of “vapor departing mechanism” and “liquid arrival mechanism” can be
found in similar works performed. For example, Peyghambarzadeh et al. [17] compared the
performance of different working fluids including water, ethanol, and methanol in a similar
heat pipe. The “vapor departing mechanism” in these working fluids is different since the
latent heat of vaporization of water is almost twice of that of ethanol and methanol.
Therefore, at similar heat input, less vapor generated in water-filled heat pipe. So, it was
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reported that Tev in water-filled heat pipe was greater than that in ethanol and methanol-filled
heat pipe [17] and as a result, the evaporator thermal resistance (Re) of water-filled heat pipe
was greater than the others.
On the other hand, the “liquid arrival mechanism” in [17] is different due to their different
surface tensions and the resulting capillary forces. Water surface tension is almost trice of
ethanol and methanol. It is predictable that the condensed liquid could return with better
performance in water-filled heat pipe. Warmer vapors received and condensed at the
condenser in water-filled heat pipe. Furthermore, the condensed liquid had less residence
time (due to the larger capillary force) for heat transfer. As a result, higher Tcv and Tav were
obtained in water-filled heat pipe [17].
5. Conclusion
In this paper, the heat transfer and the thermal resistance at different sections of the copper
heat pipe with the screen mesh wick was reported by measuring the core vapor and wall
temperatures at different operating conditions. The results showed that using butanol
solutions as working fluid improved the heat transfer performance of the heat pipe in
comparison with water. It was concluded that two mechanisms caused the heat transfer
enhancement in the heat pipe:
(1) “Vapor departing mechanism”: This affects the rate of vapor generation and its departure
from the evaporator section. Butanol solutions have lower heat of vaporization, as a result, at
the same heat input, more vapors generated in the heat pipe filled with butanol solution rather
than water. The greater mass of vapor generated in the evaporator leads to the greater heat
movement from the evaporator to the condenser. It means that vapor must be generated to
transfer heat from evaporator to condenser. Although the vapor core temperature in the water
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filled heat pipe was greater than that in butanol solutions, this cannot guarantee the heat
transfer enhancement.
It should also be mentioned that a lager mass flow rate of vapor generated in the evaporator
may result in larger pressure drop of the vapor flow from the evaporator to the condenser.
This may degrade the heat transfer performance of the heat pipe. Therefore, the generated
vapor flow rate must be optimized to have the best performance of the heat pipe. This idea
will be considered in our future works.
(2) “Liquid arrival mechanism”: This mechanism points out that how the condensed liquid in
the condenser section returns to the evaporator. The characteristics of the self-rewetting fluids
induce Marangoni flow, and as a result, the liquid arrival performs with better performance.
Some other mechanisms like gravitational force or capillary force can enhance the “Liquid
arrival mechanism”.
Nomenclature
Re
R
Q
Te,w
Tc,w
Te,v
Tc,v
hc
V
I
De,e
evaporator thermal resistance
total thermal resistance
input power
evaporator wall temperature
condenser wall temperature
vapor temperature at the evaporator section
vapor temperature at the condenser section
heat transfer coefficient at the condenser section
voltage
amperage
external diameter of the evaporator section
Di,e internal diameter of the evaporator section
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Le length of evaporator section
ke thermal conductivity of the evaporator section
Cp specific heat capacity
subscripts
e evaporator
c condenser
w wall
v vapor
Greek letters
α thermal diffusivity
ρ density
References:
[1] A. Sözen, M. Gürü, T. Menlik, M. Aktaş, Utilization of blast furnace slag nano-fluids in
two-phase closed thermo-syphon heat pipes for enhancing heat transfer, Experimental Heat
Transfer (2016).
[2] Y. Barzi, M. Assadi, Evaluation of a thermosyphon heat pipe operation and application in
a waste heat recovery system, Experimental Heat Transfer 28 (2015) 493-510.
[3] R. Savino, A. Cecere, R. Paola, Surface tension-driven flow in wickless heat pipes with
self-rewetting fluids, International Journal of Heat and Fluid Flow 30 (2009) 380-388.
[4] B. Poling, J. Prausnitz, J. O’Connell, The properties of gases and liquids, 5th
ed., McGraw-
Hill, New York, USA, 2000
[5] K. Armijo, V. Carey, An analytical and experimental study of heat pipe performance with
a working fluid exhibiting strong concentration Marangoni effects, International Journal of
Heat and Mass Transfer 64 (2013) 70-78.
[6] Y. Zhu, X. Cui, H. Han, S. Sun, The study on the difference of the start-up and heat-
transfer performance of the pulsating heat pipe with water- acetone mixtures, International
Journal of Heat and Mass Transfer 77 (2014) 834-842.
[7] R. Pachghare, A. Mahalle, Effect of pure and binary fluids on closed loop pulsating heat
pipe thermal performance, Procedia Engineering 51 (2013) 624-629.
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[8] A. Nuntaphan, J. Tiansuwan, T. Kiatsiriroat, Enhancement of heat transport in
thermosyphon air preheater at high temperature with binary working fluid: A case study of
TEG–water, Applied Thermal Engineering 22 (2002) 251-266.
[9] G. Burban, V. Ayel, A. Alexandre, P. Lagonotte, Y. Bertin, C. Romestant, Experimental
investigation of a pulsating heat pipe for hybrid vehicle applications, Applied Thermal
Engineering 50 (2013) 94-103.
[10] Kenneth M. Armijo, Van P. Carey, Prediction of binary mixture boiling heat transfer in
systems with strong Marangoni effects, Frontiers in Heat and Mass Transfer 2 (2010) 23003.
[11] R. Senthilkumar, S. Vaidyanathan, B. Sivaraman, Study of heat pipe performance using
an aqueous solution of n-Butanol, Indian Journal of Science and Technology 3 (2010) 702-
706.
[12] Hu, Yanxin, T. Liu, Xu. Li, S. Wang, Heat transfer enhancement of micro oscillating
heat pipes with self-rewetting fluid, International Journal of Heat and Mass Transfer 70
(2014) 496-503.
[13] R. Savino, N. Francescantonio, R. Fortezza, Y. Abe, Heat pipes with binary mixtures
and inverse Marangoni effects for microgravity applications, Acta Astronautica 61 (2007) 16-
26.
[14] R.C. Reid, J.M. Prausnitz, B.E. Poling, The Properties of Gases and Liquids, 4th ed.,
New York: McGraw-Hill, 1987.
[15] R.Monti, Physics of fluid in microgravity, First Edition, 2001, Taylor and Francis, New
York.
[16] R. Savino, A. Cecere, S. Van Vaerenbergh, Y. Abe, G. Pizzirusso, W. Tzevelecos, M.
Mojahed, Q. Galand, Some experimental progresses in the study of self-rewetting fluids for
the SELENE experiment to be carried in the Thermal Platform 1 hardware, Acta Astronautica
89 (2013) 179-188.
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[17] S.M. Peyghambarzadeh, S. Shahpouri, N. Aslanzadeh, M. Rahimnejad, Thermal
performance of different working fluids in a dual diameter circular heat pipe, Ain Shams
Engineering Journal 4 (2013) 855-861.
[18] J. Mustafa, S. Husain, M. Altamush Siddiqui, Experimental studies on natural
convection of water in a closed loop vertical annulus, Experimental Heat Transfer (2016).
[19] M. Sarafraz, F. Hormozi, S.M. Peyghambarzadeh, Thermal performance and efficiency
of a thermosyphon heat pipe working with a biologically ecofriendly nanofluid, International
Communications in Heat and Mass Transfer 57 (2014) 297-303.
[20] Y. Chang, C. Cheng, J. Wang, S. Chen, Heat pipe for cooling of electronic equipment,
Energy Conversion and Management 49 (2008) 3398-3404.
[21] G. Mousa, Effect of nanofluid concentration on the performance of circular heat pipe,
Ain Shams Engineering Journal 2 (2011) 63-69.
[22] R. Moffat, Describing the uncertainties in experimental results, Experimental Thermal
and Fluid Science 1 (1988) 3-17.
[23] ZhenHua Liu, QunZhi Zhu, Application of aqueous nanofluids in a horizontal mesh
heat pipe, Energy Conversion and Management 52(1) (2011) 292-300.
[24] Xiaowu Wang, Tang Yong, Chen Ping, Investigation into performance of a heat pipe
with micro grooves fabricated by extrusion–ploughing process, Energy Conversion and
Management 50 (5) (2009) 1384-1388.
[25] Lelun Jiang, Jintian Ling, Linzhen Jiang, Yong Tang, Yan Li, Wei Zhou, Jinwu Gao,
Thermal performance of a novel porous crack composite wick heat pipe, Energy
Conversion and Management 81 (2014) 10-18.
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(a)
(b)
Fig. 1. (a) Surface tension behavior in water and self-rewetting fluids (b) anomalous Marangoni
effect in self-rewetting fluids for some mixtures and at some range of temperature.

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Water
Circulator
Working
fluids
input Condenser
Vacuum pump
Data logger
DCPS
To vent
Fig. 2. A schematic presentation of the experimental setup
Fig. 3. Location of thermocouples in the heat pipe
: External : Internal

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Fig. 4. The surface tension of water / butanol solutions at different temperatures
5
10
15
20
25
30
35
40
18 28 38 48 58 68 78 88
Surfacetension(mN/m)
Temperature (°C)
Butanol solution 5 %wt [16]
Capilary method (Butanol solution 4 %wt) [This work]
Ring method (butanol solution 4 %wt) [This work]
Capilary method (Butanol solution 7 %wt) [This work]
Ring method (butanol solution 7 %wt) [This work]

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10
20
30
40
50
60
70
80
90
100
110
0 50 100 150 200 250 300 350 400
Tv(°C)
Length (mm)
Tc=15 °C
Tc=25 °C
Tc=35 °C
10
20
30
40
50
60
70
80
90
100
110
0 50 100 150 200 250 300 350 400
Tv(°C)
Length (mm)
Tc=15 °C
Tc=25 °C
Tc=35 °C
Evaporator Adiabatic Condenser
(a)
CondenserAdiabaticEvaporator
(b)

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Fig. 5. Variation of axial vapor core temperature at constant input heat flux (2400 W/m2
) for
different working fluids (a) water, (b) water / 4 %wt butanol, (c) water / 7 %wt butanol
10
20
30
40
50
60
70
80
90
100
110
0 50 100 150 200 250 300 350 400
Tv(°C)
Length (mm)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(c)

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10
20
30
40
50
60
70
80
90
100
110
0 50 100 150 200 250 300 350 400
Tw(°C)
Length (mm)
Tc=15 °C
Tc=25 °C
Tc=35 °C
10
20
30
40
50
60
70
80
90
100
110
0 50 100 150 200 250 300 350 400
Tw(°C)
Length (mm)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(a)
(b)

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Fig. 6. Variation of axial wall temperature at constant input heat flux (2400 W/m2
) for different
working fluids (a) water, (b) water / 4 %wt butanol, (c) water / 7 %wt butanol
10
20
30
40
50
60
70
80
90
100
110
0 50 100 150 200 250 300 350 400
Tw(°C)
Length (mm)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(c)

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Fig. 7. Effect of butanol concentration on the temperature at different sections of the heat pipe
20
30
40
50
60
70
80
90
0 2 4 6 8 10
T (oC)
Butanol concentration (wt.%)
Tev
Tav
Tcv
Q = 20 W
Tc = 15 oC

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0
0.02
0.04
0.06
0.08
0.1
0.12
0.14
0.16
0.18
0 0.5 1 1.5 2 2.5 3 3.5 4
he(kW/m2K)
q" (kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
0
0.2
0.4
0.6
0.8
1
1.2
0 0.5 1 1.5 2 2.5 3 3.5 4
he(kW/m2K)
q" (kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(a)
(b)

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Fig. 8. Variation of evaporator heat transfer coefficient against heat flux at different condenser
temperatures for (a) water, (b) water / 4 %wt butanol, (c) water / 7 %wt butanol
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
0 0.5 1 1.5 2 2.5 3 3.5 4
he(kW/m2K)
q" (kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(c)

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0
0.002
0.004
0.006
0.008
0.01
0.012
0 0.5 1 1.5 2 2.5 3 3.5
Re(°C/W)
q" (kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
0.004
0.006
0.008
0.01
0.012
0.014
0 0.5 1 1.5 2 2.5 3 3.5
Re(°C/W)
q" (kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(a)

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Fig. 9. Variation of evaporator thermal resistance against heat flux at different condenser
temperature for (a) water, (b) water / 4 %wt butanol, (c) water / 7 %wt butanol
0
0.002
0.004
0.006
0.008
0.01
0 0.5 1 1.5 2 2.5 3 3.5
Re(°C/W)
q'' (kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(c)

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0
0.001
0.002
0.003
0.004
0.005
0.006
0.007
0 0.5 1 1.5 2 2.5 3 3.5
Rc(°C/W)
q"(kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
0
0.0005
0.001
0.0015
0.002
0.0025
0.003
0.0035
0.004
0.0045
0 0.5 1 1.5 2 2.5 3 3.5
Rc(°C/W)
q"(kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(a)
(b)

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Fig. 10. Variation of condenser thermal resistance against heat flux at different condenser
temperature for (a) water, (b) water / 4 %wt butanol, (c) water / 7 %wt butanol
0
0.0004
0.0008
0.0012
0.0016
0.002
0.0024
0 0.5 1 1.5 2 2.5 3 3.5
Rc(°C/W)
q"(kW/m2)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(c)

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0
2
4
6
8
10
12
14
16
0 5 10 15 20 25 30
Rt(°C/W)
Q (W)
Tc=15 °C
Tc=25 °C
Tc=35 °C
0
2
4
6
8
10
12
0 5 10 15 20 25 30
Rt(°C/W)
Q (W)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(a)
(b)

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Fig. 11. Variation of heat pipe thermal resistance as a function of input heat flux at different
condenser temperature for (a) water, (b) water / 4 %wt butanol, (c) water / 7 %wt butanol
0
2
4
6
8
10
12
0 5 10 15 20 25 30
Rt(°C/W)
Q (W)
Tc=15 °C
Tc=25 °C
Tc=35 °C
(c)

Heat transfer enhancement using self-rewetting fluids

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