Thermal Science & Engineering Vol.15 No.4 (2007)[Short Note]
Heat Transport Device with Phase Change Using
Two Parallel Tubes
Adina P. CIRTOG†,Sadanari MOCHIZUKI†, Akira MURATA† and Ioana IONEL‡

Abstract
In order to achieve high heat transport performance, a new heat transport device with phase change using
two parallel tubes has been developed. The device consisted of two different diameter parallel tubes, whichconnected an electrically heated evaporator and a water-cooled condenser. Experiments were conductedsystematically using water as working fluid. Time-dependent temperatures as well as pressures at the evaporator
and the condenser were continuously measured and recorded. The heat transport rate was obtained from thetemperature rise and mass flow rate of cooling water in condenser. The present study disclosed that a simple
two-different-diameter tube system connecting a heat source and a heat sink achieved re-circulating flow and it
gave an extremely high heat transport rate. It was also found in the present study that the orientation of thedevice had appreciable influence on the performance.
Key Words: Heat transport device, Two parallel tubes, Phase-change, Re-circulating flow, Inclination
angle
1. Introduction
The rapid development of electronic equipments leads
to an increase of energy consumption which demands anefficient cooling system. In order to achieve high heattransport characteristics, many kinds of heat transport
devices (HTD) with phase change have been developed.
Above all, heat pipes (HP) and thermosiphons are wellknown [1, 2]. The need for more compact and efficient
HTD than conventional types leads to the developmentof a new type HTD which is described in this paper [3].
The uniqueness of this new HTD exists in themechanism of fluid flow. The vapor and condensate flow
separately through two different parallel tubes realizing
one-way

re-circulation

flow

condenser

throughlarger-diameter

thefrom

evaporator
tubetoandcondenser to evaporator through the smaller-diametertube. This is in contrast to the conventional HPs where
the flows in opposing directions are coexisting in thesame tube causing the entrainment limit.
Fig. 1 Schematic of experimental setup (front view).

†‡

Received: Jul. 26, 2007 Editor: Hideo INABA
Dept. of Mechanical Systems Engineering, Tokyo University of A & T (Koganei, Tokyo 184-8588, Japan)

Faculty of Mechanical Engineering, “Politehnica” University of Timisoara (300223, Romania)- 241 -C 2007 The Heat Transfer Society of Japan○

Thermal Science & Engineering Vol.15 No.4 (2007)

2. Experimental setup
A schematic of experimental setup is presented in Fig. 1 andthe side view of the test section is shown in Fig. 2. The test
section consists of two different inner-diameter copper tubes(6.5 mm and 3.5 mm, respectively, and 215 mm in length)
which connect the evaporator (Ev.) and the condenser (Co.)
whose shape are cylindrical with 40 mm in diameter and 20mm in axial length. Before charging the working fluid (purewater) the test section was evacuated to around 2 kPa through avacuum pump. A given amount of working fluid (40% of thetotal inner volume of the test section) was charged using agraduated syringe. An electrical heater was placed on thebackside of the evaporator and connected to a regulated power
supply. A water cooling coil was installed inside the condenser
and cooling passages were grooved on the backside of thecondenser as it is seen in Fig. 2. Flow behavior was observed

Fig. 2 Representation of the inclination angle,through the sight glasses attached to the front side of both
evaporator and condenser and it was recorded by digital
high-speed video camera.
Thermocouples were used for measuring temperatures of (1)inlet and outlet cooling water at condenser (2) working fluid in
the evaporator and condenser and (3) the ceramic heater

the mass flow rate and the temperature rise of coolingwater flowing through the back side of condenser
chamber and through the copper coil installed inside. The
transported heat can be expressed as follows:

attached to the evaporator. Two high precision pressure sensors

Q&(-100 ~ 100 kPa) were connected to both the evaporator andcondenser. The cooling water flow rate was measured using astop watch and a graduate cylinder. The temperature and thepressure data were fed to a data acquisition card (DAQ) andthey were recorded and analyzed using the Lab View software.
Each value of Q& shown in the following figures was obtainedc p V& Tw,outTw,in(1)

where Tw,,in, Tw,,out and V& are the inlet, outlet
temperatures and the volume flow rate of the coolingwater, respectively.

from the average of 6000 readings (or more), during the steadystate, for an interval of minimum 10 minutes.
3. Procedure
The effect of various parameters such as (a)geometrical properties (diameter and length) of the tubes,(b) orientation of the device (c) amount of charged
working fluid (pure water), on the fluid flow and heattransport characteristics were considered.
In the present study a stress was put on the effect ofdevice orientation on the heat transport performancefrom evaporator to condenser. Experiments wereperformed under different orientations of device, withthe inclination angle, , as indicated in Fig. 2.
The heat transport rate Q& was directly determined from- 242 -Fig. 3. Overall thermal resistance.

C 2007 The Heat Transfer Society of Japan○

Thermal Science & Engineering Vol.15 No.4 (2007)

Fig. 3 shows the representation of the thermal
resistances through which the heat passes, from the point
where the heat is input into the system (the heater) to theone where it is released (the cooling water). The overall
thermal resistance, Rtotal, between the heater and thecooling water was calculated using the following
relationship:

RtotalTheater

Tw,in Tw,out

2 Q&(2)

4. Measurement uncertainty

Fig. 4 Heat transport rate vs. evaporator temperature(Tev) for different inclinations angle

The uncertainty was estimated using the method of Kline
and McClintock [5]. The error range of the measurements wasbetween 15% and 3%, for low and high heat input respectively.
The majority of the experimental errors were less than 10%.
Only at the lowest heat input the uncertainty was high as theerror range relative to the low temperature difference of thecooling water resulted in a higher percentage error.

the test section is not symmetrical with respect to thecenter plane as shown in Fig. 2. The contact area
between the liquid working fluid and the hot surface ofthe evaporator, and the one between the vapor and thecold surface of condenser, changes with the inclination
angle. This plays a certain role on the characteristic ofheat

5. Results and discussions
Fig. 4 shows the heat transport rate Q& versus
evaporator temperature Tev with orientation angle

transport

phenomenonandinfluences

theperformance of the present device. Both, the boiling andas athe condensation are important in the entire process oftransporting the heat efficiently.

parameter. The figure also shows the performance of a

The higher performance of the 120° case compared toconventional HP [4] with water and similar diameter andthe 60° case indicates that the condenser is morelength. From the figure one can see that:(1) Q& increases rapidly with an increase in Tev

dominant than the evaporator in the present system.

irrespective of the inclination angles;

against Tev with the inclination angles as parameter. In(2) Although the conventional HP gives higher Q& at
low Tev, the present HTD shows far better performance at

One can observe in Fig. 6 the results for Rtotal plotted
this figure it is shown that Rtotal decreases with anincrease in Tev and that for an evaporator temperature

higher Tev, with three to four times higher heat transport
rate than the conventional HP.
Fig. 5 shows Q& vs. for three different evaporator
temperatures, Tev. One can observe that, in all three cases,
the highest value of Q& is achieved for the verticalorientation (90°). The horizontal orientation (0°) gives
the minimum heat transport rate. For the same
evaporator temperature, a small increase in theinclination from 0° to 10° brings an abrupt rise in theheat transport rate, Q& being up to three times higher.
After Q& reaches the maximum at the vertical orientation(90°), the performance decreases with a further

increase in the inclination angle. The difference between
the 60° and 120° cases occurs because the structure of- 243 -Fig. 5 Heat transport rate Q& , vs. inclination angle

C 2007 The Heat Transfer Society of Japan○Thermal Science & Engineering Vol.15 No.4 (2007)6. Conclusions
The main results of this study are:a)

It was found that the one way recirculation wasmaintained to occur and this was consider to be
the main cause for the achievement of high heattransport rate.

b) The present device showed better performance
than conventional HP.
c)

The present device could transport around
800 W for an evaporator temperature being
below 100 °C.

d) The vertical orientation with the evaporator at
the bottom gave the highest performance.
Fig. 6 Total thermal resistance vs. evaporator
temperature, with as a parameter.

References

larger than 50 °C, in case of θ =90°, the values of thethermal resistances become as low as 0.2 °C/W. For thehorizontal orientation, the thermal resistance is muchhigher, irrespective of Tev .
The main reason for the high performance of thepresent device is the unique mechanism of fluid flow. By
providing two different-diameter tubes, the flows ofvapor and condensate are separated in each tube, and theshear force at the vapor-liquid interface which exists inthe conventional HPs is avoided. Therefore, theentrainment limit does not occur in the present HTD.

[1] Stauder, F.A. and McDonald, T.W., “Experimentalstudy of a two-ohase flow thermosiphon loop heatexchanger”, ASHREA Trans. 92, Pt.2A (1986),486-496.
[2] Zou, L. et.al., Experimental investigation onparallel connection module along the exhaustcurrent direction for gravity type heat pipes , Proc.
of 11th Int. Heat Pipe Conf., Vol.1 (1999), 159-163.
[3] Onishi, T. et al., Heat transport device consisting ofcapillary tubes with boiling driven heat transfer
mechanism inside, 1st Int. Symposium on Micro andNano Technology, Hawaii (2004).
[4] Japan Association for Heat Pipe ed., Heat Pipe(2001), 104, Nikkan Kogyo Shinbusha.
[5] Kline, S.J. and McClintock, F.A., “Describinguncertainties in single sample experiments”, Mech.
Eng. ASME, 75 (1953), 3-8.

- 244 -C 2007 The Heat Transfer Society of Japan○