- Email: [email protected]
Development of a single-phase thermosiphon for cold collection and storage of radiative cooling
Applied Energy 205 (2017) 1260–1269
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Development of a single-phase thermosiphon for cold collection and storage of radiative cooling
MARK
Dongliang Zhaoa, Christine Elizabeth Martinia, Siyu Jianga, Yaoguang Maa, Yao Zhaia, Gang Tanb, ⁎ Xiaobo Yina,c, Ronggui Yanga,d, a
Department of Mechanical Engineering, University of Colorado, Boulder, CO 80309, United States Department of Civil and Architectural Engineering, University of Wyoming, Laramie, WY 82071, United States c Materials Science and Engineering Program, University of Colorado, Boulder, CO 80309, United States d Buildings and Thermal Systems Center, National Renewable Energy Laboratory, 15013 Denver West Parkway, Golden, CO 80401, United States b
H I G H L I G H T S single-phase thermosiphon is proposed for cold collection and storage of radiative cooling. • AThermal performance and fluid flow characteristics of the thermosiphon are explored. • The feasibility using single-phase thermosiphon for cold collection and storage of radiative cooling is demonstrated. • Parameters thatofaffect thermosiphon performance are experimentally investigated. •
A R T I C L E I N F O
A B S T R A C T
Keywords: Single-phase thermosiphon Radiative cooling Natural convection Cold collection Cold storage
A single-phase thermosiphon is developed for cold collection and storage of radiative cooling. Compared to the conventional nocturnal radiative cooling systems that use an electric pump to drive the heat transfer fluid, the proposed single-phase thermosiphon uses the buoyancy force to drive heat transfer fluid. This solution does not require electricity, therefore improving the net gain of the radiative cooling system. A single-phase thermosiphon was built, which consists of a flat panel, a cold collection tank, a water return tube, and a water distribution tank. Considering that outdoor radiative cooling flux is constantly changing (i.e. uncontrollable), an indoor testing facility was developed to provide a controllable cooling flux (comparable to a radiative cooling flux of 100 W/m2) for the evaluation of thermosiphon performance. The testing apparatus is a chilled aluminum flat plate that has a controlled air gap separation relative to the flat panel surface of the thermosiphon to emulate radiative cooling. With an average of 105 W/m2 cooling flux, the 18 liters of water in the thermosiphon was cooled to an average temperature of 12.5 °C from an initial temperature of 22.2 °C in 2 h, with a cold collection efficiency of 96.8%. The results obtained have demonstrated the feasibility of using a single-phase thermosiphon for cold collection and storage of radiative cooling. Additionally, the effects of the thermosiphon operation conditions, such as tilt angle of the flat panel, initial water temperature, and cooling energy flux, on the performance have been experimentally investigated. Modular design of the single-phase thermosiphon gives flexibility for its scalability. A radiative cooling system with multiple thermosiphon modules is expected to play an important role in cooling buildings and power plant condensers.
1. Introduction All matter with a temperature higher than absolute 0 K emits thermal radiation. The atmospheric transmission window allows infrared radiation at a specific wavelength range (e.g. 8–14 µm) to pass directly to the Universe without intermediate absorption and re-emission [1]. This could be utilized as a strategy to obtain “free” radiative
⁎
cooling by infrared radiation to the Universe as an infinite-size lowtemperature reservoir, which could help address global energy and environmental challenges. Nocturnal radiative cooling has attracted considerable attention worldwide, particularly for building applications [2–8]. It has been estimated that nocturnal radiative cooling can save up to 11% of the power consumption for cooling buildings in tropical climates [3].
Corresponding author at: Department of Mechanical Engineering, University of Colorado, Boulder, CO 80309, United States. E-mail address: [email protected] (R. Yang).
http://dx.doi.org/10.1016/j.apenergy.2017.08.057 Received 21 January 2017; Received in revised form 28 May 2017; Accepted 10 August 2017 0306-2619/ © 2017 Elsevier Ltd. All rights reserved.
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Nomenclature
A c g h k m p Q Q1 Q2
R Re t T Th Tc T1
T2 u
area, m2 specific heat, J/(kg K) acceleration of gravity, m/s2 height, m thermal conductivity, W/(m K) mass, kg pressure, Pa energy flux, W/m2 cooling flux from the 2nd aluminum plate to the chilled aluminum plate calculated by the first method, W/m2 cooling flux from the 2nd aluminum plate to the chilled aluminum plate calculated by the second method, W/m2 thermal resistance, K/W Reynolds number, – time, s temperature, °C flat panel inlet temperature, °C flat panel outlet temperature, °C average temperature of the chilled aluminum plate, °C
average temperature of the 2nd aluminum plate, °C velocity vector, m/s
Greek symbols
α αT δ Δt ΔT θ μ ρ
thermal diffusivity, m2/s thermal expansion coefficient, 1/K air gap thickness, m time interval, s temperature difference, °C angle, ° dynamic viscosity, m2/s density, kg/m3
Subscripts
0 amb al loss w
Radiative cooling flux depends on many parameters such as surface emissivity of the material, surface temperature, sky cloudiness, ambient air temperature, and humidity. Previous research has been focused on evaluating the dependence of radiative cooling on geographic locations, weather conditions [9–13], how to efficiently use “free” radiative cooling [14–17], as well as developing new surface materials [18–22]. A few small-scale radiative cooling systems have been demonstrated [4,23–25]. These systems usually employ a heat transfer fluid circulation loop that harnesses radiative cooling. Commonly used heat transfer fluids are water [12,15,23] and air [26]. The measured nocturnal radiative cooling flux is within the range of 40–87 W/m2 [4,23–25,27,28]. For example, Meir et al. [23] constructed a polymerbased radiative cooling system that consists of a 5.3 m2 radiator and a 280-liter water tank and achieved an average cooling flux of 55 W/m2 at night. The authors concluded that, except for hot and humid midsummer months, the radiative cooling system is capable of covering the cooling needs for a single-family house with clear sky condition. Due to the low energy flux of radiative cooling, researchers have considered to combine this technique with other methods (e.g. evaporative cooling [29,30], convective cooling [31,32]) to achieve larger cooling capabilities. Farmahini-Farahani and Heidarinejad [33] used radiative cooling to pre-cool water to increase the effectiveness of an evaporative cooler. They observed a 9% (average) efficiency increase in comparison with a stand-alone evaporative cooler. Cui et al. [28] found 11–55% cooling power increase compared to radiative cooling alone by combining radiative cooling with convective cooling. Radiative cooling can also be used in combination with other noncooling technologies to achieve multifunctional energy systems. Eicker and Dalibard [2] developed a photovoltaic–thermal (PVT) system that can produce both electricity and cooling energy. The measured radiative cooling flux is 40–65 W/m2. Furthermore, researchers have found that nocturnal radiative cooling can be used in combination with daytime solar heating [25,34]. With a surface material that has high absorptivity in both the solar spectrum and the atmospheric transmission window, Hu et al. [25] developed a hybrid system that generates solar heating at daytime and radiative cooling at night. The system achieved 86.4% heating efficiency compared to a traditional solar heating collector, along with a maximum radiative cooling flux of 50.3 W/m2. Past research on radiative cooling was primarily for nocturnal use because solar absorbance of just a few percent in the daytime will counteract the radiative cooling effect [35]. However, the recent advancement in nanophotonic and meta-materials make daytime
reference state ambient aluminum cold loss water
radiative cooling possible, which is attracting more attention to this technology [5,36]. By using a multilayer nanophotonic material over a 20-cm diameter surface area, Raman et al. [37] achieved 40.1 W/m2 daytime radiative cooling flux under 850 W/m2 solar irradiance. A temperature about 4 °C lower than the ambient temperature is achieved in a small air chamber (25 cm in diameter, a few centimeters in thickness), which demonstrated that radiative cooling can be possible for daytime use with a delicate design. Recently, the authors of this work demonstrated that a cost-effective randomized glass-polymer hybrid metamaterial can be roll-to-roll fabricated with an average of 93 W/m2 cooling flux at noon time (11 am – 2 pm) under greater than 800 W/m2 solar irradiance [38]. An average cooling flux of 110 W/m2 has been demonstrated in a 3-day continuous measurement for both day and night. The notable advantage of this novel hybrid metamaterial is not only its relatively larger cooling flux compared to other state-ofthe-art materials, but more importantly this metamaterial can be fabricated in an economically feasible way by the roll-to-roll process, which is suitable for large-scale applications. While more work needs to be done to study the reliability and the lifetime of this novel, low-cost daytime radiative cooling metamaterial, a radiative cooling system can be built to run 24 h/7 days continuously. However, harnessing radiative cooling at a low energy flux, ∼100 W/m2 or less, is a challenge. In a conventional radiative cooling system, the collection and storage of cold energy requires energy input, such as using an electric pump to drive the heat transfer fluid. This brings down the net gain of radiative cooling system. For example, Meir et al. [23] used 60 W pumping power for a radiator with an area of 5.3 m2, which means electric energy consumption is about 11.3 W/m2 and accounts for ∼20% of the total cold gain. Also, it should be noted that the electric energy consumed is in a much higher quality than the cold gain according to the second law of thermodynamics. Therefore, it is desirable to minimize electricity use for the collection and storage of radiative cooling. In fact, due to the low energy flux of radiative cooling, the required mass flow rate of heat transfer fluid (e.g. water) to acquire the cold energy can be low. It is thus possible to use passive technique to realize cold collection and storage. A single-phase thermosiphon, which circulates a fluid based on temperature difference-induced natural convection, is very promising [39,40]. Basu et al. [41] presented a comprehensive review of analyses and applications of single-phase thermosiphons. Most literature work on single-phase thermosiphon are mainly focused on finding reliable modeling approaches, analyzing 1261
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2.2. Construction of a single-phase thermosiphon
dynamic stability of thermosiphon, as well as improving performance using nanofluids. The applications of single-phase thermosiphon are mostly for hot water collection (e.g. solar water heater [42]) and heat dissipation (e.g. cooling of nuclear reactor core [43,44]). To the best of the authors’ knowledge, the use of single-phase thermosiphon for cooling down a heat transfer fluid to a temperature lower than ambient has never been demonstrated. In addition, solar water heaters work under a solar irradiance at ∼1000 W/m2. It is very challenging to demonstrate a single-phase thermosiphon that can work under a very low energy flux, ∼100 W/m2 of radiative cooling. In this study, a single-phase thermosiphon module is developed to demonstrate passive cold collection and storage for radiative cooling technology. The modular design of the single-phase thermosiphon offers great flexibility for its scalability. A radiative cooling system consists of a few thermosiphon modules can be applied to cool buildings to lower the electricity bill. A large scale radiative cooling system can potentially be employed for supplemental cooling of power plant condensers to increase thermal-to-electric energy conversion efficiency [45]. The objectives of this study are: (1) to develop a thermosiphon testing apparatus that can provide controllable low cooling flux for indoor tests because outdoor radiative cooling is uncontrollable, (2) to experimentally demonstrate a single-phase thermosiphon that can cool down heat transfer fluid to a temperature lower than ambient under a cooling energy flux that is comparable to radiative cooling (i.e. ∼100 W/m2), and (3) to experimentally investigate the effect of operating conditions on the performance of thermosiphon, including the tilt angle of the thermosiphon flat panel, initial water temperature, and the cooling flux.
To construct a single-phase thermosiphon shown in Fig. 1c, a 1 m × 1 m × 0.006 m (L × W × H) twin-wall transparent polycarbonate panel is used as the flat panel for the thermosiphon. Inside the flat panel there are multiple parallel channels, each having a 5 mm × 10 mm (H × W) cross-sectional area. Polycarbonate panel is used because it shows long-term durability for outdoor greenhouse applications [46]. It is also economically feasible for large scale deployment of radiative cooling application. Square-shaped PVC tubes serve as the cold collection tank and water distribution tank while a cylindrical shaped PVC tube is used as the water return tube. The crosssectional side width of the cold collection tank and water distribution tank are 0.1 m and 0.05 m, respectively. The diameter of the water return tube is 0.025 m (1 inch). The entire thermosiphon is thermally insulated except for the top surface of the flat panel where radiative cooling flux is applied. The cold collection tank and water distribution tank are insulated by a 7.5-cm thick spray foam. The bottom of flat panel is insulated by 5-cm thick insulation board. The water return tube is insulated by 2.5-cm thick pipe insulation. The total volume of water filled in the thermosiphon is 18.0 l.
2.3. Testing apparatus for a single-phase thermosiphon Radiative cooling flux can be a complicated function of several parameters including surface emissivity of the radiative cooling
2. System description and experimental setup As shown in Fig. 1a, conventional radiative cooling systems usually require an electric pump to drive the heat transfer fluid, which consumes high-grade electric energy and reduces the net gain of the cooling system. A single-phase thermosiphon for cold collection and storage of radiative cooling is proposed in this work, as shown in Fig. 1b. The fluid flow in the single-phase thermosiphon is self-driven natural circulation by a temperature difference induced by radiative cooling without extra energy input. To test thermosiphon performance under controllable cooling flux conditions, a thermosiphon testing apparatus that can emulate the low energy flux of radiative cooling has been developed. 2.1. Workings principle of a single-phase thermosiphon As shown in Fig. 1b, the proposed single-phase thermosiphon consists of a flat panel, a cold collection tank, a water return tube, and a water distribution tank. In this study, water is filled in the thermosiphon to serve as the working fluid. To mimic the operation of radiative cooling, a flat chilled aluminum plate with a temperature lower than the thermosiphon temperature will be put on top of the flat panel with a controlled air gap separation to provide a controllable cooling flux of ∼100 W/m2 through heat conduction of the air gap. The flat panel is tilted at a few degrees (θ ) to create a height difference (h ) between its inlet and outlet (marked by red and blue dots in Fig. 1b respectively). Due to the cooling flux to the chilled aluminum plate, a temperature difference (i.e. Th−Tc ) is established across the inlet and outlet of the flat panel. When the water inside the flat panel is cooled down, it enters the cold collection tank, and pushes the relatively high temperature water out from the bottom of the cold collection tank. The relatively high temperature water then enters the water distribution tank through the water return tube. From the water distribution tank, water is evenly distributed back into the flat panel to be cooled down further. Apparently, radiative cooling-induced buoyancy drives a closed-loop fluid circulation inside the thermosiphon, with no extra electricity consumption.
Fig. 1. (a) Schematic of a conventional nocturnal radiative cooling system. An electric pump is used to drive the heat transfer fluid for cold collection of radiative cooling. The electricity consumption of the pump reduces the net gain of the radiative cooling system. (b) A cross-sectional diagram and (c) a 3-D illustration of the single-phase thermosiphon module proposed in this work for cold collection and storage of radiative cooling. The single-phase thermosiphon consists of four different parts: a flat panel, a cold collection tank, a water return tube, and a water distribution tank. Radiative cooling induces buoyancy that drives a close-loop fluid circulation inside the thermosiphon. No extra energy consumption is needed for cold collection and storage using the proposed singlephase thermosiphon module.
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solve energy equation). The flat panel was meshed using a Swept grid, while all other computational domain was meshed using a Tetrahedral grid. After grid independence check, a simulation mesh consisted of 277,000 elements has been employed. The initial state of water inside the thermosiphon was assumed to be stationary, i.e., with zero velocity, and water temperature was set to be uniform. A cooling flux comparable to radiative cooling was applied at the top surface of the thermosiphon flat panel. All other surfaces of the thermosiphon were thermally insulated. The thermal resistance values were calculated using the thermal conductivity and thickness of the insulation material.
material, sky cloudiness, surface temperature, ambient air temperature, and humidity. Outdoor radiative cooling flux with large variation cannot be directly applied to test a single-phase thermosiphon while it is in the development stage. Therefore, a testing apparatus with controllable cooling flux is developed in this work to characterize thermosiphon performance. Such a testing system emulating the controlled low energy flux of radiative cooling enables testing of the single-phase thermosiphon at indoor environment. The key component of the testing apparatus is a chilled aluminum plate that has the same surface area as the thermosiphon flat panel and is located directly above the flat panel of thermosiphon with a controllable air gap (Fig. 2). The aluminum plate is cooled down to ∼5 °C by circulating chilled water inside the parallel aluminum tubing which is attached to the top of the aluminum plate. High thermal conductivity thermal grease is used between the aluminum tubing and aluminum plate to ensure good thermal contact. Heat conduction from thermosiphon flat panel to chilled aluminum plate through air gap mimics the low energy flux of radiative cooling. The cooling flux is controlled at 37–150 W/m2 by changing the air gap separation from 8.4 mm to 1.9 mm. Fig. 3 shows both the schematic and the photo of the singlephase thermosiphon and its testing apparatus developed in this work. The chilled aluminum plate is located above the single-phase thermosiphon, and both are held by a support structure located underneath.
4. Results and discussion 4.1. Thermosiphon performance testing using chilled aluminum plate Before any testing of the single-phase thermosiphon, the chilled aluminum plate was tested to find out its temperature uniformity and its cooling flux as a function of air gap separations. Extensive tests on single-phase thermosiphon performance was then conducted using the chilled aluminum plate. 4.1.1. Temperature uniformity and cooling flux of chilled aluminum plate To correctly simulate radiative cooling flux, two performance indexes are investigated: (1) the temperature uniformity of the chilled aluminum plate, and (2) the cooling flux that can be obtained using the chilled aluminum plate. Eight (8) T-type thermocouples are embedded into the chilled aluminum plate for temperature measurement (denoted as AL1 to AL8 in Fig. 4a). The thermocouples are distributed at different locations with various distances from the aluminum tubing. Chilled water at ∼0 °C is circulated through the aluminum tubing to cool down the aluminum plate. Testing results are shown in Fig. 4b. The chilled aluminum plate is initially at room temperature, after about 15 min, the chilled aluminum plate reaches steady-state at 5.4 ± 0.5 °C. The temperate is very uniform, with less than 1 °C temperature difference across the whole 1 m × 1 m chilled plate. To obtain a cooling flux, a 2nd aluminum plate with the same dimensions as the chilled aluminum plate is placed underneath the chilled aluminum plate. In the transient test, the initial temperature of the 2nd aluminum plate is at room temperature. The chilled aluminum plate is at ∼5 °C. The two plates are separated by acrylic separators of specific thickness from 1.9 mm to 8.4 mm. The 2nd aluminum plate is also embedded with eight (8) thermocouples. The entire testing platform is insulated by polyisocyanurate rigid foam insulation board (indicated by yellow color in Fig. 5). The cooling flux from the 2nd aluminum plate to the chilled aluminum plate can be predicted by two different methods. The first method calculates the cooling flux by using air gap thermal resistance between chilled aluminum plate and 2nd aluminum plate. Due to the
3. Numerical simulation of single-phase thermosiphon performance Numerical simulations on three-dimensional transient fluid flow and heat transfer in a single-phase thermosiphon are performed using COMSOL Multiphysics. A numerical model for the single-phase thermosiphon has been built (thermosiphon geometry shown in Fig. 1c). The coupled three-dimensional Navier-Stokes equation and the energy equation are solved for the velocity and temperature fields based on the following assumptions: (1) Fluid flow is laminar due to the low flow velocity. This assumption is later confirmed by simulation results. (2) No slip boundary condition is applied at the interior walls of the thermosiphon module. (3) Boussinesq approximation is applied for the buoyancy-driven fluid flow, which suggests that fluid density difference is ignored except where they appear in terms multiplied by gravity of acceleration (see Eq. (3)). The Governing equations are as follows Continuity equation: (1)
∇·u = 0 where u is fluid flow velocity vector (m/s). Momentum equation:
ρ
∂u + ρ (u ·∇) u = −∇p + μ∇2 u + ρg ∂t
(2)
ρg = ρ0 g [1−αT (T −T0)]
(3) 3
where ∇p is the pressure force on element per unit volume (N/m ), ρ is fluid density (kg/m3), g is the acceleration due to gravity (m/s2), μ is dynamic viscosity (m2/s), αT is thermal expansion coefficient (1/ K ), and the subscript “0 ” is a reference state. Energy equation:
∂T + u·∇T = α∇2 T ∂t
Fig. 2. Schematic of the testing apparatus for single-phase thermosiphon by using a chilled aluminum plate to simulate radiative cooling. The aluminum plate is cooled down to ∼5 °C by circulating chilled water inside aluminum tubing. Heat conduction from thermosiphon flat panel to chilled aluminum plate through air gap mimics the low energy flux of radiative cooling. The cooling flux is controlled at 37–150 W/m2 by changing the air gap separation from 8.4 mm to 1.9 mm.
(4) 2
where α is thermal diffusivity (m /s), and T is temperature (K). A numerical model was developed in COMSOL 5.1. The Laminar Flow module (used to solve continuity and momentum equations) was coupled iteratively with the Heat Transfer in Fluids module (used to 1263
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Fig. 3. (a) Exploded schematic view of the entire experimental setup including a single-phase thermosiphon, a chilled aluminum plate, and a support structure. (b) Photo of the assembled experimental setup. Different from the exploded schematic view, experimental setup in this photo is well insulated by using spray foam and insulation boards. The red box at bottom right of this photo contains ice water. The ice water is circulated through the tubes attached to aluminum plate to maintain the chilled aluminum plate at ∼5 °C. (For interpretation of the references to colour in this figure legend, the reader is referred to the web version of this article.)
Q1 (t ) =
Rair =
T2−T1 Rair
(5)
δair kair ·A
(6)
where T1 and T2 are average temperatures of the chilled aluminum plate and the 2nd aluminum plate, respectively. Rair is the thermal resistance of the air gap. δair is the air gap thickness. kair is thermal conductivity of air and A is the total heat transfer area between the two aluminum plates. The second method is through using temperature change of 2nd aluminum plate. Cooling flux Q2 can be calculated as shown in Eq. (7).
Q2 (t ) =
cal mal ΔT2 Δt
(7) nd
where cal is aluminum heat capacity, mal is the mass of 2 aluminum plate, ΔT2 is the temperature change of the 2nd aluminum plate over time interval Δt . Testing results for using a 3.1 mm air gap separation are shown in Fig. 6. The changes of the averaged temperatures in the chilled aluminum plate and the 2nd aluminum plate are plotted in Fig. 6a. Both Q1 and Q2 are plotted in Fig. 6b. Results show that the cooling flux is a linear function of temperature difference between these two aluminum plates. We found that Q1 is 5–21% higher (depending on ΔT , see Fig. 6b) than Q2 which is due to thermal loss from the ambient (Q1 = Q2 + Qloss ). In fact, thermal loss exists throughout the test, therefore, Q2 is more appropriate to represent cooling flux when conducting tests for the thermosiphon. With the chilled aluminum plate maintained at 5 °C, the closer the 2nd aluminum plate temperature is to the ambient, the smaller the thermal loss. The cooling energy flux can be changed by varying the air gap separation. Tests were conducted under 5 different air gap separations: 1.9 mm, 3.1 mm, 4.1 mm, 6.2 mm, and 8.4 mm, respectively. Cooling fluxes as a function of temperature difference (ΔT ) are summarized in Table 1. For the thermosiphon tests, ΔT between the chilled aluminum plate and the flat panel of the thermosiphon is usually from 5 °C to 20 °C. Therefore, by adjusting the air gap separation between the chilled aluminum plate and the flat panel, a transient cooling flux
Fig. 4. (a) Eight (8) T-type thermocouples (i.e. AL1-AL8) are embedded into 8 small holes drilled at various locations of the chilled aluminum plate. (b) Temperature measurement shows great uniformity across the whole 1 m × 1 m chilled aluminum plate, with less than 1 °C difference.
small separation and low emissivity (∼0.05) of aluminum plate, this method assumes heat transfer mechanism through the air gap is pure conduction by neglecting both convection and radiation. With very small area coverage of the acrylic separators (neglected), transient cooling flux Q1 is given by:
Fig. 5. Transient cooling flux test using the chilled aluminum plate with a 2nd aluminum plate. The 2nd aluminum plate has the same dimensions as the chilled aluminum plate. The chilled aluminum plate is placed a few millimeters (1.9–8.4 mm) above the 2nd aluminum plate where the air gap separation is controls the cooling flux. Cooling flux from the 2nd aluminum plate to the chilled aluminum plate can be calculated by using Eqs. (5) and (7).
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Fig. 6. (a) Change of the averaged temperatures of the chilled aluminum plate and the 2nd aluminum plate during a 30-min test. Temperature are uniformly distributed on both aluminum plates, which can be observed from the error bar. (b) The transient cooling energy flux from the 2nd aluminum plate to the chilled aluminum plate (Q1 , predicted by Eq. (5)), and cold gain inside the 2nd aluminum plate (Q2 , predicted by Eq. (7)). The discrepancy between Q1 and Q2 is due to cold loss to the ambient.
estimated based on the total water mass in the thermosiphon, radiative cooling flux, the targeted temperature drop, and the estimated thermal loss to the ambient. Eighteen (18) T-type thermocouples were placed inside the thermosiphon to measure the local water temperatures, among which nine (9) were inside the polycarbonate flat panel (PC1-PC9, see Fig. 7a), and nine (9) were inside the cold collection tank (CCT1-CCT9). Fig. 7b shows a cross-sectional view of the cold collection tank with thermocouple locations. PC1, PC2 and PC3 are located at the inlet of the flat panel, while PC7, PC8 and PC9 are located at the outlet of the flat panel. Temperature change at the inlet (i.e. average (PC1, PC2, PC3), midway (i.e. average (PC4, PC5, PC6) and outlet (i.e. average (PC7, PC8, PC9) of the flat panel are plotted in Fig. 7c. Observed evidence (marked by black circles) shows that water inside the thermosiphon is self-circulating. The average temperature at the inlet, midway and outlet of flat panel all experienced a small dip, but in a time sequence. This is because at the beginning the water inside the thermosiphon is stationary, when the water inside the flat panel was cooled down by the chilled aluminum plate, a temperature difference was established to drive the water flow, the relatively cool water started to move downward to the cold collection tank, while the relatively warm water from water distribution tank filled into the flat panel, which resulted in a temperature increase. Since it takes time for the relatively warm water to flow through the flat panel from inlet to outlet, the relatively warm water continuously affected flat panel’s inlet, midway and outlet temperatures in a time sequence. The temperature difference between the inlet and outlet of the flat panel is defined as the working temperature difference of the thermosiphon. Fig. 7d shows the transient working temperature difference of the thermosiphon. The working temperature difference increased rapidly to 8.5 °C at the beginning (i.e. the “start-up” period), then it became relatively stable (i.e. “steady-state”). The “start-up” period lasts about 25 min. At the “steady-state”, the working temperature difference decreased gradually. The reason of the decreasing trend is because the cooling flux from the flat panel to the chilled aluminum plate decreases with time due to the lowering of the flat panel temperature. Lower cooling flux results in lower fluid flow velocity, which results in a smaller working temperature difference. Reasonable agreement has been achieved between experimental results and numerical data. Some minor discrepancy is also observed at the “quasi-steady-state” of the experiment, which is possibly because the roughness of polycarbonate flat panel and PVC water return tube are not accounted for in the numerical model, and the wall roughness has some effect on fluid flow.
Table 1 Cooling flux from the 2nd aluminum plate to the chilled aluminum plate as a function of air gap separations. Air gap separation (mm)
Cooling flux (W/m2)
1.9 3.1 4.1 6.2 8.4
Q2 Q2 Q2 Q2 Q2
= = = = =
15.46ΔT −13.8 12.42ΔT −17.6 8.42ΔT −13.5 6.43ΔT −13.3 4.96ΔT −12.4
between 37 W/m2 and 150 W/m2 can be achieved. This part of the work demonstrates that the effectiveness of using heat conduction through a small air gap to simulate the low flux from radiative cooling (∼100 W/m2). Transient and average cooling flux can be readily calculated during the thermosiphon tests by measuring the transient temperature of the thermosiphon flat panel.
4.1.2. Thermosiphon performance A case study that demonstrates the thermal performance and fluid flow characteristics of the single-phase thermosiphon was carried out. Both experimental and numerical approaches were employed and results were compared. In this case study, the flat panel was tilted at 10°. The air gap separation between the chilled aluminum plate and the flat panel of thermosiphon was set at 1.9 mm. The testing room temperature is ∼21 °C. At the experiment preparation stage, the chilled aluminum plate was cooled down and the single-phase thermosiphon was filled with 23 °C water. To avoid cooling effects at this preparation stage, the chilled aluminum plate and the thermosiphon were separated at a distance greater than 10 cm with thermal insulation board in between. Once the chilled aluminum plate temperature reached ∼5 °C (with ± 0.5 °C temperature uniformity), it was then placed on top of the thermosiphon by using a few 1.9-mm thick acrylic separators (10mm in diameter). The transient test began right after the chilled aluminum plate was set up. For outdoor application, once the thermosiphon is cooled down, radiative cooling flux decreases due to the lower surface temperature. To maintain a relatively high radiative cooling flux, thermosiphon operation time needs to be controlled. At the end of a cooling process, the cooled water in thermosiphon will be removed and the thermosiphon will be refilled with relatively high temperature water. In this study, the thermosiphon testing period was controlled to be 2 h, which was 1265
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Fig. 7. (a) A 3-dimensional schematic showing the locations of thermocouples inside the flat panel, (b) a cross-sectional view through the middle of the cold collection tank showing thermocouple locations, (c) temperature change at the inlet, midway and outlet of the flat panel during a 2-h thermosiphon cooling test, (d) thermosiphon working temperature difference change during the 2-h cooling test, (e) average temperature change inside the cold collection tank during the 2-h thermosiphon test, (f) energy flux from the flat panel to the chilled aluminum plate during the 2-h thermosiphon test. Average energy flux for the 2-h test was 105 W/m2.
thermosiphon cold collection is calculated to be 733.9 kJ/ 758.1 kJ = 96.8%. The average temperature of the 18-liter water inside the thermosiphon is cooled down to 12.5 °C from 22.2 °C.
For both numerical data and experimental results, a little bit of overshoot for the working temperature difference at the end of “start-up” period is observed, which suggests that the thermosiphon fluid flow is capable of self-adjusting per different energy fluxes. Good agreement has been achieved between numerical and experimental results for temperature change inside the cold collection tank (Fig. 7e). Energy balance analysis is then conducted. Fig. 7f shows that the cooling energy flux decreased throughout the 2-h period. Larger cooling energy flux during the beginning of the test is due to the larger temperature difference between the flat panel to the chilled aluminum plate. The integration of cooling energy flux over the 2-h period gives the total energy flow from the thermosiphon to the chilled aluminum plate of 758.1 kJ. Average cooling flux for the 2-h period is 105 W/m2. Using initial and final temperatures and volume of each part of the thermosiphon (detailed in Table 2), the total cold storage inside the thermosiphon is calculated to be 733.9 kJ. The efficiency of
Table 2 Breakdown of cold storage inside the single-phase thermosiphon.
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Component
Flat panel
Water distribution tank
Water return tube
Cold collection tank
Volume (L) Initial temperature (°C) Final temperature (°C) Cold storage (kJ)
4.5 20.8
2.3 22.5
0.7 21.6
10.5 22.8
9.6
15.1
13.1
13.1
211.1
71.1
25.3
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A numerical model was built in COMSOL Multiphysics to predict fluid flow characteristics. Due to the low energy flux, fluid flow velocity inside the flat panel is very slow (in the order of 10−4 m/s). Fluid flow velocity inside the flat panel changes with time and location as plotted in Fig. 8a. It is observed that fluid velocity at all locations across the flat panel increased rapidly at the beginning and reached a maximum (at about 1.2 × 10−3 m/s) by the end of 25-min “start-up” period. After that, the thermosiphon runs into the “quasi-steady-state”, and the velocity decreased gradually due to the decreasing of the cooling flux. It is also observed that the fluid flow velocity is maximum at the center of the flat panel always. Water velocity at the edge of the flat panel is about 75% of the velocity at the center. The reason is because water flow at the center of the flat panel has the shortest flow path, and thus, the smallest flow resistance, for larger velocity. The Reynolds numbers inside the flat panel and the water return tube versus time are plotted in Fig. 8b. The Reynolds number inside the water return tube is larger because the water return tube has a smaller cross-sectional area and higher fluid flow velocity. However, both Reynolds numbers are still within laminar region, which suggest that fluid flow inside the thermosiphon is laminar at all times, and this validates the assumption made for the simulation.
8.5
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Flat Panel Tilt Angle (o) Fig. 9. Working temperature difference and the cold collection tank temperature drop of the thermosiphon after 2-h operation as a function of flat panel tilt angle. A larger tilt angle gives a larger cold collection tank temperature drop, but a smaller working temperature difference.
temperature difference decreases when the tilt angle increases. Note that the thermosiphon working temperature difference is the final-state temperature difference at the end of 2-h testing period. The reason is because a larger tilt angle gives larger height difference between flat panel inlet and outlet, which requires smaller working temperature difference. Also, larger tilt angle gives larger cold collection tank temperature drop. The average cooling fluxes for the 6°, 8°, 10°, 12° and14° tile angles were 122 W/m2, 129 W/m2, 132 W/m2, 137 W/m2, and 146 W/m2, respectively. A larger tile angle results in a larger average cooling flux. This is due to larger water mass flow rate inside the thermosiphon caused by larger buoyancy at the larger tilt angle, which is also confirmed by numerical results. By the end of 2-h cooling periods, the mass flow rates at the center of flat panel are 5.4 × 10−4 m/s, 5.8 × 10−4 m/s, 6.3 × 10-4 m/s, 7.0 × 10-4 m/s, 7.8 × 10-4 m/s for tilt angles 6°, 8°, 10°, 12°, 14°, respectively. Compared to the 6° tilt angle, the cold collection tank temperature drop increased by 23.7% for the 14° tilt angle.
4.2. Parametric study of single-phase thermosiphon 4.2.1. Effect of flat panel tilt angle on thermosiphon performance A larger tilt angle of the flat plat results in a larger height difference between the inlet and the outlet of the flat panel (see Fig. 1b), giving a greater buoyancy force. For an outdoor radiative cooling application, the final tilt angle will be determined along with other parameters that affect thermosiphon performance, such as radiation shape factor from the thermosiphon to the sky and daytime solar absorption. On one hand, a smaller tilt angle will give a larger radiative shape factor to the sky. On the other hand, to minimize solar absorption at the daytime, the thermosiphon will be tilted to facing the north in northern hemisphere. A larger tilt angle will give a smaller solar irradiance and thus is preferred. The final application of the thermosiphon will dictate the optimal operating condition. The thermosiphon was tested under different tilt angles from 6° to 14°, with 2° intervals to quantify the effect of tilt angle on thermosiphon performance. A tilt angle between 6° and 14° offers sufficient height difference for thermosiphon fluid flow, along with enough radiation shape factor from the thermosiphon to the sky. All tests used 3.1 mm air gap separation with an initial thermosiphon temperature at ∼29 °C for a period of 2-h. Fig. 9 shows that the thermosiphon working
4.2.2. Effect of initial water temperature on thermosiphon performance Keep the thermosiphon working at a relatively high temperature gives a higher radiative cooling flux. To quantify the effect of initial water temperature on thermosiphon performance, the thermosiphon
Fig. 8. (a) Fluid flow velocity inside flat panel versus time and location, (b) Reynolds numbers for fluid flow in the flat panel and the water return tube. Both Reynolds numbers are within the laminar region.
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4.2.3. Effect of cooling flux on thermosiphon performance The operation of the thermosiphon is strongly affected by weather conditions. In cloudy weather conditions, radiative cooling can be blocked by clouds and cooling flux is greatly reduced. Therefore, it is desirable to investigate the thermosiphon performance under different cooling fluxes. Five different cooling fluxes were tested, corresponding to 5 difference air gap separations (i.e. 1.9 mm, 3.1 mm, 4.1 mm, 6.2 mm and 8.4 mm) between the chilled aluminum plate and the flat panel. These 5 air gap separations were selected because they can provide an average cooling flux comparable to different levels of radiative cooling flux. The average cooling fluxes for these 5 air gap separations were obtained at 144 W/m2, 132 W/m2, 111 W/m2, 96 W/ m2, and 74 W/m2 respectively. Initial temperatures inside the thermosiphon were all set to ∼29 °C and the flat panel angles were all 10°. All tests were 2-h long. With increasing air gap separation, both the temperature drop of the cold collection tank and the working temperature difference of the final-state thermosiphon decreases due to the reduced average cooling flux (Fig. 11). The reason is that lower cooling flux results in lower fluid flow velocity inside thermosiphon, which in turn reduces the flow resistance. A lower flow resistance requires a smaller working temperature difference to drive the flow. A lower cooling flux also results in a smaller temperature drop in the cold collection tank due to energy balance. From the 8.4 mm air gap test results, the final-state cooling flux was 65 W/m2 and the thermosiphon is still working, which suggests that the thermosiphon can work on partially cloudy days.
Initial Temperature (oC)
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Fig. 10. Working temperature difference and the cold collection tank temperature drop of thermosiphon at the final state as a function of the initial water temperature. Both cold collection tank temperature drop and thermosiphon working temperature difference increase with initial temperature.
5. Conclusions A single-phase thermosiphon and its testing apparatus has been developed for cold collection and storage of radiative cooling. This is the very first demonstration for cooling down liquid inside the singlephase thermosiphon to a temperature lower than ambient induced by a very small cooling flux, ∼100 W/m2. Both experimental and numerical methods have been used to characterize the thermal performance and flow characteristics of the thermosiphon. The main conclusions that can be drawn from this study are the following:
• The
Air Gap Separation (mm)
•
Fig. 11. Working temperature difference and cold collection tank temperature drop of thermosiphon at the final state as a function of the air gap separation (i.e. cooling flux). Both cold collection tank temperature drop and thermosiphon working temperature difference decreased with increasing air gap separation (i.e. decreasing cooling flux).
was tested under 5 different initial temperatures, 24.4 °C, 26.6 °C, 27.6 °C, 29.1 °C, and 33.7 °C, respectively. These initial temperatures are selected because for either building cooling application or supplemental cooling application for power plant condensers, the water that needs to be cooled has a temperature within this range [45,47]. Air gap separations are all set at 3.1 mm, and the tilt angles of the flat panel are all set at 10°. All tests are 2-h long. A larger initial temperature gives a larger cooling flux through the air gap. The average cooling fluxes for these 5 different initial temperatures were 108 W/m2, 120 W/m2, 123 W/m2, 132 W/m2, and 158 W/m2, respectively. Both the working temperature difference and cold collection tank temperature drop of the thermosiphon increases with the initial temperature (Fig. 10). By increasing the initial temperature from 24.4 °C to 33.7 °C, the old collection tank temperature drop increased from 8.5 °C to 14.5 °C. For outdoor applications, the radiative cooling system operating time needs to be optimized to achieve both a targeted cold water temperature and a relatively high radiative cooling flux.
• • •
thermosiphon testing apparatus (i.e. the chilled aluminum plate) can provide transient cooling flux between 37 W/m2 and 150 W/m2, which is at the same level as radiative cooling flux. With an average cooling flux of 105 W/m2, 18 liters of water in the thermosiphon was cooled to an average temperature of 12.5 °C from an initial temperature of 22.2 °C in 2 h, with a cold collection efficiency of 96.8%. The testing room temperature is ∼21 °C. This has demonstrated the feasibility of a single-phase thermosiphon for cold collection and storage of radiative cooling. This also demonstrated that the thermosiphon can cool down a heat transfer fluid to a temperature lower than ambient by using a cooling flux comparable to radiative cooling. A larger flat panel tilt angle, higher initial water temperature, and larger cooling flux are desirable for achieving larger cold collection tank temperature drop while the final application of the thermosiphon will dictate the optimal operating conditions. At a lower cooling flux testing condition (65 W/m2), which simulates a partial cloudy day, the thermosiphon can still work, but with a smaller cold collection and storage capability. A radiative cooling system that has cooling capacity in kW-scale or even in MW-scale can be achieved by using multiple thermosiphon modules connected.
Acknowledgements The authors acknowledge the financial support of this work from the US Department of Energy’s ARPA-E (Contract No. DE-AR0000580). 1268
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Development of a single-phase thermosiphon for cold collection and storage of radiative cooling
MARK
Dongliang Zhaoa, Christine Elizabeth Martinia, Siyu Jianga, Yaoguang Maa, Yao Zhaia, Gang Tanb, ⁎ Xiaobo Yina,c, Ronggui Yanga,d, a
Department of Mechanical Engineering, University of Colorado, Boulder, CO 80309, United States Department of Civil and Architectural Engineering, University of Wyoming, Laramie, WY 82071, United States c Materials Science and Engineering Program, University of Colorado, Boulder, CO 80309, United States d Buildings and Thermal Systems Center, National Renewable Energy Laboratory, 15013 Denver West Parkway, Golden, CO 80401, United States b
H I G H L I G H T S single-phase thermosiphon is proposed for cold collection and storage of radiative cooling. • AThermal performance and fluid flow characteristics of the thermosiphon are explored. • The feasibility using single-phase thermosiphon for cold collection and storage of radiative cooling is demonstrated. • Parameters thatofaffect thermosiphon performance are experimentally investigated. •
A R T I C L E I N F O
A B S T R A C T
Keywords: Single-phase thermosiphon Radiative cooling Natural convection Cold collection Cold storage
A single-phase thermosiphon is developed for cold collection and storage of radiative cooling. Compared to the conventional nocturnal radiative cooling systems that use an electric pump to drive the heat transfer fluid, the proposed single-phase thermosiphon uses the buoyancy force to drive heat transfer fluid. This solution does not require electricity, therefore improving the net gain of the radiative cooling system. A single-phase thermosiphon was built, which consists of a flat panel, a cold collection tank, a water return tube, and a water distribution tank. Considering that outdoor radiative cooling flux is constantly changing (i.e. uncontrollable), an indoor testing facility was developed to provide a controllable cooling flux (comparable to a radiative cooling flux of 100 W/m2) for the evaluation of thermosiphon performance. The testing apparatus is a chilled aluminum flat plate that has a controlled air gap separation relative to the flat panel surface of the thermosiphon to emulate radiative cooling. With an average of 105 W/m2 cooling flux, the 18 liters of water in the thermosiphon was cooled to an average temperature of 12.5 °C from an initial temperature of 22.2 °C in 2 h, with a cold collection efficiency of 96.8%. The results obtained have demonstrated the feasibility of using a single-phase thermosiphon for cold collection and storage of radiative cooling. Additionally, the effects of the thermosiphon operation conditions, such as tilt angle of the flat panel, initial water temperature, and cooling energy flux, on the performance have been experimentally investigated. Modular design of the single-phase thermosiphon gives flexibility for its scalability. A radiative cooling system with multiple thermosiphon modules is expected to play an important role in cooling buildings and power plant condensers.
1. Introduction All matter with a temperature higher than absolute 0 K emits thermal radiation. The atmospheric transmission window allows infrared radiation at a specific wavelength range (e.g. 8–14 µm) to pass directly to the Universe without intermediate absorption and re-emission [1]. This could be utilized as a strategy to obtain “free” radiative
⁎
cooling by infrared radiation to the Universe as an infinite-size lowtemperature reservoir, which could help address global energy and environmental challenges. Nocturnal radiative cooling has attracted considerable attention worldwide, particularly for building applications [2–8]. It has been estimated that nocturnal radiative cooling can save up to 11% of the power consumption for cooling buildings in tropical climates [3].
Corresponding author at: Department of Mechanical Engineering, University of Colorado, Boulder, CO 80309, United States. E-mail address: [email protected] (R. Yang).
http://dx.doi.org/10.1016/j.apenergy.2017.08.057 Received 21 January 2017; Received in revised form 28 May 2017; Accepted 10 August 2017 0306-2619/ © 2017 Elsevier Ltd. All rights reserved.
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Nomenclature
A c g h k m p Q Q1 Q2
R Re t T Th Tc T1
T2 u
area, m2 specific heat, J/(kg K) acceleration of gravity, m/s2 height, m thermal conductivity, W/(m K) mass, kg pressure, Pa energy flux, W/m2 cooling flux from the 2nd aluminum plate to the chilled aluminum plate calculated by the first method, W/m2 cooling flux from the 2nd aluminum plate to the chilled aluminum plate calculated by the second method, W/m2 thermal resistance, K/W Reynolds number, – time, s temperature, °C flat panel inlet temperature, °C flat panel outlet temperature, °C average temperature of the chilled aluminum plate, °C
average temperature of the 2nd aluminum plate, °C velocity vector, m/s
Greek symbols
α αT δ Δt ΔT θ μ ρ
thermal diffusivity, m2/s thermal expansion coefficient, 1/K air gap thickness, m time interval, s temperature difference, °C angle, ° dynamic viscosity, m2/s density, kg/m3
Subscripts
0 amb al loss w
Radiative cooling flux depends on many parameters such as surface emissivity of the material, surface temperature, sky cloudiness, ambient air temperature, and humidity. Previous research has been focused on evaluating the dependence of radiative cooling on geographic locations, weather conditions [9–13], how to efficiently use “free” radiative cooling [14–17], as well as developing new surface materials [18–22]. A few small-scale radiative cooling systems have been demonstrated [4,23–25]. These systems usually employ a heat transfer fluid circulation loop that harnesses radiative cooling. Commonly used heat transfer fluids are water [12,15,23] and air [26]. The measured nocturnal radiative cooling flux is within the range of 40–87 W/m2 [4,23–25,27,28]. For example, Meir et al. [23] constructed a polymerbased radiative cooling system that consists of a 5.3 m2 radiator and a 280-liter water tank and achieved an average cooling flux of 55 W/m2 at night. The authors concluded that, except for hot and humid midsummer months, the radiative cooling system is capable of covering the cooling needs for a single-family house with clear sky condition. Due to the low energy flux of radiative cooling, researchers have considered to combine this technique with other methods (e.g. evaporative cooling [29,30], convective cooling [31,32]) to achieve larger cooling capabilities. Farmahini-Farahani and Heidarinejad [33] used radiative cooling to pre-cool water to increase the effectiveness of an evaporative cooler. They observed a 9% (average) efficiency increase in comparison with a stand-alone evaporative cooler. Cui et al. [28] found 11–55% cooling power increase compared to radiative cooling alone by combining radiative cooling with convective cooling. Radiative cooling can also be used in combination with other noncooling technologies to achieve multifunctional energy systems. Eicker and Dalibard [2] developed a photovoltaic–thermal (PVT) system that can produce both electricity and cooling energy. The measured radiative cooling flux is 40–65 W/m2. Furthermore, researchers have found that nocturnal radiative cooling can be used in combination with daytime solar heating [25,34]. With a surface material that has high absorptivity in both the solar spectrum and the atmospheric transmission window, Hu et al. [25] developed a hybrid system that generates solar heating at daytime and radiative cooling at night. The system achieved 86.4% heating efficiency compared to a traditional solar heating collector, along with a maximum radiative cooling flux of 50.3 W/m2. Past research on radiative cooling was primarily for nocturnal use because solar absorbance of just a few percent in the daytime will counteract the radiative cooling effect [35]. However, the recent advancement in nanophotonic and meta-materials make daytime
reference state ambient aluminum cold loss water
radiative cooling possible, which is attracting more attention to this technology [5,36]. By using a multilayer nanophotonic material over a 20-cm diameter surface area, Raman et al. [37] achieved 40.1 W/m2 daytime radiative cooling flux under 850 W/m2 solar irradiance. A temperature about 4 °C lower than the ambient temperature is achieved in a small air chamber (25 cm in diameter, a few centimeters in thickness), which demonstrated that radiative cooling can be possible for daytime use with a delicate design. Recently, the authors of this work demonstrated that a cost-effective randomized glass-polymer hybrid metamaterial can be roll-to-roll fabricated with an average of 93 W/m2 cooling flux at noon time (11 am – 2 pm) under greater than 800 W/m2 solar irradiance [38]. An average cooling flux of 110 W/m2 has been demonstrated in a 3-day continuous measurement for both day and night. The notable advantage of this novel hybrid metamaterial is not only its relatively larger cooling flux compared to other state-ofthe-art materials, but more importantly this metamaterial can be fabricated in an economically feasible way by the roll-to-roll process, which is suitable for large-scale applications. While more work needs to be done to study the reliability and the lifetime of this novel, low-cost daytime radiative cooling metamaterial, a radiative cooling system can be built to run 24 h/7 days continuously. However, harnessing radiative cooling at a low energy flux, ∼100 W/m2 or less, is a challenge. In a conventional radiative cooling system, the collection and storage of cold energy requires energy input, such as using an electric pump to drive the heat transfer fluid. This brings down the net gain of radiative cooling system. For example, Meir et al. [23] used 60 W pumping power for a radiator with an area of 5.3 m2, which means electric energy consumption is about 11.3 W/m2 and accounts for ∼20% of the total cold gain. Also, it should be noted that the electric energy consumed is in a much higher quality than the cold gain according to the second law of thermodynamics. Therefore, it is desirable to minimize electricity use for the collection and storage of radiative cooling. In fact, due to the low energy flux of radiative cooling, the required mass flow rate of heat transfer fluid (e.g. water) to acquire the cold energy can be low. It is thus possible to use passive technique to realize cold collection and storage. A single-phase thermosiphon, which circulates a fluid based on temperature difference-induced natural convection, is very promising [39,40]. Basu et al. [41] presented a comprehensive review of analyses and applications of single-phase thermosiphons. Most literature work on single-phase thermosiphon are mainly focused on finding reliable modeling approaches, analyzing 1261
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2.2. Construction of a single-phase thermosiphon
dynamic stability of thermosiphon, as well as improving performance using nanofluids. The applications of single-phase thermosiphon are mostly for hot water collection (e.g. solar water heater [42]) and heat dissipation (e.g. cooling of nuclear reactor core [43,44]). To the best of the authors’ knowledge, the use of single-phase thermosiphon for cooling down a heat transfer fluid to a temperature lower than ambient has never been demonstrated. In addition, solar water heaters work under a solar irradiance at ∼1000 W/m2. It is very challenging to demonstrate a single-phase thermosiphon that can work under a very low energy flux, ∼100 W/m2 of radiative cooling. In this study, a single-phase thermosiphon module is developed to demonstrate passive cold collection and storage for radiative cooling technology. The modular design of the single-phase thermosiphon offers great flexibility for its scalability. A radiative cooling system consists of a few thermosiphon modules can be applied to cool buildings to lower the electricity bill. A large scale radiative cooling system can potentially be employed for supplemental cooling of power plant condensers to increase thermal-to-electric energy conversion efficiency [45]. The objectives of this study are: (1) to develop a thermosiphon testing apparatus that can provide controllable low cooling flux for indoor tests because outdoor radiative cooling is uncontrollable, (2) to experimentally demonstrate a single-phase thermosiphon that can cool down heat transfer fluid to a temperature lower than ambient under a cooling energy flux that is comparable to radiative cooling (i.e. ∼100 W/m2), and (3) to experimentally investigate the effect of operating conditions on the performance of thermosiphon, including the tilt angle of the thermosiphon flat panel, initial water temperature, and the cooling flux.
To construct a single-phase thermosiphon shown in Fig. 1c, a 1 m × 1 m × 0.006 m (L × W × H) twin-wall transparent polycarbonate panel is used as the flat panel for the thermosiphon. Inside the flat panel there are multiple parallel channels, each having a 5 mm × 10 mm (H × W) cross-sectional area. Polycarbonate panel is used because it shows long-term durability for outdoor greenhouse applications [46]. It is also economically feasible for large scale deployment of radiative cooling application. Square-shaped PVC tubes serve as the cold collection tank and water distribution tank while a cylindrical shaped PVC tube is used as the water return tube. The crosssectional side width of the cold collection tank and water distribution tank are 0.1 m and 0.05 m, respectively. The diameter of the water return tube is 0.025 m (1 inch). The entire thermosiphon is thermally insulated except for the top surface of the flat panel where radiative cooling flux is applied. The cold collection tank and water distribution tank are insulated by a 7.5-cm thick spray foam. The bottom of flat panel is insulated by 5-cm thick insulation board. The water return tube is insulated by 2.5-cm thick pipe insulation. The total volume of water filled in the thermosiphon is 18.0 l.
2.3. Testing apparatus for a single-phase thermosiphon Radiative cooling flux can be a complicated function of several parameters including surface emissivity of the radiative cooling
2. System description and experimental setup As shown in Fig. 1a, conventional radiative cooling systems usually require an electric pump to drive the heat transfer fluid, which consumes high-grade electric energy and reduces the net gain of the cooling system. A single-phase thermosiphon for cold collection and storage of radiative cooling is proposed in this work, as shown in Fig. 1b. The fluid flow in the single-phase thermosiphon is self-driven natural circulation by a temperature difference induced by radiative cooling without extra energy input. To test thermosiphon performance under controllable cooling flux conditions, a thermosiphon testing apparatus that can emulate the low energy flux of radiative cooling has been developed. 2.1. Workings principle of a single-phase thermosiphon As shown in Fig. 1b, the proposed single-phase thermosiphon consists of a flat panel, a cold collection tank, a water return tube, and a water distribution tank. In this study, water is filled in the thermosiphon to serve as the working fluid. To mimic the operation of radiative cooling, a flat chilled aluminum plate with a temperature lower than the thermosiphon temperature will be put on top of the flat panel with a controlled air gap separation to provide a controllable cooling flux of ∼100 W/m2 through heat conduction of the air gap. The flat panel is tilted at a few degrees (θ ) to create a height difference (h ) between its inlet and outlet (marked by red and blue dots in Fig. 1b respectively). Due to the cooling flux to the chilled aluminum plate, a temperature difference (i.e. Th−Tc ) is established across the inlet and outlet of the flat panel. When the water inside the flat panel is cooled down, it enters the cold collection tank, and pushes the relatively high temperature water out from the bottom of the cold collection tank. The relatively high temperature water then enters the water distribution tank through the water return tube. From the water distribution tank, water is evenly distributed back into the flat panel to be cooled down further. Apparently, radiative cooling-induced buoyancy drives a closed-loop fluid circulation inside the thermosiphon, with no extra electricity consumption.
Fig. 1. (a) Schematic of a conventional nocturnal radiative cooling system. An electric pump is used to drive the heat transfer fluid for cold collection of radiative cooling. The electricity consumption of the pump reduces the net gain of the radiative cooling system. (b) A cross-sectional diagram and (c) a 3-D illustration of the single-phase thermosiphon module proposed in this work for cold collection and storage of radiative cooling. The single-phase thermosiphon consists of four different parts: a flat panel, a cold collection tank, a water return tube, and a water distribution tank. Radiative cooling induces buoyancy that drives a close-loop fluid circulation inside the thermosiphon. No extra energy consumption is needed for cold collection and storage using the proposed singlephase thermosiphon module.
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solve energy equation). The flat panel was meshed using a Swept grid, while all other computational domain was meshed using a Tetrahedral grid. After grid independence check, a simulation mesh consisted of 277,000 elements has been employed. The initial state of water inside the thermosiphon was assumed to be stationary, i.e., with zero velocity, and water temperature was set to be uniform. A cooling flux comparable to radiative cooling was applied at the top surface of the thermosiphon flat panel. All other surfaces of the thermosiphon were thermally insulated. The thermal resistance values were calculated using the thermal conductivity and thickness of the insulation material.
material, sky cloudiness, surface temperature, ambient air temperature, and humidity. Outdoor radiative cooling flux with large variation cannot be directly applied to test a single-phase thermosiphon while it is in the development stage. Therefore, a testing apparatus with controllable cooling flux is developed in this work to characterize thermosiphon performance. Such a testing system emulating the controlled low energy flux of radiative cooling enables testing of the single-phase thermosiphon at indoor environment. The key component of the testing apparatus is a chilled aluminum plate that has the same surface area as the thermosiphon flat panel and is located directly above the flat panel of thermosiphon with a controllable air gap (Fig. 2). The aluminum plate is cooled down to ∼5 °C by circulating chilled water inside the parallel aluminum tubing which is attached to the top of the aluminum plate. High thermal conductivity thermal grease is used between the aluminum tubing and aluminum plate to ensure good thermal contact. Heat conduction from thermosiphon flat panel to chilled aluminum plate through air gap mimics the low energy flux of radiative cooling. The cooling flux is controlled at 37–150 W/m2 by changing the air gap separation from 8.4 mm to 1.9 mm. Fig. 3 shows both the schematic and the photo of the singlephase thermosiphon and its testing apparatus developed in this work. The chilled aluminum plate is located above the single-phase thermosiphon, and both are held by a support structure located underneath.
4. Results and discussion 4.1. Thermosiphon performance testing using chilled aluminum plate Before any testing of the single-phase thermosiphon, the chilled aluminum plate was tested to find out its temperature uniformity and its cooling flux as a function of air gap separations. Extensive tests on single-phase thermosiphon performance was then conducted using the chilled aluminum plate. 4.1.1. Temperature uniformity and cooling flux of chilled aluminum plate To correctly simulate radiative cooling flux, two performance indexes are investigated: (1) the temperature uniformity of the chilled aluminum plate, and (2) the cooling flux that can be obtained using the chilled aluminum plate. Eight (8) T-type thermocouples are embedded into the chilled aluminum plate for temperature measurement (denoted as AL1 to AL8 in Fig. 4a). The thermocouples are distributed at different locations with various distances from the aluminum tubing. Chilled water at ∼0 °C is circulated through the aluminum tubing to cool down the aluminum plate. Testing results are shown in Fig. 4b. The chilled aluminum plate is initially at room temperature, after about 15 min, the chilled aluminum plate reaches steady-state at 5.4 ± 0.5 °C. The temperate is very uniform, with less than 1 °C temperature difference across the whole 1 m × 1 m chilled plate. To obtain a cooling flux, a 2nd aluminum plate with the same dimensions as the chilled aluminum plate is placed underneath the chilled aluminum plate. In the transient test, the initial temperature of the 2nd aluminum plate is at room temperature. The chilled aluminum plate is at ∼5 °C. The two plates are separated by acrylic separators of specific thickness from 1.9 mm to 8.4 mm. The 2nd aluminum plate is also embedded with eight (8) thermocouples. The entire testing platform is insulated by polyisocyanurate rigid foam insulation board (indicated by yellow color in Fig. 5). The cooling flux from the 2nd aluminum plate to the chilled aluminum plate can be predicted by two different methods. The first method calculates the cooling flux by using air gap thermal resistance between chilled aluminum plate and 2nd aluminum plate. Due to the
3. Numerical simulation of single-phase thermosiphon performance Numerical simulations on three-dimensional transient fluid flow and heat transfer in a single-phase thermosiphon are performed using COMSOL Multiphysics. A numerical model for the single-phase thermosiphon has been built (thermosiphon geometry shown in Fig. 1c). The coupled three-dimensional Navier-Stokes equation and the energy equation are solved for the velocity and temperature fields based on the following assumptions: (1) Fluid flow is laminar due to the low flow velocity. This assumption is later confirmed by simulation results. (2) No slip boundary condition is applied at the interior walls of the thermosiphon module. (3) Boussinesq approximation is applied for the buoyancy-driven fluid flow, which suggests that fluid density difference is ignored except where they appear in terms multiplied by gravity of acceleration (see Eq. (3)). The Governing equations are as follows Continuity equation: (1)
∇·u = 0 where u is fluid flow velocity vector (m/s). Momentum equation:
ρ
∂u + ρ (u ·∇) u = −∇p + μ∇2 u + ρg ∂t
(2)
ρg = ρ0 g [1−αT (T −T0)]
(3) 3
where ∇p is the pressure force on element per unit volume (N/m ), ρ is fluid density (kg/m3), g is the acceleration due to gravity (m/s2), μ is dynamic viscosity (m2/s), αT is thermal expansion coefficient (1/ K ), and the subscript “0 ” is a reference state. Energy equation:
∂T + u·∇T = α∇2 T ∂t
Fig. 2. Schematic of the testing apparatus for single-phase thermosiphon by using a chilled aluminum plate to simulate radiative cooling. The aluminum plate is cooled down to ∼5 °C by circulating chilled water inside aluminum tubing. Heat conduction from thermosiphon flat panel to chilled aluminum plate through air gap mimics the low energy flux of radiative cooling. The cooling flux is controlled at 37–150 W/m2 by changing the air gap separation from 8.4 mm to 1.9 mm.
(4) 2
where α is thermal diffusivity (m /s), and T is temperature (K). A numerical model was developed in COMSOL 5.1. The Laminar Flow module (used to solve continuity and momentum equations) was coupled iteratively with the Heat Transfer in Fluids module (used to 1263
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Fig. 3. (a) Exploded schematic view of the entire experimental setup including a single-phase thermosiphon, a chilled aluminum plate, and a support structure. (b) Photo of the assembled experimental setup. Different from the exploded schematic view, experimental setup in this photo is well insulated by using spray foam and insulation boards. The red box at bottom right of this photo contains ice water. The ice water is circulated through the tubes attached to aluminum plate to maintain the chilled aluminum plate at ∼5 °C. (For interpretation of the references to colour in this figure legend, the reader is referred to the web version of this article.)
Q1 (t ) =
Rair =
T2−T1 Rair
(5)
δair kair ·A
(6)
where T1 and T2 are average temperatures of the chilled aluminum plate and the 2nd aluminum plate, respectively. Rair is the thermal resistance of the air gap. δair is the air gap thickness. kair is thermal conductivity of air and A is the total heat transfer area between the two aluminum plates. The second method is through using temperature change of 2nd aluminum plate. Cooling flux Q2 can be calculated as shown in Eq. (7).
Q2 (t ) =
cal mal ΔT2 Δt
(7) nd
where cal is aluminum heat capacity, mal is the mass of 2 aluminum plate, ΔT2 is the temperature change of the 2nd aluminum plate over time interval Δt . Testing results for using a 3.1 mm air gap separation are shown in Fig. 6. The changes of the averaged temperatures in the chilled aluminum plate and the 2nd aluminum plate are plotted in Fig. 6a. Both Q1 and Q2 are plotted in Fig. 6b. Results show that the cooling flux is a linear function of temperature difference between these two aluminum plates. We found that Q1 is 5–21% higher (depending on ΔT , see Fig. 6b) than Q2 which is due to thermal loss from the ambient (Q1 = Q2 + Qloss ). In fact, thermal loss exists throughout the test, therefore, Q2 is more appropriate to represent cooling flux when conducting tests for the thermosiphon. With the chilled aluminum plate maintained at 5 °C, the closer the 2nd aluminum plate temperature is to the ambient, the smaller the thermal loss. The cooling energy flux can be changed by varying the air gap separation. Tests were conducted under 5 different air gap separations: 1.9 mm, 3.1 mm, 4.1 mm, 6.2 mm, and 8.4 mm, respectively. Cooling fluxes as a function of temperature difference (ΔT ) are summarized in Table 1. For the thermosiphon tests, ΔT between the chilled aluminum plate and the flat panel of the thermosiphon is usually from 5 °C to 20 °C. Therefore, by adjusting the air gap separation between the chilled aluminum plate and the flat panel, a transient cooling flux
Fig. 4. (a) Eight (8) T-type thermocouples (i.e. AL1-AL8) are embedded into 8 small holes drilled at various locations of the chilled aluminum plate. (b) Temperature measurement shows great uniformity across the whole 1 m × 1 m chilled aluminum plate, with less than 1 °C difference.
small separation and low emissivity (∼0.05) of aluminum plate, this method assumes heat transfer mechanism through the air gap is pure conduction by neglecting both convection and radiation. With very small area coverage of the acrylic separators (neglected), transient cooling flux Q1 is given by:
Fig. 5. Transient cooling flux test using the chilled aluminum plate with a 2nd aluminum plate. The 2nd aluminum plate has the same dimensions as the chilled aluminum plate. The chilled aluminum plate is placed a few millimeters (1.9–8.4 mm) above the 2nd aluminum plate where the air gap separation is controls the cooling flux. Cooling flux from the 2nd aluminum plate to the chilled aluminum plate can be calculated by using Eqs. (5) and (7).
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Fig. 6. (a) Change of the averaged temperatures of the chilled aluminum plate and the 2nd aluminum plate during a 30-min test. Temperature are uniformly distributed on both aluminum plates, which can be observed from the error bar. (b) The transient cooling energy flux from the 2nd aluminum plate to the chilled aluminum plate (Q1 , predicted by Eq. (5)), and cold gain inside the 2nd aluminum plate (Q2 , predicted by Eq. (7)). The discrepancy between Q1 and Q2 is due to cold loss to the ambient.
estimated based on the total water mass in the thermosiphon, radiative cooling flux, the targeted temperature drop, and the estimated thermal loss to the ambient. Eighteen (18) T-type thermocouples were placed inside the thermosiphon to measure the local water temperatures, among which nine (9) were inside the polycarbonate flat panel (PC1-PC9, see Fig. 7a), and nine (9) were inside the cold collection tank (CCT1-CCT9). Fig. 7b shows a cross-sectional view of the cold collection tank with thermocouple locations. PC1, PC2 and PC3 are located at the inlet of the flat panel, while PC7, PC8 and PC9 are located at the outlet of the flat panel. Temperature change at the inlet (i.e. average (PC1, PC2, PC3), midway (i.e. average (PC4, PC5, PC6) and outlet (i.e. average (PC7, PC8, PC9) of the flat panel are plotted in Fig. 7c. Observed evidence (marked by black circles) shows that water inside the thermosiphon is self-circulating. The average temperature at the inlet, midway and outlet of flat panel all experienced a small dip, but in a time sequence. This is because at the beginning the water inside the thermosiphon is stationary, when the water inside the flat panel was cooled down by the chilled aluminum plate, a temperature difference was established to drive the water flow, the relatively cool water started to move downward to the cold collection tank, while the relatively warm water from water distribution tank filled into the flat panel, which resulted in a temperature increase. Since it takes time for the relatively warm water to flow through the flat panel from inlet to outlet, the relatively warm water continuously affected flat panel’s inlet, midway and outlet temperatures in a time sequence. The temperature difference between the inlet and outlet of the flat panel is defined as the working temperature difference of the thermosiphon. Fig. 7d shows the transient working temperature difference of the thermosiphon. The working temperature difference increased rapidly to 8.5 °C at the beginning (i.e. the “start-up” period), then it became relatively stable (i.e. “steady-state”). The “start-up” period lasts about 25 min. At the “steady-state”, the working temperature difference decreased gradually. The reason of the decreasing trend is because the cooling flux from the flat panel to the chilled aluminum plate decreases with time due to the lowering of the flat panel temperature. Lower cooling flux results in lower fluid flow velocity, which results in a smaller working temperature difference. Reasonable agreement has been achieved between experimental results and numerical data. Some minor discrepancy is also observed at the “quasi-steady-state” of the experiment, which is possibly because the roughness of polycarbonate flat panel and PVC water return tube are not accounted for in the numerical model, and the wall roughness has some effect on fluid flow.
Table 1 Cooling flux from the 2nd aluminum plate to the chilled aluminum plate as a function of air gap separations. Air gap separation (mm)
Cooling flux (W/m2)
1.9 3.1 4.1 6.2 8.4
Q2 Q2 Q2 Q2 Q2
= = = = =
15.46ΔT −13.8 12.42ΔT −17.6 8.42ΔT −13.5 6.43ΔT −13.3 4.96ΔT −12.4
between 37 W/m2 and 150 W/m2 can be achieved. This part of the work demonstrates that the effectiveness of using heat conduction through a small air gap to simulate the low flux from radiative cooling (∼100 W/m2). Transient and average cooling flux can be readily calculated during the thermosiphon tests by measuring the transient temperature of the thermosiphon flat panel.
4.1.2. Thermosiphon performance A case study that demonstrates the thermal performance and fluid flow characteristics of the single-phase thermosiphon was carried out. Both experimental and numerical approaches were employed and results were compared. In this case study, the flat panel was tilted at 10°. The air gap separation between the chilled aluminum plate and the flat panel of thermosiphon was set at 1.9 mm. The testing room temperature is ∼21 °C. At the experiment preparation stage, the chilled aluminum plate was cooled down and the single-phase thermosiphon was filled with 23 °C water. To avoid cooling effects at this preparation stage, the chilled aluminum plate and the thermosiphon were separated at a distance greater than 10 cm with thermal insulation board in between. Once the chilled aluminum plate temperature reached ∼5 °C (with ± 0.5 °C temperature uniformity), it was then placed on top of the thermosiphon by using a few 1.9-mm thick acrylic separators (10mm in diameter). The transient test began right after the chilled aluminum plate was set up. For outdoor application, once the thermosiphon is cooled down, radiative cooling flux decreases due to the lower surface temperature. To maintain a relatively high radiative cooling flux, thermosiphon operation time needs to be controlled. At the end of a cooling process, the cooled water in thermosiphon will be removed and the thermosiphon will be refilled with relatively high temperature water. In this study, the thermosiphon testing period was controlled to be 2 h, which was 1265
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Fig. 7. (a) A 3-dimensional schematic showing the locations of thermocouples inside the flat panel, (b) a cross-sectional view through the middle of the cold collection tank showing thermocouple locations, (c) temperature change at the inlet, midway and outlet of the flat panel during a 2-h thermosiphon cooling test, (d) thermosiphon working temperature difference change during the 2-h cooling test, (e) average temperature change inside the cold collection tank during the 2-h thermosiphon test, (f) energy flux from the flat panel to the chilled aluminum plate during the 2-h thermosiphon test. Average energy flux for the 2-h test was 105 W/m2.
thermosiphon cold collection is calculated to be 733.9 kJ/ 758.1 kJ = 96.8%. The average temperature of the 18-liter water inside the thermosiphon is cooled down to 12.5 °C from 22.2 °C.
For both numerical data and experimental results, a little bit of overshoot for the working temperature difference at the end of “start-up” period is observed, which suggests that the thermosiphon fluid flow is capable of self-adjusting per different energy fluxes. Good agreement has been achieved between numerical and experimental results for temperature change inside the cold collection tank (Fig. 7e). Energy balance analysis is then conducted. Fig. 7f shows that the cooling energy flux decreased throughout the 2-h period. Larger cooling energy flux during the beginning of the test is due to the larger temperature difference between the flat panel to the chilled aluminum plate. The integration of cooling energy flux over the 2-h period gives the total energy flow from the thermosiphon to the chilled aluminum plate of 758.1 kJ. Average cooling flux for the 2-h period is 105 W/m2. Using initial and final temperatures and volume of each part of the thermosiphon (detailed in Table 2), the total cold storage inside the thermosiphon is calculated to be 733.9 kJ. The efficiency of
Table 2 Breakdown of cold storage inside the single-phase thermosiphon.
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Component
Flat panel
Water distribution tank
Water return tube
Cold collection tank
Volume (L) Initial temperature (°C) Final temperature (°C) Cold storage (kJ)
4.5 20.8
2.3 22.5
0.7 21.6
10.5 22.8
9.6
15.1
13.1
13.1
211.1
71.1
25.3
426.5
Working Temperature Difference (oC)
A numerical model was built in COMSOL Multiphysics to predict fluid flow characteristics. Due to the low energy flux, fluid flow velocity inside the flat panel is very slow (in the order of 10−4 m/s). Fluid flow velocity inside the flat panel changes with time and location as plotted in Fig. 8a. It is observed that fluid velocity at all locations across the flat panel increased rapidly at the beginning and reached a maximum (at about 1.2 × 10−3 m/s) by the end of 25-min “start-up” period. After that, the thermosiphon runs into the “quasi-steady-state”, and the velocity decreased gradually due to the decreasing of the cooling flux. It is also observed that the fluid flow velocity is maximum at the center of the flat panel always. Water velocity at the edge of the flat panel is about 75% of the velocity at the center. The reason is because water flow at the center of the flat panel has the shortest flow path, and thus, the smallest flow resistance, for larger velocity. The Reynolds numbers inside the flat panel and the water return tube versus time are plotted in Fig. 8b. The Reynolds number inside the water return tube is larger because the water return tube has a smaller cross-sectional area and higher fluid flow velocity. However, both Reynolds numbers are still within laminar region, which suggest that fluid flow inside the thermosiphon is laminar at all times, and this validates the assumption made for the simulation.
8.5
16 14
8.0
12 7.5
10
7.0
8 6
6.5
4 6.0 5.5
2 6
8
10
12
14
0
Cold Collection Tank Temperature Drop (oC)
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Flat Panel Tilt Angle (o) Fig. 9. Working temperature difference and the cold collection tank temperature drop of the thermosiphon after 2-h operation as a function of flat panel tilt angle. A larger tilt angle gives a larger cold collection tank temperature drop, but a smaller working temperature difference.
temperature difference decreases when the tilt angle increases. Note that the thermosiphon working temperature difference is the final-state temperature difference at the end of 2-h testing period. The reason is because a larger tilt angle gives larger height difference between flat panel inlet and outlet, which requires smaller working temperature difference. Also, larger tilt angle gives larger cold collection tank temperature drop. The average cooling fluxes for the 6°, 8°, 10°, 12° and14° tile angles were 122 W/m2, 129 W/m2, 132 W/m2, 137 W/m2, and 146 W/m2, respectively. A larger tile angle results in a larger average cooling flux. This is due to larger water mass flow rate inside the thermosiphon caused by larger buoyancy at the larger tilt angle, which is also confirmed by numerical results. By the end of 2-h cooling periods, the mass flow rates at the center of flat panel are 5.4 × 10−4 m/s, 5.8 × 10−4 m/s, 6.3 × 10-4 m/s, 7.0 × 10-4 m/s, 7.8 × 10-4 m/s for tilt angles 6°, 8°, 10°, 12°, 14°, respectively. Compared to the 6° tilt angle, the cold collection tank temperature drop increased by 23.7% for the 14° tilt angle.
4.2. Parametric study of single-phase thermosiphon 4.2.1. Effect of flat panel tilt angle on thermosiphon performance A larger tilt angle of the flat plat results in a larger height difference between the inlet and the outlet of the flat panel (see Fig. 1b), giving a greater buoyancy force. For an outdoor radiative cooling application, the final tilt angle will be determined along with other parameters that affect thermosiphon performance, such as radiation shape factor from the thermosiphon to the sky and daytime solar absorption. On one hand, a smaller tilt angle will give a larger radiative shape factor to the sky. On the other hand, to minimize solar absorption at the daytime, the thermosiphon will be tilted to facing the north in northern hemisphere. A larger tilt angle will give a smaller solar irradiance and thus is preferred. The final application of the thermosiphon will dictate the optimal operating condition. The thermosiphon was tested under different tilt angles from 6° to 14°, with 2° intervals to quantify the effect of tilt angle on thermosiphon performance. A tilt angle between 6° and 14° offers sufficient height difference for thermosiphon fluid flow, along with enough radiation shape factor from the thermosiphon to the sky. All tests used 3.1 mm air gap separation with an initial thermosiphon temperature at ∼29 °C for a period of 2-h. Fig. 9 shows that the thermosiphon working
4.2.2. Effect of initial water temperature on thermosiphon performance Keep the thermosiphon working at a relatively high temperature gives a higher radiative cooling flux. To quantify the effect of initial water temperature on thermosiphon performance, the thermosiphon
Fig. 8. (a) Fluid flow velocity inside flat panel versus time and location, (b) Reynolds numbers for fluid flow in the flat panel and the water return tube. Both Reynolds numbers are within the laminar region.
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4.2.3. Effect of cooling flux on thermosiphon performance The operation of the thermosiphon is strongly affected by weather conditions. In cloudy weather conditions, radiative cooling can be blocked by clouds and cooling flux is greatly reduced. Therefore, it is desirable to investigate the thermosiphon performance under different cooling fluxes. Five different cooling fluxes were tested, corresponding to 5 difference air gap separations (i.e. 1.9 mm, 3.1 mm, 4.1 mm, 6.2 mm and 8.4 mm) between the chilled aluminum plate and the flat panel. These 5 air gap separations were selected because they can provide an average cooling flux comparable to different levels of radiative cooling flux. The average cooling fluxes for these 5 air gap separations were obtained at 144 W/m2, 132 W/m2, 111 W/m2, 96 W/ m2, and 74 W/m2 respectively. Initial temperatures inside the thermosiphon were all set to ∼29 °C and the flat panel angles were all 10°. All tests were 2-h long. With increasing air gap separation, both the temperature drop of the cold collection tank and the working temperature difference of the final-state thermosiphon decreases due to the reduced average cooling flux (Fig. 11). The reason is that lower cooling flux results in lower fluid flow velocity inside thermosiphon, which in turn reduces the flow resistance. A lower flow resistance requires a smaller working temperature difference to drive the flow. A lower cooling flux also results in a smaller temperature drop in the cold collection tank due to energy balance. From the 8.4 mm air gap test results, the final-state cooling flux was 65 W/m2 and the thermosiphon is still working, which suggests that the thermosiphon can work on partially cloudy days.
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Fig. 10. Working temperature difference and the cold collection tank temperature drop of thermosiphon at the final state as a function of the initial water temperature. Both cold collection tank temperature drop and thermosiphon working temperature difference increase with initial temperature.
5. Conclusions A single-phase thermosiphon and its testing apparatus has been developed for cold collection and storage of radiative cooling. This is the very first demonstration for cooling down liquid inside the singlephase thermosiphon to a temperature lower than ambient induced by a very small cooling flux, ∼100 W/m2. Both experimental and numerical methods have been used to characterize the thermal performance and flow characteristics of the thermosiphon. The main conclusions that can be drawn from this study are the following:
• The
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Fig. 11. Working temperature difference and cold collection tank temperature drop of thermosiphon at the final state as a function of the air gap separation (i.e. cooling flux). Both cold collection tank temperature drop and thermosiphon working temperature difference decreased with increasing air gap separation (i.e. decreasing cooling flux).
was tested under 5 different initial temperatures, 24.4 °C, 26.6 °C, 27.6 °C, 29.1 °C, and 33.7 °C, respectively. These initial temperatures are selected because for either building cooling application or supplemental cooling application for power plant condensers, the water that needs to be cooled has a temperature within this range [45,47]. Air gap separations are all set at 3.1 mm, and the tilt angles of the flat panel are all set at 10°. All tests are 2-h long. A larger initial temperature gives a larger cooling flux through the air gap. The average cooling fluxes for these 5 different initial temperatures were 108 W/m2, 120 W/m2, 123 W/m2, 132 W/m2, and 158 W/m2, respectively. Both the working temperature difference and cold collection tank temperature drop of the thermosiphon increases with the initial temperature (Fig. 10). By increasing the initial temperature from 24.4 °C to 33.7 °C, the old collection tank temperature drop increased from 8.5 °C to 14.5 °C. For outdoor applications, the radiative cooling system operating time needs to be optimized to achieve both a targeted cold water temperature and a relatively high radiative cooling flux.
• • •
thermosiphon testing apparatus (i.e. the chilled aluminum plate) can provide transient cooling flux between 37 W/m2 and 150 W/m2, which is at the same level as radiative cooling flux. With an average cooling flux of 105 W/m2, 18 liters of water in the thermosiphon was cooled to an average temperature of 12.5 °C from an initial temperature of 22.2 °C in 2 h, with a cold collection efficiency of 96.8%. The testing room temperature is ∼21 °C. This has demonstrated the feasibility of a single-phase thermosiphon for cold collection and storage of radiative cooling. This also demonstrated that the thermosiphon can cool down a heat transfer fluid to a temperature lower than ambient by using a cooling flux comparable to radiative cooling. A larger flat panel tilt angle, higher initial water temperature, and larger cooling flux are desirable for achieving larger cold collection tank temperature drop while the final application of the thermosiphon will dictate the optimal operating conditions. At a lower cooling flux testing condition (65 W/m2), which simulates a partial cloudy day, the thermosiphon can still work, but with a smaller cold collection and storage capability. A radiative cooling system that has cooling capacity in kW-scale or even in MW-scale can be achieved by using multiple thermosiphon modules connected.
Acknowledgements The authors acknowledge the financial support of this work from the US Department of Energy’s ARPA-E (Contract No. DE-AR0000580). 1268
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