Experimental study of the optimal design and performance of a mixed-flow dew-point indirect evaporative cooler

Research Paper Experimental study of the optimal design and performance of a mixed-flow dew-point indirect evaporative cooler Alessandra Urso a,b , Eloy Velasco-G´ omez b , Ana Tejero-Gonz´ alez b,* , Manuel Andr´ es-Chicote b , Francesco Nocera a a Department of Civil Engineering and Architecture (DICAR), University of Catania, 95125 Catania, Italy b Departamento de Ingeniería Energ´ etica y Fluidomec´ anica, Grupo de Investigaci´ on de Termotecnia, Instituto de las Tecnologías Avanzadas de la Producci´ on (ITAP), Escuela de Ingenierías Industriales, Universidad de Valladolid, Paseo del Cauce No. 59, 47011 Valladolid, Spain ARTICLE INFO Keywords: Cooling energy poverty Experimental study Dew point indirect evaporative cooler Mixed-flow configuration Polycarbonate plates ABSTRACT Evaporative cooling technologies represent a promising alternative to face the emerging cooling energy poverty, owing to their low production and operative costs. High performance can be achieved through more complex designs, such as dew point indirect evaporative cooling (DIEC) systems. Recent literature explores improvements on these systems, like flow configuration, materials, and water distribution. This study proposes a compact mixed-flow DIEC system made of polycarbonate plates covered with two possible wicking materials. Three water distribution systems are also studied. The prototypes are experimentally characterised by varying the inlet air temperature, humidity, volume flow, and working-to-intake air ratio. The best performing design is evaluated in terms of temperature drop, dew-point effectiveness, wet-bulb effectiveness, and cooling capacity. Results are consistent with those in the literature for equivalent heat transfer areas. The use of a wicking material improves the cooling capacity by up to 1.45. The type of material is less relevant, which enables to select the most economic and accessible option. External nozzles for water distribution offers temperature drops of more than 1 ◦C and better cooling capacities of approximately 100 W than inlet water distributors. 1. Introduction Energy poverty is an emerging topic in the European Union, especially after the Covid-19 pandemic and the energy crisis due to the war in Ukraine [1]. It could be defined as the incapability of a householder to avail adequate domestic energy services, namely space heating, cooling, or cooking [2]. In particular, the problem of cooling energy poverty is still not wellexplored in European scientific literature [2], even if it has been gaining more attention amid the rising temperatures due to global warming and the consequent increase in energy demand for indoor space cooling. As reported in [3], the increase in cooling demand may force the lowest income households into summer energy poverty. This problem may become more evident in southern European countries, such as Greece, Italy, and Spain, which are more vulnerable to high temperatures and heatwaves [3]. Among the existing alternative cooling systems to mechanical vapor compression, evaporative coolers (ECs) are drawing interest in recent years for their higher coefficient of performance (COP) [4] and the possibility of low production costs [5,6]. In addition, ECs have a simple and compact design, i.e., they can be installed in buildings easily [7]. Despite these undisputable advantages, ECs do not outperform in the market [5]. The main drawback of the ECs is that their effectiveness is strongly dependent on the humidity levels in the air supply, which is drawn from the outside. Therefore, they find applicability only in the hot-and-dry climate. By considering the European K¨ oppen–Geiger map, hot and dry summer seasons are characteristic of the Mediterranean coast, and they are expected to be dominant in the rest of southern Europe in future decades [8]. Moreover, recent studies have attempted to integrate dehumidification methods with ECs, which extends the application of ECs to temperate and humid areas [6]. Therefore, the limited climate applicability of evaporative coolers cannot justify their scarce diffusion in the present and future European markets. Despite the scientific advancement of this technology, persistent challenges, e.g., water consumption and complexity of equipment, especially those of water distributors [5], should be solved to commercialize these systems. * Corresponding author. E-mail address: [email protected] (A. Tejero-Gonz´ alez). Contents lists available at ScienceDirect Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng https://doi.org/10.1016/j.applthermaleng.2024.124294 Received 29 April 2024; Received in revised form 6 August 2024; Accepted 29 August 2024 Applied Thermal Engineering 257 (2024) 124294 Available online 7 September 2024 1359-4311/© 2024 The Author(s). Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license ( http://creativecommons.org/licenses/bync-nd/4.0/ ). 1.1. Evaporative cooling technologies and working principles ECs leverage the effect of water evaporation to cool down the ambient air: when air interacts with water, the sensible heat in the air is converted to latent heat by an isenthalpic process. Based on the interaction between the supply air (i.e., the air to be introduced into the indoor space) and water, ECs are classified into direct evaporative coolers (DEC) and indirect evaporative coolers (IEC) [9]. In DEC, the supply air comes in direct contact with water. This system provides a higher temperature drop, but inevitably increases the moisture content in the supply air. In some cases, this renders the system less effective in guaranteeing adequate hygrothermal comfort in occupied spaces [10]. Conversely, in IEC, humidification of the supply (or primary) air is avoided by combining the evaporative cooling process in a working (or secondary) airstream with a heat exchanger [11]. To improve the performance of IEC and overcome the limitations of DEC, a study proposed the dew-point indirect evaporative cooler (DPIEC) [12]. A DPIEC is an IEC in which part of the supply air in the dry channel is diverted to the wet channel as the working fluid. Thus, the working fluid is pre-cooled and to a temperature below its wet-bulb temperature. This temperature eventually reaches the dew-point temperature as the process is repeated cyclically. The Maisotsenko cycle maximizes this potential, and its application is now prevalent in various fields [13]. DPIECs offer a higher temperature drop than conventional ECs, as well as higher cooling capacity, provided the ratio between secondary and intake air is properly dimensioned. This has the advantage of an improved COP, reduced energy consumption, compactness for easy installation in buildings [14], raw material conservation, and reduced miscellaneous costs. 1.2. Improved designs of DPIEC systems Despite the high energy efficiency and low environmental impact of DPIEC systems, several issues related to their design, such as optimal water supply, promotion of auto-wicking, and selection of durable materials with the best thermal transfer properties and water retention [12], must be addressed. Various studies have focused on the effects of flow configuration, material, and water supply system. 1.2.1. Flow configuration The flow configuration refers to the direction of primary air flow relative to the secondary air flow. Typically, DPIEC systems are designed in the counter-flow or cross-flow configuration. In counter-flow, the primary air flows in the opposite direction of the secondary air, whereas in cross-flow, the primary air flows in the transverse direction of the secondary air. Counter-flow DPIEC systems have demonstrated better cooling effectiveness and capacity than the cross-flow DPIEC [14], despite larger pressure drops. However, this comes at the cost of higher electrical power consumption [15]. Sohani et al. [16] found out that the counter-flow configuration performed better in arid areas, while in other climates the cross-flow configuration was the better alternative. On contrary, [17] obtained higher exergy efficiency for the cross-flow configuration, in addition to less water consumption, with comparable electricity consumption and carbon emissions than the counter-flow configuration. In addition, as reported in [18], the cross-flow configuration can potentially save more installation space. Moreover, efforts to fabricate and commercialize counter-flow DPIECs have encountered several difficulties at the early design phase [15]. This demonstrates that cross-flow DPIECs are more preferred in the market. Additionally, the complex design of the counter-cross configuration may disturb the air distribution, as per a numerical analysis by [19]. In fact, it revealed that a non-uniform air distribution in complex geometries reduced the efficiency of air cooling. As an alternative, [20] proposed a counter–cross-flow DPIEC that consists of both counter- and cross-flow channels. On the one hand, the performance of this device is comparable to that of the counter-flow DPIEC, with the advantage of reduced volume and weight of the cooler, while providing the same air volume. Other alternative configurations were proposed by [21–23]. In particular, [22] tested a mixedflow DPIEC in which the secondary air flowed into a plate without a constrained direction. They found that this configuration ensured better Nomenclature Acronyms and abbreviations EC Evaporative cooler COP Coefficient of performance DEC Direct evaporative cooler HMX Heat and mass exchange IEC Indirect evaporative cooler DPIEC Dew-point indirect evaporative cooler AHU Air handling unit OP-in Orifice plate in the inlet channel OP-out Orifice plate in the outlet channel TH-in Temperature and relative humidity sensor in the inlet air entrance TH1-out Temperature and relative humidity sensor in the primary air exit TH2-out Temperature and relative humidity sensor in the secondary air exit Tw Temperature sensor in the water tank CoF counter-flow CrF cross-flow Symbols L d Dry channel length (cm) W d Dry channel width (mm) H d Dry channel height (mm) L w Wet channel length (cm) W w Wet channel width (cm) G w Wet channel gap (mm) ˙ mMass flow (kg/h) RWorking-to-intake air ratio TTemperature (◦C) RH Relative humidity (%) ˙ Vair volume flow (m 3 /s) K Characteristic of OP ΔpPressure drop (Pa) vSpecific volume (m 3 /kg) ε Effectiveness ˙ qCooling capacity (W) A Area (m 2 ) V velocity (m/s) Subscripts tot Total air pr Primary air sec Secondary air wb Wet bulb dp Dew point w Water HT Heat transfer A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 2 cooling performance and effectiveness than the conventional ones. Although this device is expected to be facile and inexpensive to produce owing to its simpler design, the study did not analyse the compactness, cost, or water consumption of the system. 1.2.2. Materials The material selected that can facilitate high heat transfer between wet and dry channels, as well as an even distribution and storage of water on the wet side [24]. Regarding the thermal conductivity of the plates, thicknesses below 0.25 mm would result in a low thermal resistance irrespective of the thermal properties of the material [25]. Moreover, the use of high-porosity materials that absorb and store water inside the pores would improve the saturation of the working fluid (air) throughout the wet channels. However, humidification of the dry channels must be avoided. The strategy adopted in recent studies [22,24,26–28] involved selection of the plate material and a wicking material to cover the surface of the wet channels. The plate material forms the rigid, low-thermal- resistance, and impermeable heat exchanger. With this aim, Al, polyvinyl chloride (PVC), and polypropylene (PP) are commonly used as the plate material. Meanwhile, the function of the wicking material is to absorb and retain water, while also possessing a low thermal resistance. Fibrous materials are commonly used as wicking materials [29]. Among them, cloth fabrics are widely preferred for their availability, durability, and pliability. Moreover, they should induce lower thermal resistance and pressure drop [7]. Reference [27] used a PP plate coated with nylon fabric on the wet side. Similarly, [22] used Al coated with cotton. Instead, [26] used PVC sheets coated with an unspecified hydrophilic material. Reference [24] experimentally tested several fibrous materials: flocking fibres, Coolplus fibre, Spunlace nonwoven fabrics, and Kraft paper. Ultimately, they chose Coolplus fibre to coat the Al channel walls. As an alternative, the fibrous materials can be used to form a composite membrane material along with a hydrophobic material on the side of the dry channel [29]. For example, [30] used a polymer–plantfibre composite material and a plastic bracket. Notably, the performance of the material used in DPIEC is tested to determine the water absorption and retention [29], which might influence the saturation rate of the secondary air. However, the literature contains few studies that have compared the performances of different materials by applying them to DPIEC. For example, [20] compared a prototype made of Al with the same prototype made of polystyrene covered with nylon. The performance of the DPIEC was evaluated in terms of temperature drop and dew-point effectiveness. In addition, other aspects should be considered in the selection of channel materials, e.g., durability, workability, availability, last, and costs. 1.2.3. Water supply system The water supply system must uniformly distribute the water droplets through the wicking material to enhance the saturation of secondary air and reduce heat transfer resistance between the wet and dry paths [31]. It must also minimize the water and electric consumptions, as well as allow free passage of air throughout the channels. The most used water-supply system in DPIECs consists of a water spray installed at the top of the wet side. However, this type of system exhibits the disadvantages of large water consumption, water droplet drift, and non-uniform water distribution [24]. To overcome these challenges, studies on IECs, such as [32] and [33], focused on optimizing the nozzle configuration. In particular, [32] experimentally tested and compared five commonly used spray nozzles (spiral, conical, square, sector type, target impact) and found out that the spiral type enhanced the uniformity of distribution the most. Additionally, they proposed an intermittent spray to save water and energy consumption. Reference [33] proposed and experimentally validated a three-dimensional spray model based on computational fluid dynamics. Thus, they designed an optimal configuration based on the droplet diameter (0.25 mm), pressure (1.5 bar), flowrate (5.4 l/min), spray cone angle (68◦) and distance between nozzles (80 mm). The study found out that the optimized solution could increase the wet-bulb efficiency and COP by 15.1 % and 17.6 %, respectively. A recent study on IEC by [34] proposed intermittent spraying instead of continuous spraying. Although the performance decreased, a noticeable reduction in energy consumption was observed. To improve the water distribution in DPIECs, [35] constructed a water distributor composed of a series of small tubes drilled and inserted inside the wet channel. Along with the choice of a water-holding material, this device adopted an intermittent water-supply scheme to minimize the water usage and water pump-power consumption. However, this system was expected to induce problems such as a large pressure drop and channel blockage [24]. Alternatively, [24] leveraged the capillary action of wet materials to design an automatic wicking DPIEC. This system was composed of a parallel dry channel and wet channel placed vertically and a water channel installed at the bottom of the wet channel, coated with a wet material. Thus, water was wicked automatically by the wet material through capillary action by forming a stable water film in the wet channel. Reference [36] studied the optimal configuration between the water supply and the working fluid (air) flow. They observed that higher evaporation was achieved when the air flowed to the water distribution in a crossflow manner. Moreover, [37] numerically evaluated the performance of a DPIEC by changing the supplied fluid to the wet channel. They determined that using a hybrid nanofluid instead of water provided a significant improvement in combination with surface modification, but it was not very fruitful in the case of flat channels. 1.3. Originality and scope of the article The literature reveals that more complex and high-performance designs of DPIEC systems in aspects including flow configuration, channel materials, and water-supply system have been investigated in recent years. Regarding flow configuration, the counter-flow configuration performs the best, despite being more complex in construction than the cross-flow configuration. A mixed configuration was proposed to balance effectiveness and ease of design. However, it was not compact. This study proposed a compact mixed-flow configuration and compared its performance with those of the current mixed-flow design. Regarding material selection, the prototype of the proposed system has a design similar to the existing one: the use of plate materials covered with a wicking material on the wet side. However, unlike with the previous design, we use polycarbonate as the plate material. Two wicking options were studied through direct comparison of their effects on the prototype’s performance. Notably, published works on the wicking materials did not study their performance directly on a tested prototype. Finally, concurring with existing knowledge, uniform water distribution is chosen. While most studies analysed different spray devices, a few also proposed direct supply of water into the wet channels through small pipes. This study compared the performances of the prototype with three different water distributors: two spray-type distributors, one placed outside and the other inside the wet channels, and an inside pipeending distributor. The experiments were carried out by varying the inlet air temperature, humidity, volume flow, and working-to-intake air ratio. The performance of the proposed DPIEC was evaluated in terms of temperature drop, dew-point effectiveness, wet-bulb effectiveness, and cooling A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 3 capacity. The results were validated against data from published work. 2. Methodology 2.1. Prototype design and construction In this study, a novel DPIEC is developed and experimentally tested. The device is composed of overlapping modular elements made of polycarbonate (Fig. 1). The base element consisted of dry channels and a wet plate, with the latter covered with a wicking material on one side. At the end of the dry channels, a few air paths were blocked and holes were drilled to drive the primary air through the wet channels, forming the secondary air stream. Next, the secondary air stream was vented through the top opening of the device (Fig. 2). As the air stream through the wet channels moves in a hybrid counter–cross-flow configuration, the device is more appropriately defined as a mixed-flow-configuration DPIEC. A water distributor was installed at the top of the device to supply water to the wet channels. Fig. 3 illustrates a photograph of the assembled device, along with the inlet and outlet of the air flow. The device is tested without a wicking material and for two different cloth materials: synthetic and cotton fabrics, as shown in Fig. 4. Moreover, different types of water distributors were tested (Fig. 5). Distributors 1 and 2 consisted of a circular pipe with additional shorter pipes that penetrated the wet channels. Distributor 1 distributed water simply under the influence of gravity, while Distributor 2 had small nozzles at the end of the pipes to spray water and distribute water more uniformly through the wet channels. Distributor 3 consisted of a circular pipe equipped with nozzles that uniformly spray water over the top of the wet channels. Videos of the operation of the three distributors are provided as supplementary material with this manuscript. The structural parameters of the device are summarized in Table 1. 2.2. Experimental set up The test bench is composed of an air handling unit (AHU), a water tank, a water pump, connection ducts, orifice plates for airflow measurement (OP-in and OP-out), and temperature and relative humidity sensors (TH-in, TH1-out, TH2-out, Tw), as displayed in Fig. 6. A water tank was placed below the prototype, and a submergible pump guaranteed the circulation of water till the water distributor. The power consumption of the pump ranged from 10–20 W with a maximum flow rate of 16 l/min. Moreover, the prototype was connected to the AHU via flexible air ducts with a diameter of 25 cm. An additional duct with the same characteristics transported the supply air from the prototype to the laboratory. The working air ratio of the secondary air to the primary air flow was controlled with a dumper placed at the exit of the primary airstream duct. First, the air exhausted from the laboratory was treated in the AHU to reach the required test conditions. Subsequently, it was transported to the prototype. In the dry channels, a fraction of the primary air treated in the prototype was diverted into the wet channels to act as the working (secondary) air. The fraction of the unused treated (primary) air was then pumped to the target indoor space. During the test, the following factors were monitored: •total, primary, and secondary air flow rates ˙ mtot,˙ mpr, and ˙ msec; •total, primary, and secondary air temperatures Ttot,Tpr, and Tsec; •total, primary, and secondary air relative humidity RHtot,RHpr, and RHsec; •water temperature Tw. The air mass flow rate is determined indirectly through the pressure drop measured at the previously calibrated orifice plates, while the other measurements (air dry-bulb temperature, air relative humidity, and water temperature) were measured with the previously calibrated sensors described in Table 2. To ensure steady-state conditions, all measurements were taken after at least 20 min of operation the experimental set up. To measure the inlet and outlet primary air flow characteristics, two orifice plates were placed between the AHU and the device (OP-in in Fig. 6) and between the device and the supply duct (OP-out in Fig. 6), respectively. The orifice plate allows for calculating the air volumetric flow rate ˙ V by measuring the pressure difference Δpupstream and downstream of the plate (Equation 1). ˙ V=k• ΔP √(1) The values of coefficient kin (Equation 1) is a characteristic specific to the orifice plate. Δpwas measured by datalogger Testo 435–4 as the mean value of the instant values of static pressure after 30 s. The measurement was conducted after a period of at least 20 min Fig. 1. Three-dimensional (3D) view (a) and photographs (b) of the base element. A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 4 (corresponding to 120 measurements) since the activation of the AHU to guarantee a constant air temperature. Furthermore, to avoid modification of the flux line next to the orifice, the pressure was measured downstream at almost four times the diameter of the ducts. From this point, the air volumetric flow rate ˙ Vof the inlet and outlet primary air were converted to the air mass flow rates by calculating the specific volume v(Equation 2 and 3). ˙ m=˙ V/v(2) Fig. 2. Schematic section of the prototype and 3D flow paths. Fig. 3. Photograph of prototype. elements assembled. A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 5 Specific volume vwas obtained through the known psychrometric equations [38] by measuring the temperature and relative humidity for each stage and by considering the altitude (690 m.a.s.l.) of Valladolid, Spain where the laboratory is located. Thus, by considering the mass balance in (Equation 4), the secondary air mass flow is determined (Equation 5). ˙ mtot =˙ mpr +˙ msec (3) ˙ msec =˙ mtot −˙ mpr (4) Furthermore, the working-to-intake air ratio R is calculated as follows (Equation 6): R= ˙ msec ˙ mtot =1− ˙ mpr ˙ mtot (5) The air dry-bulb temperature and relative humidity were directly measured using the sensors described in Table 2. These sensors were located inside the duct at the entrance of the HMX (TH-in in Fig. 6), inside the duct at the exit of the HMX (TH1-out in Fig. 4. View of materials tested. Fig. 5. Photographs and schematics of the tested water distributors. Table 1 Description of the structural parameters of the prototype. Parameters Symbol Specification/Values Volume of Heat and Mass eXchange (HMX) 31 cm ×25 cm ×30.5 cm Flow configuration Mixed flow Dry channel length L d 30 cm Dry channel width W d 9 mm Dry channel height H d 9 mm Number of dry channels N d 8 x 28 Wet channel length L w 30 cm Wet channel height H w 30 cm Wet channel gap G w 9 mm Number of wet channels N w 8 Plate thickness 0.5 mm Channel material Polycarbonate Wicking material None, synthetic cloth, cotton cloth Water distributor type 1, type 2, type 3 A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 6 Fig. 6), and at the top of the HMX at the exhaust (TH2-out in Fig. 6), respectively. The measurements were recorded every 10 s during the experiment. The water temperature was directly measured using a sensor (Tw in Fig. 6) installed inside the water tank. 2.3. Testing of the prototype The prototype was tested by varying the wicking material and the water distributor, under different inlet air temperature conditions, Fig. 6. Experimental set up (a) photograph and (b) scheme. Table 2 Description of sensors (accuracy, model, measured parameters). Parameter Instrument Range Accuracy Static pressure Data logger Testo 435–4 0–25 hPa ±0.02 hPa (0–2 hPa) ±1% of measured values (remaining range) Air dry bulb temperature Testo 175 H1 Temperature and relative humidity sensor −20–+55 ◦C±0.4 ◦C Air relative humidity Testo 175 H1 Temperature and relative humidity sensor 0 %–100 % ±2% RH (from 2 %–98 % RH) at +25 ◦C Water temperature Testo 175 T2 Temperature sensor −35–+55 ◦C±0.5 ◦C Table 3 Description of the operative conditions in each experimental test. Test Wicking material Water distributor Ttot RHtot ˙ mtot R Material Cotton cloth None Synthetic cloth Type 3 30, 35 ◦C No control 17–46 % 389 ±11 kg/h 0.5 Distributor Cotton cloth Type 1 Type 2 Type 3 30, 35, 40 ◦C HR control 50 % No control 26–49 % 367 ±11 kg/h 0.5 Operative conditions Cotton cloth Type 3 Ambient, 30, 35, 40 ◦C 25–48 % 394 ±17 kg/h 0.4–0.8 A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 7 without humidity control. Further, some experiments were repeated by setting the relative humidity at 50 %. Further experiments were performed to evaluate the effect of the operative conditions. Table 3 lists the experiments performed and the respective operative conditions. 2.4. Performance evaluation To evaluate the performance of the DPIEC and compare it with published data, the temperature drop, effectiveness, and cooling capacity were calculated based on the experimental values. In particular, the temperature drop is defined as the difference between total air temperature Ttot and primary air temperature Tpr. ΔT=Ttot −Tpr (6) Regarding the effectiveness, three different definitions were considered. In all cases, the temperature drop was compared to a maximum depression, which is representative of an ideal condition. Therefore, by considering the wet-bulb depression in the primary channel (Equation 7) as the ideal temperature drop, wet-bulb effectiveness ε wb is expressed as Equation 8. WBDpr =Ttot −WBTtot (7) ε wb =ΔT WBDpr (8) where WBTtot is the wet-bulb temperature of the total air mass. By considering the dew-point depression (Equation 9) as the ideal temperature drop, dew-point effectiveness ε dp is expressed as Equation 10. DPDpr =Ttot −DPTtot (9) ε dp =ΔT DPDpr (10) where DPTtot is the dew-point temperature of the total air mass, which is the same as that of the secondary air under the hypothesis of no mass exchange between the primary and secondary channels. Finally, by considering the wet-bulb depression in the secondary channel (Equation 11), a second wet bulb effectiveness ε wbʹis expressed as Equation 12. WBDsec =Ttot −WBTsec,in =Ttot −WBTpr (11) ε wbʹ=ΔT WBDsec (12) where WBTsec,in is the wet-bulb temperature of the air at the entrance of the secondary channel, which is the same as that of primary air WBTpr. Hence, the cooling capacity is calculated by Equation 13. ˙ q=˙ mpr(ca+xincv)(Ttot −Tpr)(13) where ˙ mpr is the primary air flow rate, xin is the inlet air humidity ratio, and cma is the specific heat of moist air, which is expressed by Equation (14). cma = (ca+xincv)(14) where caand cvare the specific heat values of the dry air and vapor, respectively. 2.5. Uncertainty analysis Uncertainties in the values of the studied parameters were determined through the root sum squared method [39]. For the temperature drop, because the same type of sensor was used for both temperature measurements: uΔT= ( ∂ ΔT ∂ Ttot uT)2 +( ∂ ΔT ∂ Tpr uT)2 √(15) where the uncertainty of the measuring equipment (uT) corresponds to the accuracy of the calibrated sensor. As indicated in Table 2, the accuracy of the temperature sensor was ±0.4 ◦C, and accuracy of the temperature measurement was improved by calibrating it to uT= 0.2◦C. Thus, the uncertainty in the temperature drop became 0.28 ◦C. For the cooling capacity ( ˙ q), because the inlet humidity ratio in the primary airstream was maintained at a constant value, the specific heat of the moist air can be considered a constant. Having the primary air mass flow calculated by Equations (1) and (2), the uncertainty of the cooling capacity is uq= ( ∂ q ∂ ΔPuΔP)2 +( ∂ q ∂ Ttot uT)2 +( ∂ q ∂ Tpr uT)2 √(16) Since ∂ q ∂ Ttot and ∂ q ∂ ΔPare increasing functions, but ∂ q ∂ Tpr is a decreasing function, the maximum uncertainty must be calculated in terms of the maximum values of ΔP and T tot , but minimum T pr . Given the uncertainty of the instrumentation (uT=0.2◦C, uΔP= 2Pa), the uncertainty of the cooling capacity resulted in 24 W. Considering the psychrometric equations required to obtain the dewpoint temperature of inlet (total) air [38], the uncertainty of dew-point effectiveness: u ε DP = ( ∂ε DP ∂ HRuHR)2 +( ∂ε DP ∂ Ttot uT)2 +( ∂ε DP ∂ Tpr uT)2 √(17) For the uncertainty introduced by the accuracy of the instrumentation (uT=0.2◦C, uHR =0,02), the uncertainty of dew-point effectiveness is 0.06. Finally, the root sum squared method cannot be applied to formulate the wet-bulb effectiveness, given the iterative calculation of the wetbulb temperature. Therefore, uncertainty was estimated as the maximum difference between the values obtained from the measurements and those achievable within the limits of accuracy of the measuring equipment (uT=0.2◦C, uHR =0,02). The results show that the calculated values were always conservative and differed by less than 0.06. 3. Results and discussion 3.1. Test results As an example of the experimental procedure and data logging, Fig. 7 presents the temperatures and relative humidities registered during one test. The test conditions were: air flow rate =350–391 kg/h; working-to- intakeair ratio of 0.5; and primary air inlet temperatures =25, 30, 35 ◦C, and 40 ◦C under uncontrolled humidity conditions and 35, 30, and 25 ◦C under relative humidity of 50 %. Steady-state operating conditions were identified, and the average values were determined to calculate the performance of the system. 3.2. Selection of the optimal wicking material and water distributor This section describes the results of the experiments performed to determine the optimal wicking material and water distributor. In particular, the wicking material was investigated by considering the device with Distributor 3. The test was repeated for the inlet temperature range of 30–35 ◦C without RH control. Therefore, based on thermal-hygrometric conditions of the laboratory, the relative humidity A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 8 of the inlet air ranged from 17 %–46 %. The inlet mass air flow rate was 389 ±11 kg/h, and the working-to-intake air ratio was approximately 0.5 for each test (Table 3). Table 4 lists the cooling capacities achieved by the tested device as a function of the inlet air temperature. The results in the table confirm that the presence of a cloth on the surface of the wet channels improves the performance of the prototype. This could be expected for the enhanced water distribution over the working area. However, the material of the fabric does not significantly influence the result. Compared to the case the wicking material, wicking by the synthetic cloth improves the cooling capacity by 27 % and 30 % at 30 and 35 ◦C, respectively. Meanwhile, wicking by the cotton cloth improves it by slightly higher margins (36 % and 45 %, respectively). Without the wicking material, the achievable temperature drop and wetbulb effectiveness decrease by 0.9 ◦C and 0.09 at the inlet air temperature of 35 ◦C, respectively. Because the difference in cooling capacity between the wicking material types is lower than its accuracy, the use of cotton cloth over synthetic cloth cannot be justified in terms of the improvement it offers. Consequently, the selection of the type of wicking material should be evaluated in relation to its cost, availability, and maintenance. However, it is worth noting that these cloths are easily accessible in the market or available as reused material; however, the possible decay of natural fibres must be also considered. In the following experiments, cotton cloth Fig. 7. Example of registered data during a test. Table 4 Cooling capacity of different wicking materials. Inlet air temperature (◦C) Cooling capacity (W) No wicking material Synthetic cloth Cotton cloth 30 122.8 156.5 166.8 35 157.0 203.5 227.0 A. Urso et al. Applied Thermal Engineering 257 (2024) 124294 9