An additively manufactured novel polymer composite heat exchanger for dry cooling applications

International Journal of Heat and Mass Transfer 147 (2020) 118889

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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt

An additively manufactured novel polymer composite heat exchanger for dry cooling applications M.A. Arie, D.M. Hymas, F. Singer, A.H. Shooshtari ⇑, M. Ohadi Advanced Heat Exchangers and Process Intensification (AHX-PI) Laboratory, Department of Mechanical Engineering, University of Maryland, College Park, MD 20742, USA

a r t i c l e

i n f o

Article history: Received 14 December 2018 Received in revised form 22 July 2019 Accepted 12 October 2019

Keywords: Next-generation heat exchangers Polymer composite heat exchangers Additive manufacturing Fused Filament Fabrication (FFF) Dry cooling Metal fiber composites

a b s t r a c t The work presented in this paper focuses on the design and thermal characterization of a novel polymer composite heat exchanger (HX) produced by an innovative additive manufacturing process. The heat exchanger represents a gas to liquid configuration in which the gas side removes heat from the liquid side in a cross-flow arrangement. The novel HX utilizes a cross media approach in which, unlike the conventional HXs, the hot and cold sides are directly connected to each other through high conductivity metal fiber fins on the gas side protruding through the walls of the liquid side, thus eliminating the wall resistance separating the hot and cold sides. The HX demonstrates superior thermal performance at reduced pressure drops while also benefiting from the lighter weight and the lower cost that the polymer structure introduces. A 350-W water-to-air heat exchanger was fabricated using a fused filament fabrication (FFF) technique with a novel/patent pending printer head which was developed to produce the metal fiber composite structure of the heat exchanger. The results of the heat exchanger characterization tests show that it yields up to 220% and 125% improvement in heat flow rate over mass (Q =m) and heat flow rate over volume (Q =V), respectively, when compared to comparable state-of-the-art plate fins HX configurations. This study in particular demonstrates the impact of additive manufacturing in realizing potentially transformative heat exchanger technologies that may otherwise be very difficult to achieve with conventional fabrication methods. Ó 2019 Elsevier Ltd. All rights reserved.

1. Introduction Compared to wet cooling towers in which water is utilized to enhance rate of cooling through evaporation or other mechanisms, dry cooling uses ambient air for cooling with net zero use of water, thus avoiding water losses associated with evaporation and draining that is typical in many wet cooling towers. With increasing demand for water in many parts of the world, especially in areas where freshwater supplies are sparse, the need for more efficient and cost-effective dry cooling technologies is more than ever a priority. This includes the need for innovative fin designs; lighter weight/lower cost heat exchangers with improved thermal performance and without additional pumping power penalties. However, realizing such improvements without resorting to innovative materials and manufacturing techniques may be difficult. Various efforts have been reported on enhancing the performance of existing conventional heat transfer surfaces by adding measures such as shaped reentrant cavities and internal ribs [1], ⇑ Corresponding author at: Department of Mechanical Engineering, 2181 Glenn L. Martin Hall, Building 088, University of Maryland, College Park, MD 20742, USA. E-mail address: [email protected] (A.H. Shooshtari). https://doi.org/10.1016/j.ijheatmasstransfer.2019.118889 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.

vortex generators [2], or piezoelectric translation agitators (PTA) [3]. Other works involve developing new heat transfer surfaces by introducing porous media [4], oblique fins [5], or by using impinging jets to enhance the heat transfer rate [6,7]. However, fabrication of these surfaces is generally a complicated and laborintensive process. Moreover, most of these technologies are accompanied by a significant increase in pressure drop, rendering them less attractive for most applications. Several previous studies have focused on the development of polymer heat exchangers such as polymer plate heat exchangers [8–12] and polymeric hollow fiber heat exchangers [13,14]. Compared to metals, polymers have lower thermal conductivity; therefore, the size of a conventional polymer heat exchanger is usually much larger than an equivalent metallic heat exchanger. However, polymers offer better corrosion resistance, chemical resistance, lower density (lower mass), and lower cost than metals. Harris et al. fabricated a microchannel heat exchanger from polymethylmethacrylate (PMMA) and compared its performance to an identical heat exchanger fabricated from nickel [10]. The results for the same pressure drop showed that while the nickel heat exchanger had higher volumetric heat transfer densityðQ =ðV DTÞÞ, the PMMA

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Nomenclature A C COP cp Dwire Dair f h k keff kc ke L Lc m _ m N NTU Dp Q Q =m Q =ðmDTÞ Q =V Q =ðV DTÞ R Re S T DT t U

v

W

area [m2] heat capacity rate [W/K] coefficient of performance [-] specific heat [J/kgK] wire diameter [m] air-side hydraulic diameter [m] friction factor [-] heat transfer coefficient [W/m2K] thermal conductivity [W/mK] effective thermal conductivity [W/mK] coefficient of contraction [-] coefficient of expansion [-] length [m] fin height [m] mass [kg] mass flow rate [kg/s] total number of wire [-] number of transfer units [-] pressure drop [Pa] heat flow rate [W] heat flow rate over mass [W/kg] gravimetric heat transfer density [W/kgK] heat flow rate over volume [W/m3] volumetric heat transfer density [W/m3K] thermal resistance [K/W] Reynolds number [-] wire spacing [m] temperature [°C] temperature difference [°C] thickness [m] overall heat transfer coefficient [W/m2K] velocity [m/s] width [m]

q g l r

density [kg/m3] fin efficiency [-] dynamic viscosity [kg/ms] area ratio of minimum free flow area over frontal area [-]

Subscripts air air-side int interface area chn channel conv convection cond conduction fin fins in inlet H heat transfer area no-flow No-flow direction out outlet stack stack water water-side wall polymer wall wire wire Abbreviations AM additive manufacturing ABS acrylonitrile butadiene styrene CMHX cross-media polymer composite heat exchanger DMLS direct metal laser sintering FFF fused filament fabrication HX heat exchanger LPFS louvered plate-fin surface PMMA polymethylmethacrylate PPFS plain plate-fin surface WPFS wavy plate-fin surface

Greek Letters e effectiveness [-]

heat exchanger was superior in terms of gravimetric heat transfer density (ðQ =ðmDTÞ). To address the issue of the polymer’s low thermal conductivity, two approaches are generally pursued in heat exchangers. One approach involves reducing the wall thickness (~100 mm) to compensate for the low thermal conductivity; this approach requires a new heat exchanger design [15,16]. The other approach involves the development of filled polymer composite materials with thermal conductivities much higher than regular polymers [17–23]. This is achieved by filling the polymer matrix with thermally conductive material such as metallic, carbon-based, and ceramic fillers [24]. One such composite reported is a polymer matrix with graphene nanoplatelet fillers. It was reported that polyphenylene sulfide with 29.3% volume graphene nanoplatelets can achieve a thermal conductivity of 4.4 W/mK [23]. Another composite was developed from a polymer with carbon nanotube fillers. Depending upon the properties of the carbon nanotube fillers, their orientation, and the overall volume fraction of the fillers, a composite material thermal conductivity up to 41 W/mK, along the carbon nanotube alignment direction, has been reported [18]. However, the costs of such composite materials are much higher than regular polymers, and most often the composites exhibit anisotropic thermal characteristics. The high loadings of filler material also increase the weight of the material, lead to rough surfaces, and generate problems in the production process. In addition, the reported thermal performance of heat exchangers fabricated using filled polymers has shown infe-

rior performance when compared to metallic heat exchangers since the thermal conductivity of such structures is significantly lower that of metals. Robinson et al. studied the use of carbon fiberreinforced polyamide in fabricating a polymer finned plate heat exchanger [11]. It was reported that although the filling material improved the heat transfer performance by 70%, this performance was still 28% lower than a titanium heat exchanger. In another work, Glade et al. reported the development of a polymer composite tube out of polypropylene with 50% volume graphene to be used for heat exchanger application [25]. Despite significant improvement in thermal conductivity, its overall heat transfer coefficient was still 14–20% lower than copper-nickel and aluminum brass tube heat exchangers. In addition, the graphene filler increased the density of the polypropylene by 67%. In the present work, we present a new approach involving a novel cross-media polymer composite heat exchanger (CMHX). This CMHX yields an effective thermal conductivity of 130 W/mK without significant cost increase. The CMHX is fabricated using an additive manufacturing (AM) process first demonstrated by our team as reported in Hymas et al. [26]. As most commonly known, in an AM process objects are built layer-by-layer from a pre-programmed digital model. Compared to conventional fabrication techniques, AM offers multiple advantages, such as freedom of design, lead-time reduction, and waste material reduction. Freedom of design is especially important in heat exchanger design, and AM allows complex geometries to be fabricated which are challenging to fabricate using conventional

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methods. In recent years, several works on heat exchanger fabrication using AM have been reported, such as manifold-microchannel heat exchangers by Arie et al. [27–30] and Zhang et al. [31], a micro heat sink by Robinson et al. [32], a webbed tube heat exchanger by Cevellos [33], a thin wall polymer heat exchanger by Arie et al. [15], and microchannel heat exchangers by Tsopanos et al. [34]. A more detailed review of additively manufactured heat exchangers can be found in [12,30,35,36]. Compared to other heat exchangers which have been successfully fabricated using AM, CMHX is unique, as it is fabricated out of two materials (metal fiber and polymer) at once. To the best of the authors’ knowledge, this is the first time a metal fiber composite heat exchanger has been fabricated using AM. This paper reports results for a 350 W CMHX prototype which was fabricated, experimentally tested, and its performance was compared with state-of-the-art dry cooling heat exchangers to evaluate the performance advantages of the CMHX.

2. Heat exchanger geometry and fabrication The fabricated and tested CMHX is an air-water heat exchanger consisting of continuous high conductivity metal wires which pass directly through the polymer tube walls as shown in Fig. 1. The air and the water flow in cross-flow configuration. The polymer tubes are used to direct the flow and form a physical barrier between the air side and the water side. The continuous metallic wires thermally couple the air-side and the water-side fluid flows, as shown in Fig. 1 (a). Since heat is predominantly transferred from the water side to the air side via the metallic wires, the CMHX is not limited by the low conductivity of the polymer tube wall. In addition, the wires also act as a flow disruptor, which increases the heat transfer coefficient in both sides. The wires are in an inline configuration with a spacing of Sair in the air-flow direction and Swater in the water-flow direction, as shown in Fig. 1(b). In the current work, the polymer tubes were fabricated from ABS (acrylonitrile butadiene styrene), and aluminum wires were used as the metallic fibers. A summary of the CMHX geometrical variables is provided in Table 1. The CMHX was fabricated using a modified Fused Filament Fabrication process (FFF), an AM technique [37]. The printer consists of two print heads: one head deposits the polymer to form the CMHX channel walls, while the other head embeds the metallic wires into the polymer structure as it is being built. The fabrication process is shown in Fig. 2 and consists of the following steps. (1) The channel walls are fabricated using the polymer head. (2) After a certain height (Swater ) is reached, the metal head will lay down the metallic wires into the top of the polymer structure. (3) The polymer head will continue to fabricate the channel walls. (4) The process is repeated until the entire heat exchanger is fully formed. A more detailed description of the fabrication process can be found in [26]. Fig. 3 shows the fabricated 350 W CMHX with the water-side headers attached. The water-side headers were fabricated separately using an off-the-shelf FFF printer and then connected to the heat exchanger using marine epoxy. The heat exchanger core (excluding the headers) has an overall volume of 390 cm3 and overall mass of 0.127 kg. Since the CMHX was fabricated out of ABS with a glass transition temperature of 105 °C, the heat exchanger is capable to operate under water-side temperatures up to 70 °C [38]. Moreover, the pressure test has shown that the fabricated CMHX is capable of operating with no leakage up to a system pressure of 30 Psig.

3. Experimental test setup The schematic diagram of the CMHX experimental test setup is shown in Fig. 4. The air side consisted of an open loop with a blower

Fig. 1. CMHX (a) Full-view; (b) side cross-sectional view.

to drive the flow, a flow meter to measure the air flow rate, an offthe-shelf heat exchanger to control the ingoing air temperature,

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M.A. Arie et al. / International Journal of Heat and Mass Transfer 147 (2020) 118889 Table 1 Geometrical variables of the CMHX. Air Side Sair W chn;air N air

1.29 mm 12 mm 21

Water Side Swater W chn;water N water

1.93 mm 5 mm 68

Metal Wire Dwire

0.5 mm

Polymer Wall twall

1 mm

Overall Size Lair Lwater Lnoflow

29 mm 132 mm 81 mm

sured pressure drop and temperature on both the air and the water sides. On the water side, both the temperature and pressure were measured in the upstream and downstream of the manifold. The upstream and downstream measurements were chosen to allow for a better flow mixing. Two thermocouples were used to measure the inlet water temperature and a differential thermopile, with accuracy of 0.04 °C, was used for water-side temperature drop measurement. To evaluate the pressure losses in the manifold, CFD simulations of the manifold were performed. The results show that the pressure drop in the manifold is negligible compared to the pressure drop in the heat exchanger core (<1%) for the entire test range. On the air-side, the pressure drop was measured directly next to the inlet/outlet faces of the heat exchanger. For the temperature measurement, a honeycomb structure was added in the inlet and exit to promote air mixing. The temperature was measured using T-type thermocouple in four evenly spaced locations in both the inlet and exit manifolds.

4. Experimental procedure and data reduction

and the experimental test section (including the CMHX) at the end of the loop. To control the air flow rate, a variable speed controller was used. The water side consisted of a closed loop with a chiller to control the water temperature and drive the flow, a flow meter to measure the water flow rate, and the experimental test section. Differential pressure transducers and thermocouples (T-types) mea-

Polymer printing head

Experimental testing was performed to evaluate the performance of the CMHX. Two sets of experiments were conducted. First, air-side flow rate was varied from 12 g/s to 33 g/s while the water-side flow rate was kept constant at 33 g/s. Second, the water-side flow rate was varied from 6 g/s to 33 g/s while the air-side flow rate was kept constant at 30 g/s. Table 2 summarizes the experimental test conditions. For each run, the air-side

Metal fibers

Polymer tubes

Metal fibers printer

Print the first layer of polymer tubes

Hot water in

Inlay the metal fibers on top of the printed tube

Metal fibers Metal fibers printing head

Cold air in Metal fibers were cut Repeat the process until the desired number of layers is reached

Cut the metal fibers at the point of extrusion

Fig. 2. The CMHX fabrication process summary.

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M.A. Arie et al. / International Journal of Heat and Mass Transfer 147 (2020) 118889

Fig. 3. Fabricated CMHX (With water-side headers attached).

Fig. 4. Schematic diagram of the experimental setup.

inlet and exit temperatures (T in;air and T out;air ), water-side inlet temperature and temperature variation (T in;water and DT water ), air_ air and m _ water ), and air-side side and water-side mass flow rates (m and water-side pressure drops (Dpair and Dpwater ), were recorded. To evaluate the performance of the CMHX, the heat flow rate of the heat exchanger (Q ), heat exchanger effectiveness (e), overall heat transfer coefficient (U), and air-side heat transfer coefficient (hair ) were calculated. The heat exchanger heat flow rate was evaluated by taking the average of the air-side and water-side heat flow rates (Q air and Q water ) as calculated by Eqs. (1)–(3). For the entire experimental testing, the energy balance between the airand the water-sides was determined to be within ±16%. The source of the error in the energy balance was mainly due to instrument uncertainty at low heat flow rate conditions.

_ air cp;air ðT out;air  T in;air Þ Q air ¼ m

ð1Þ

_ water cp;water ðT in;water  T out;water Þ Q water ¼ m

ð2Þ

Table 2 The CMHX experimental conditions. Temperature Boundary Conditions T in;air T in;water

20 °C 40 °C

Varying Air-side _ water m _ air m

33 g/s 12–33 g/s

Varying Water-side _ water m _ air m

Q¼

6–33 g/s 30 g/s

ðQ air þ Q water Þ 2

ð3Þ

Then, the heat exchanger effectiveness can be evaluated as:

e¼

Q
 min ðC water ; C air Þ  T in;water  T in;air

ð4Þ

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M.A. Arie et al. / International Journal of Heat and Mass Transfer 147 (2020) 118889

_ water cp;water , C air ¼ m _ air cp;air , and cp is the specific where C water ¼ m heat. To calculate the overall heat transfer coefficient, the NTU-e method for unmixed cross flow on both sides was used. First, NTU was calculated as a function of the effectiveness and the specific heat ratio (C r ) as shown in Eq. (5).

h

n

h

i

e ¼ 1  exp ð1=C r ÞNTU0:22 exp C r ðNTUÞ0:78  1

oi

ð5Þ

min ðC water ;C air Þ where C r ¼ maxðC . water ;C air Þ

Thereafter, the overall heat transfer coefficient was calculated as:

U¼

NTU  minðC water ; C air Þ Aint

ð6Þ

where Aint is the nominal interface area (Aint ¼ 2Lwater Lair ðNchn;water  1Þ) and N chn;water is total number of water-side channels (N chn;water = 5 for this case). For the overall heat transfer coefficient, it was assumed that the heat transfer into the polymer wall was negligible. This is a valid assumption, as the thermal conductivity of the aluminum is 800 times larger than that for polymers. The overall heat transfer coefficient can be related to the air-side heat transfer coefficient by thermal resistance network analysis on the metallic wire as shown in Eq. (7).

U¼


1  Rconv ;air þ Rconv ;water þ Rcond wire Aint

ð7Þ

The water-side thermal resistance (Rconv ;water;wire ) was evaluated using a Wilson plot method where 1=UA (overall heat transfer coefficient multiplied by the area) was plotted versus vNwater. Here v water is water velocity at the inlet and N=0.8 as described in [39]. On the
 other hand, the conduction thermal resistance Rcond;wire was evaluated as shown in Eq. (8) assuming heat transfers only through wires and not polymer walls.

Rcond;wire ¼

twall kwire Awire;int

ð8Þ

where twall is thickness of the wall separating the air- and waterside, kwire is the thermal conductivity of the wire, and Awire;int is the interface cross sectional area of the wire  2 Dwire N air N water 2ðNchn;water  1Þ). The 2ðN chn;water  1Þ (Awire;int ¼ p 2 term represents the total number of interfacial walls between air and water passages in the heat exchanger which for this case is a total of 8. Thereafter, the air-side thermal resistance (Rconv ;air;wire ) was evaluated as a function of U, Rconv ;water;wire , and Rcond;wire using Eq. (7). Lastly, the air-side heat transfer coefficient was evaluated using Eqs. (9)–(11).

Rconv ;air;wire ¼

gfin ¼

1 hair gfin AH;air

tanh ðSLc Þ ðSLc Þ

sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 4hair S¼ kwire Dwire

ð9Þ

W chn;air 2

Uncertainty propagation analysis was performed to measure the inaccuracy due to the measurement devices used: the flow meters, pressure transducers, and thermocouples. A list of the measurement equipment used in the testing and their accuracy is shown in Table 3. The uncertainty in pressure drop can be directly calculated based on the equipment accuracy as listed in Table 3. However, it is more complex to evaluate the uncertainties in the heat transfer parameters, such as the heat flow rate, the overall heat transfer coefficient, and the air-side heat transfer coefficient. Therefore, a method explained in NIST Technical Note 1297 [40] was used. Based on the method, the uncertainty U Z of a calculated quantity Z, which is a function of X 1 ,X 2 , . . . ,X N with uncertainty of U X 1 ; U X 2 ;    ; U X N , can be calculated as:

vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u uX  @Z 2 2 UZ ¼ t U Xi @X i i

ð12Þ

The uncertainty in Q and U were evaluated using Eq. (12) as a function of the uncertainty in the water-side and air-side flow rates, and the inlet and exit temperatures. For the air-side heat transfer coefficient uncertainty evaluation, the uncertainty of the slope of 1=UA versus vNwater of the Wilson plot needs to be evaluated first using Eq. (12). Thereafter, the uncertainty in the heat transfer coefficient can be evaluated as a function of the uncertainty of the Wilson plot slope, the flow rate and the temperature. The maximum errors in Q , U, and hair were determined to be ±4%, ±6%, and ± 16%, respectively. 6. Results and discussion 6.1. Heat transfer performance The heat transfer performance results are shown in Fig. 5 as a _ air ) with a constant water flow function of the air-side flow rate (m rate of 33 g/s. Due to the relatively high accuracy on the air-side flow meter (±2%), the error bar in the air-side flow rate is not visible. The results are also shown as a function of air-side Reynolds number (Reair ), where Reynolds number is defined as:

Re ¼

v in Dwire q

ð13Þ

l

where v in is the velocity at the entrance of the heat exchanger, and Dwire is the diameter of the metallic wires. The heat flow rate (Q ), the overall heat transfer coefficient (U), and the air-side heat transfer coefficient (hair ) of the CMHX are shown in Fig. 5. A heat flow rate up to 330 W, an overall heat transfer coefficient up to 150 W/m2K, and an air-side heat transfer coefficient up to 325 W/m2K were achieved. The graphs show that the heat flow rate, the overall heat transfer coefficient, and the air-side heat transfer coefficient follow an increasing trend as the flow rate increases. This trend was expected, as the metallic wires continuously disrupted the flow, which kept the flow in the developing

ð10Þ Table 3 List of measured parameters and their accuracies.

ð11Þ

where gfin is the fin efficiency, AH;air is the total air-side heat transfer area (AH;air ¼ pDwire W chn;air N air N water ðN chn;water  1Þ) where W chn;air represents the width of air passage, Lc is fin height (Lc ¼

5. Uncertainty analysis

= 6 mm for this case), and Dwire is the wire diameter.

Measured Parameters

Accuracy

Air-side flow rate Water-side flow rate Temperature Temperature difference Air-side pressure drop Water-side pressure drop

2%of measured value 1%of measured value 0:5o C 0:04o C 0.14% FS of 5 inH2O 0.25% FS of 5.5 kPa

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Fig. 6. Overall heat transfer coefficient results for the 350 W CMHX (the air flow rate was maintained constant at 30 g/s).

water-side thermal resistance is still half of air-side thermal resistance. On the other hand, the base thermal conduction resistance contributes only to about 5% of the total resistance. 6.2. Pressure drop performance The air-side pressure drop (Dpair ) and water-side pressure drop (Dpwater ) as a function of flow rate are plotted in Fig. 7. The resulting trend was obtained as expected, where the pressure drop increases as the flow rate increases. An air-side pressure drop up to 71 Pa was recorded, while a water-side pressure drop of up to 800 Pa was recorded. 6.3. Comparison to conventional heat exchangers

Fig. 5. Heat transfer performance results for the 350 W CMHX (the water flow rate was maintained constant at 33 g/s): (a) Heat flow rate and overall heat transfer coefficient; (b) air-side heat transfer coefficient.

region and in turn increased both the heat flow rate and the overall heat transfer coefficient. The overall heat transfer coefficient as a function of the water_ water ) and Reynold number (Rewater ) for a constant side flow rate (m air flow rate of 30 g/s is shown in Fig. 6. Water-side Reynolds numbers were evaluated using Eq. (13). Due to the high accuracy of the water-side flow meter (±1%), the error bar in the water-side flow rate is not visible. Compared to the case of varying the air-side flow rate, varying the water-side flow rate does not significantly affect the overall heat transfer coefficient. Increasing the water-side flow rate from 6 g/s to 33 g/s increases the overall heat transfer coefficient from 120 W/m2K to only 145 W/m2K (20% increase). On the other hand, increasing the air-side flow rate from 12 g/s to 36 g/s increases the overall heat transfer coefficient from 90 W/m2K to 153 W/m2K (70% increase). This shows that the air-side thermal resistance is more dominant than the water-side thermal resistance, rendering the air-side as the limiting factor for this heat exchanger. At water flow rate of 33 g/s, the water-side thermal resistance contributes to 10% of the total resistance while airside thermal resistance contributes to 85% of the total resistance. The water-side thermal resistance increases as the water flow rate reduces. However, even at a very low water flow rate of 6 g/s, the

To rate the performance of the CMHX, its performance was compared to state-of-the-art heat exchangers. The current state-of-the-art air-water heat exchangers consist of rectangular channels in the water side and fin surfaces (such as plain plate fins, louvered fins, or wavy fins) on the air side, as explained by Arie et al. [41]. For this comparison, twelve different types of commercially available plate-fin heat exchanger surfaces, as listed by Kays

Fig. 7. The plots of the air-side pressure drop and water-side pressure drop, of the 350 W CMHX.

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and London [42], were selected. The major geometrical parameters of the heat exchanger surfaces are listed in Table 4, and other details can be found in [42]. The conventional heat exchangers were sized with the objective of minimizing mass. The sizing was performed by varying the airside geometry as listed in Table 4 while on the water side, the channels size was allowed to vary. The overall size of the heat exchanger was unconstrained and allowed to vary. The performance constraints involved matching the following parameters of the conventional heat exchangers to those of the CMHX: heat flow rate (Q ) (or effectiveness (e)), air-side and water-side mass _ water ), inlet air temperature (T in;water ), inlet _ air and m flow rates (m water temperature (T in;air ), and air-side and the water-side pressure drops (Dpair and Dpwater ). The sizing process was performed numerically. The water-side performance (rectangular channels) was determined using available correlations for rectangular channels while the air-side performance (fin surfaces) was evaluated based on the thermal performance data for common fins as provided by Kays and London [42]. For a direct comparison with the experimental results, the air-side pressure drops of the conventional surfaces also include the entrance and exit pressure losses due to contraction, expansion, and flow acceleration/deceleration, which can be evaluated using Eq. (14):

COP ¼

Q _ air =qair þ Dpwater m _ water =qwater Þ ðDpair m

ð15Þ

Fig. 8 shows that the CMHX yields higher performance over all the conventional heat exchangers for the entire test range. Compared to the wavy fins, as shown in Fig. 8(a), up to 165% increase in the Q =m is possible for the same coefficient of performance. Compared to the louvered fins and the plain plate fins, up to 220% and 175% increase, respectively, in the Q =m is possible for the same coefficients of performance, as shown in Fig. 8(b) and (c). The significantly lower mass of the CMHX is due to the use of the polymer fluid channels, which have significantly lower density than all-metal heat exchangers. The significant mass reduction can also be attributed to the pin-fin geometry, which has a higher surface area to volume ratio than a similar flat plate. Fig. 9 compares the heat flow rate over volume (Q =V) between the CMHX and conventional heat exchangers. Since the major advantage of the CMHX is the use of low density, polymer fluid channels, the volume advantage of the CMHX is not as high as the mass advantage. However, a significant gain in Q =V is still observed: compared to the wavy fins, the louvered fins, and the plain plate fins, up to 55%, 125% and 120% increase, respectively, in Q =V is possible for the same coefficient of performance (Fig. 9). These results show that the CMHX could reduce both volume and mass compared to all-metal heat exchangers. 6.4. Effective thermal conductivity calculation

ð14Þ

where v in;air is the inlet air velocity, kc is the coefficient of contraction, ke is the coefficient of expansion, r is the area ratio (minimum free flow area divided by the frontal area), and f is the friction factor [43]. Eq. (14) is derived by assuming constant air density from the heat exchanger entrance to its exit. This assumption is valid due to the low temperature difference between the hot side and the cold side of the HX. Details of the heat exchangers sizing method can be found in [41]. The performances of the CMHX and the conventional heat exchangers are compared in Fig. 8, where the heat flow rate over mass (Q =m) is plotted as function of the coefficient of performance (COP). The COP was evaluated as the ratio of heat flow rate over total pumping power in both the air and the water sides as:

To further compare the CMHX with the state-of-the-art polymer composite heat exchangers, the effective thermal conductivity (keff ) of the CMHX was evaluated. The effective thermal conductivity represents the thermal conductivity required by a similar design heat exchanger to achieve the same thermal performance as the current cross media concept if it is fabricated out of a single material only. The effective thermal conductivity can be evaluated by analyzing the thermal resistance network of the heat exchanger. To this end, heat transfer can be considered through both the wires and the polymer wall as derived below:

1

UAint ¼ Rconv ;air þ Rconv ;water þ Rcond wire 1

þ Rconv ;air þ Rconv ;water þ Rcond wall

ð16Þ

Table 4 Conventional Heat Exchanger Surfaces [42] Plain Plate-Fin Surface (PPFS) Symbol

Fin Pitch (fin/in)

Plate Spacing (in)

PPFS PPFS PPFS PPFS PPFS

11.1 15.08 10.27 14.77 19.86

0.25 0.418 0.544 0.33 0.25

1 2 3 4 5

Louvered Plate-Fin Surface (LPFS) Symbol

Fin Pitch (fin/in)

Plate Spacing (in)

Louver gap (in)

Louver Spacing (in)

LPFS LPFS LPFS LPFS LPFS

6.06 6.06 11.1 6.06 11.1

0.25 0.25 0.25 0.25 0.25

0.0275 0.065 0.02 0.0275 0.0275

3/8 3/8 3/4 1/2 1/2

1 2 3 4 5

Wavy Plate-Fin Surface (WPFS) Symbol WFPFS 1 WFPFS 2

Fin Pitch (fin/in) 11.44 11.5

Plate Spacing (in) 0.413 0.375

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Fig. 8. Mass-based performance comparison between CMHX and conventional HXs: (a) wavy fin HX; (b) louvered fin HX; (c) plain plate fin HX.

Fig. 9. Volume-based performance comparison between CMHX and conventional HXs: (a) wavy fin HX; (b) louvered fin HX; (c) plain plate fin HX.

The water-side thermal resistance on the wire (Rconv ;water;wire ) was determined using the Wilson plot method as described in section 4. The wire’s conduction thermal resistance (Rcond;wire ) and airside thermal resistance (Rconv ;air;wire ) were evaluated using previously defined Eqs. (8)–(9). The major difference is that the effective thermal conductivity was used instead of the wire thermal conductivity. The wall-side thermal resistances were evaluated using Eqs. (17)–(19) below:

Rconv ;air;wall

1 ¼ hair;wall Awall;int

ð17Þ

Rcond;wall ¼

twall keff Awall;int

Rconv ;water;wall ¼

1 hwater;wall Awall;int

ð18Þ

ð19Þ


 where the wall interface area Awall;int was evaluated as: Awall;int ¼ ðAint  Awire;int Þ. As there are no fins on the polymer wall, the air-side and water-side heat transfer areas are equal to the wall

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M.A. Arie et al. / International Journal of Heat and Mass Transfer 147 (2020) 118889

interface area. The air-side and water-side heat transfer coefficients on the wall (hair;wall and hwater;wall ) were numerically approximated by using CFD simulations to be about 20% of the air-side and waterside heat transfer coefficients on the wire (hair and hwater ). Using the previously calculated overall heat transfer coefficient (U) as shown in Fig. 5 and Fig. 6, air-side heat transfer coefficient on the wall as shown in Fig. 5, and Wilson plot results for Rconv ;water;wire and hwater;wall , the effective thermal conductivity was determined to be 130 ± 10 W/mK for the CMHX. This thermal conductivity is much higher than that for a polymer heat exchanger (k ~ 0.2 W/mK) or a heat exchanger fabricated out of state-of-theart carbon filament filled polymer, which has reported thermal conductivity of up to 41 W/mK along the carbon nanotube alignment direction [18]. 7. Conclusions In this study, a novel composite polymer heat exchanger (CMHX) was fabricated, experimentally tested, and its performance results compared to respective state-of-the-art heat exchangers. The novel CMHX utilized the innovative cross-media heat exchanger concept, which was described in this paper. The heat exchanger was fabricated using a patent-pending, fused filament fabrication 3D printer head. Experimental results demonstrated an overall heat transfer coefficient and the air-side heat transfer coefficient of 150 W/m2K and 325 W/m2K, respectively, for air-side flow rates of 33 g/s and water-side flow rate of 33 g/ s. Compared to the conventional heat exchangers, the CMHX displayed up to 220% and 125% increase in Q =m and Q =V, respectively, for the same coefficient of performance. The corresponding effective thermal conductivity of 130 W/(mK) for the CMHX is several times higher than that for the state-of-theart polymer/polymer composite heat exchangers. Moreover, as the CMHX was additively manufactured using commercially available ABS filaments and aluminum wires, a lower mass-produced cost is projected when compared to state of the art respective HXs. Lastly, this work demonstrates the impact of additive manufacturing in realizing high performance, low cost air-to-liquid HXs for dry cooling applications and beyond. Declaration of Competing Interest

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The authors declare that there is no conflict of interest. [19]

Acknowledgements Financial support of this work by the U.S. Department of Energy, ARPA-E division, under award number DE- AR0000584 is gratefully acknowledged. The views and opinions of authors expressed herein do not necessarily state or reflect those of the United States Government or any agency thereof. The authors would like to take this opportunity to thank ARPA-E Program Director, Dr. David Tew, and other ARPA-E staff including Dr. Adison Stark, Mr. Geoffrey Short, and Mr. Joel Fetter for their technical insight and advice over the course of this project.

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Appendix A. Supplementary material Supplementary data to this article can be found online at https://doi.org/10.1016/j.ijheatmasstransfer.2019.118889.

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