Thermal Performance of a Heat Pipe with Two Dissimilar Condensers ...

Journal of Applied Science and Engineering, Vol. 15, No. 2, pp. 123-129 (2012)

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Thermal Performance of a Heat Pipe with Two Dissimilar Condensers for a Medium-Temperature Thermal Storage System Min Kyu Park1 and Joon Hong Boo2* 1

2

Graduate School, Korea Aerospace University, School of Aerospace and Mechanical Engineering, Korea Aerospace University, 200-1 Hwajeon, Goyang, Gyeonggi-do, 412-791 Korea

Abstract An experimental study was conducted to investigate the thermal performance of a heat pipe having two dissimilar condenser sections which were subject to different boundary conditions. The first condenser dissipated heat to the surrounding air through annular fins by natural convection, while the second one was cooled by a liquid by forced convection at different temperatures. The container and the wick were made of stainless steel and the working fluid was Dowtherm-A, which is used for medium-temperature applications. The diameter and length of the heat pipe were 25.4 mm and 1 m, respectively. The maximum thermal load was 1 kW and the operating temperature of the heat pipe was around 250 °C. The liquid fill charge was adjusted so that the first condenser section might work as an evaporator when the original evaporator was inactive in a vertical configuration. The experimental results were analyzed in terms of thermal resistance and effective thermal conductivity against input heat flux and operating temperature. Key Words: Heat Pipe, Medium-High Temperature, Dissimilar Condensers, Experiment

1. Introduction The medium-temperature range for a heat pipe is normally between 550 and 750 K, based on the operating temperature [1]. Heat pipes have been considered as promising means for effective heat transfer in energy transport and storage systems operating in a mediumhigh temperature range, such as concentrated solar thermal energy systems. Although water is the best working fluid for heat pipes operating below 500 K unless electric insulation is required, it is not suitable to be used in the medium temperature range in general. This renders considerable difficulties in fabricating medium temperature heat pipes. Mercury being excluded due to its toxicity, there are only a few options, to date, for heat pipe working fluids *Corresponding author. E-mail: [email protected]

in the medium temperature range: synthetic fluids such as Flutec PP2 and PP9, and Dowtherm-A. Among these, Flutec fluids are usually employed for dielectric applications, and the maximum temperatures allowable for use with PP2 and PP9 are known to be 160 °C and 225 °C. On the other hand, Dowtherm-A, which is also known by the commercial name Thermex, is a eutectic mixture of diphenyl and diphenyl ether and can be used in the range of 150 to 400 °C (420 to 670 K). From the viewpoint of a usable temperature range, Dowtherm-A is the most useful working fluid for medium-temperature heat pipes [2]. Therefore, it is meaningful to investigate the performance characteristics of a Dowtherm-A heat pipe in various configurations and operating conditions. A recent experimental study on the fundamental aspects of Dowtherm-A heat pipes can be found in Park et al. [3] for a thermosyphon type with a single evaporator and condenser.

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Min Kyu Park and Joon Hong Boo

This study focused on the use of a heat pipe for a thermal energy storage system charging and discharging heat at 400 °C and 220 °C, respectively, in a typical solar thermal power plant. The proposed thermal energy storage system consisted of a solid-liquid phase-change material (PCM, hereafter), typically nitrates of alkali metals, and a penetrating heat pipe. The advantage of using a PCM is that it may contribute to a volume reduction of the thermal storage system utilizing the latent heat. In the study, the heat pipe interfaced with the heat source and sink at heat exchanging sections located below and above the PCM. During a heat charging mode, the heat pipe part interfacing with the heat source functioned as an evaporator and the portion embedded in the PCM worked as a condenser section. During the heat discharging mode however, the heat pipe part embedded in the PCM functioned as an evaporator and the part interfacing with the heat sink worked as a condenser. A similar configuration was investigated experimentally by Liu et al. [4] using a thermosyphon made of a 950-mm long copper pipe filled with acetone. Unfortunately, however, the result of their study is not applicable to medium-temperature devices since the heat source was below 100 °C and the melting temperature of the PCM was 52 °C (that of industrial paraffin wax). In an experimental study by Lee et al. [5] on a Dowtherm-A heat pipe with an internal screen wick, the working fluid charge was greater than the whole evaporator internal volume to ensure a startup during the heat discharging mode. The operating temperature of the heat pipe was nearly 320 °C and the effective thermal conductivity reached 6,000 W/m-K. However, their study was limited to the thermal performance of the heat pipe with a single condenser simulating a heat charging mode. This study is aimed at conducting an experimental study on a Dowtherm-A heat pipe with two dissimilar condenser sections subject to different boundary conditions which might be encountered in a medium-temperature application.

in the actual thermal storage system would be embedded in a PCM. For convenience of the experimental setup however, it was cooled by natural convection in ambient air. It was taken under consideration that typical PCMs in the medium temperature range normally exhibit very low thermal conductivity values (typically, 0.1 to 0.5 W/m-K), which can be approximated conservatively by that of air (typically, 0.05 W/m-K). The second condenser section was cooled by forced convection with liquid through a cooling block. Annular fins were attached only to the first condenser to enhance heat transfer. The heat pipe container and fins were made of stainless steel. The diameter and length of the heat pipe were 25.4 mm and 1 m, respectively. The lengths of the evaporator, the first condenser, and the second condenser were 200, 470, and 200 mm, respectively, with two adiabatic sections of 65 mm in between the three sections (See Figure 1). Heat was supplied to the evaporator by a ceramic mold type electric heater simulating the high-tempera-

2. Heat Pipe and Experimental Setup 2.1 Fabrication of a Dowtherm-A Heat Pipe The Dowtherm-A heat pipe was fabricated with two different condenser sections. The first condenser section

Figure 1. Schematic of the heat pipe with thermocouple locations.

Thermal Performance of a Heat Pipe with Two Dissimilar Condensers for a Medium-Temperature Thermal Storage System 125

ture steam in an actual solar thermal system. To simulate a heat discharging mode, heat had to be supplied to the Condenser 1 (in Figure 1) section while the Evaporator (in Figure 1) was kept insulated. To ensure a heat pipe startup during a heat discharging mode, the working fluid charge was 133.3 to 150.5 ml, which corresponded to 155 to 175% of the original evaporator (Evaporator in Figure 1) internal volume. These values corresponded to 23 to 32% of the internal volume of the evaporator (Condenser 1 in Figure 1) during a heat discharging mode, and 372 to 420% based on the total void volume of the wick. The specification of the heat pipe is summarized in Table 1.

cylindrical wall and for annular fin surfaces by natural convection. The heat loss was estimated by substracting the heats discharged through Condenser 1 and Condenser 2 from Qin. Temperature readings at 11 different locations along the axis of the heat pipe were measured to investigate the heat transport performance. In the following figures, every data point represented an average value of the measurements for ten minutes at least, after the operation reached a steady state for a given thermal load and cooling condition.

2.2 Experimental Setup and Experimental Methods The temperatures of the heat pipe wall were measured by K-type thermocouples at the locations depicted in Figure 1. A data aquisition system connected with a PC was used to measure, monitor, and store the temperature data from the thermocouples. The sampling period of the data acquisition was 12 channels per second. The A cooling block was attached to the heat pipe wall in the second condenser section (Condenser 2) at the top. The flow rate, and the coolant temperatures at the inlet and outlet of the cooling block were used to estimate the amount of heat actually transported through the heat pipe, Qrec.

Figures 2 and 3 compare steady-state axial temperature distributions of the heat pipe against fill charge and coolant inlet temperature variations, for the cases in which the Condenser 1 was cooled by natural convection and was insulated, while the heat inputs were the same as at 900 W. The abscissas of Figures 2 and 3 represent a normalized axial position, which was defined as the distance from the bottom end of the heat pipe, x, divided by the total length of the heat pipe, Lhp (See Figure 1). The operating temperature of the heat pipe was assumed to be the same as the wall temperature in the adiabatic section, as normally adopted in heat pipe studies. It was observed that the operating temperature of the heat pipe decreased with the fill charge amount for the same coolant inlet temperature. The heat pipe wall temperatures in the Condenser 1 as well as those in the evaporator increased as much as 25 to 50 °C when the Condenser 1 was inactive by insulation. In addition, the operating temperature re-

(1) where m& and Cp represent the mass flow rate and specific heat of the coolant, respectively. A series of experiments was conducted during this study against variations in the working fluid charge, heat input, and coolant inlet temperature. The combination of the last two was utilized to acquire a desired operating temperature for the heat pipe so that three different coolant inlet temperatures of 40, 60 and 80 °C were imposed for every heat input, Qin from the electric heater. Based on the comparison between Qin and Qrec, the heat loss of the setup was estimated to be less than 15% in most cases except when a dry-out occurred for a single-condenser (Condenser 2 only) operation, where the calculated heat loss exceeded 20%. The amount of heat discharged through Condenser 1 was estimated by combining two equations for a vertical

3. Results and Discussion

Table 1. Specification of the heat pipe Container (mm) Wick Working fluid

Length (mm)

Fin (mm)

Stainless steel 316L, O.D.: 25.4, Thickness: 2 Stainless steel 304, 2 layers of mesh No. 40 screen Dowtherm-A Fill charge: 372~420% (based on void volume of the wick) Evaporator: 200, Adiabatic section 1: 65, Condenser 1: 470, Adiabatic section 2: 65, Condenser 2: 200 Thickness: 1, Height: 32.5, Pitch: 10

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Figure 2. Axial temperature variation of the heat pipe as a function of coolant temperature (for a 900-W heat input, Condenser1 natural cooling).

Figure 4. Thermal resistance of the heat pipe as a function of heat flux and fill charge ratio (for a coolant inlet temperature of 80 °C).

Figure 3. Axial temperature variation of the heat pipe as a function of coolant temperature (for a 900-W heat input with Condenser1 insulated).

Figure 5. Thermal resistance of the heat pipe as a function of the operating temperature and fill charge ratio (for a coolant inlet tempreature of 80 °C).

mained constant, with a maximum of 5 °C fluctuation at 275 °C, when Condenser 1 was insulated, while the coolant temperature and fill charge ratio varied. Figures 4 and 5 represent the variation of thermal resistance as functions of the heat flux and the operating temperature, respectively. The results are shown only for a coolant inlet temperature of 80 °C since the highest operating temperatures were achieved among three different coolant inlet conditions. The thermal resistance of

the heat pipe in this study was calculated by the following equation, considering that two different condensers had to be combined properly to represent the average condenser surface temperature:

(2) where Tevp and Tcond represent the average wall tempera-


tures of the evaporator and condenser, respectively. Tcond was determined by the following equation. (3) where the subscripts cond 1 and cond 2 denote the values for Condenser 1 and Condenser 2, respectively. The thermal resistance reduced monotonically with the increase of heat flux in every case. The minimum thermal resistance of 0.12 °C/W was achieved for a heat pipe with dissimilar condensers with fill charge of 420% and operating temperature of 268.1 °C, which corresponded to an input heat flux of 69 kW/m2. When Condenser 1 was insulated, the minimum thermal resistance for the same heat flux was about 0.24 °C/W. In general, the thermal resistance value of the case in which Condenser 1 was insulated was less than a half of the value with insulation. The effective thermal conductivity, keff, of a heat pipe can be calculated by the following equation

discharging mode, where the original evaporator was kept inactive, the heat was input through the Condenser 1 in Figure 1, and the heat was removed only from Condenser 2. Figures 8 and 9 summarize the experimental results in terms of thermal resistance and effective thermal conductivity as a function of input heat flux. Both figures represent the results for a typical coolant inlet temperature of 80 °C. The effective thermal conductivity for the discharging mode exhibited a linear increase with input heat flux,

(4) where Ac is the cross-sectional area of the heat pipe, and Leff is an effective transport length calculated by Ladia + 0.5(Levp + Lcond 1 + Lcond 2), which was 0.9 m and 0.533 m for the heat pipes with single condenser and the one with dissimilar condensers, respectively. Figures 6 and 7 summarize the effective thermal conductivity as functions of heat flux and operating temperature, respectively. In general, the effective thermal conductivity of the heat pipe increased with heat flux and operating temperature. While it was obvious that the effective thermal conductivity increased with the fluid fill charge for a heat pipe with an inactive Condenser 1, no definite trends could be stated for a heat pipe with an active Condenser 1. However, the absolute value of the effective thermal conductivity of the heat pipe with an active Condenser 1 exhibited values more than twice as high. The maximum keff value was about 8,900 W/m-K, which corresponded to about 22 times of the thermal conductivity of commercial copper. The same liquid fill charge ratios and coolant inlet temperatures were imposed for the same heat pipe for the

Figure 6. Effective thermal conductivity of the heat pipe as a function of heat flux (for a coolant inlet temperature of 80 °C).

Figure 7. Effective thermal conductivity of the heat pipe as a function of the operating temperature (for a coolant inlet temperature of 80 °C).

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Figure 8. Thermal resistance of the heat pipe as a function of input heat flux (for coolant inlet temperature of 80 °C, heat discharging mode).

Figure 9. Effective thermal conductivity of the heat pipe as a function of input heat flux (for coolant inlet temperature of 80 °C, heat discharging mode).

and those for the fill charge ratio of 420% showed higher values than other fill charges. The maximum input heat flux before a dry-out occurred was 37.3 kW/m2, and the corresponding effective thermal conductivity was 5,200 W/m-K, which was 58% of the maximum value achieved for the charging mode simulation. The corresponding thermal resistance was about 25% higher than the minimum value for the charging mode.

sistance variation was insensitive with the liquid fill charge for the heat pipe with dissimilar condensers. (3) Despite a narrower operating range of input heat flux and a lower temperature, the heat pipe with two dissimilar condensers had a superior effective thermal conductivity to the heat pipe with a single condenser. (4) During the charging mode, dry-out occurred only for a heat pipe having an inactive Condenser 1 with a liquid fill charge of 372% at a heat flux of 66 kW/m2. Dry-out was not observed for heat pipes with higher liquid fill charges up to a heat flux of 75 kW/m2, which was the maximum heater capacity. For a heat pipe with dissimilar condensers, dry-out occurred for heat fluxes between 60 and 70 kW/m2. (5) For the discharging mode, a liquid fill charge of 420% exhibited the best performance among three different fill charges ratios.

4. Conclusion The following conclusions can be stated based on the observations derived from the experiments conducted in this study. (1) The Dowtherm-A heat pipe having two dissimilar condensers during a charging mode exhibited an effective thermal conductivity as much as twenty-two times that of commercial copper for the operating temperature near 270 °C. The heat pipe with an inactive Condenser 1 was able to reach the highest operating temperature of 322 °C, where the maximum effective thermal conductivity exhibited half of the above value. During the discharging mode it reduced to a lower value, which was still thirteen times higher than that of commercial copper. (2) For the heat pipe with an inactive Condenser 1, the thermal resistance reduced distinctly with the increase of working fluid fill charge. However, the thermal re-

Acknowledgement This work was supported by 2006 Korea Aerospace University faculty research grant. This work was also supported by the New & Renewable Energy Technology Development Program of the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant funded by the Korea Government Ministry of Knowledge and Economy under the Project Number 2005N-SO17-P-01.


Nomenclature A: C p: k: L: & m: Q: Rth: T: T:

area (m2) specific heat (kJ/kg-K) thermal conductivity (W/m-K) length (m) mass flow rate (kg/s) thermal load (W) thermal resistance (K/W) temperature (°C) average temperature (°C)

Subscripts cond: condenser cond 1: Condenser 1 (see Figure1) cond 2: Condenser 2 (see Figure1) e: exit eff: effective evp: evaporator hp: heat pipe i: inlet in: input rec: recovered

References [1] Faghri, A., Heat Pipe Science and Technology, Taylor and Francis, pp. 19-24 (1995). [2] Reay, D. and Kew, P., Heat Pipes, Elsevier, pp. 108-114 (2006). [3] Park, K. H., Lee, Y. S., Na, S. H. and Chang, K. C., “An Experimental Study on the Operating Characteristics of the Naphthalene and Dowtherm Heat Pipe,” Proc. KSME Annual Meeting, pp. 1966-1971 (2006). [4] Liu, Z., Wang, Z. and Ma, C., “An Experimental Study on Heat Transfer Characteristics of Heat Pipe Heat Exchanger with Latent Heat Storage. Part I: Charging Only and Discharging only Modes,” Energy Conversion and Management, Vol. 47, pp. 944-966 (2006). [5] Lee, S. K., Kwak, H. H., Boo, J. H., Kim, J. K. and Kang, Y. H., “Thermal Performance of a Dowtherm-A Heat Pipe for the Thermal Storage System at Medium-High Temperature,” Proc. KSME Fall Annual Conference, pp. 1645-1650 (2009).

Manuscript Received: Mar. 12, 2012 Accepted: May 18, 2012