Numerical and experimental investigation of the indoor air quality and thermal comfort performance of a low energy cooling windcatcher with heat pipes and extended surfaces

Renewable Energy 145 (2020) 744e756

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Numerical and experimental investigation of the indoor air quality and thermal comfort performance of a low energy cooling windcatcher with heat pipes and extended surfaces John Kaiser Calautit a, *, Paige Wenbin Tien a, Shuangyu Wei a, Katrina Calautit a, **, Ben Hughes b a b

Department ofArchitecture and Built Environment, University of Nottingham, Nottingham, NG7 2TN, UK Department of Mechanical and Aerospace Engineering, University of Strathclyde, Glasgow G1 1XJ, UK

a r t i c l e i n f o

a b s t r a c t

Article history: Received 19 November 2018 Received in revised form 1 June 2019 Accepted 10 June 2019 Available online 18 June 2019

This work builds on previous experience in windcatcher design, maximising the ventilation rate of the windcatcher, whilst integrating low energy cooling technologies. The present study aims to investigate the thermal comfort and indoor air quality in buildings ventilated with a passive cooling windcatcher integrated with heat pipes and extended surface using numerical modelling, wind tunnel and far-field testing in the UAE during a summer month. Results of the scaled wind tunnel tests showed that the addition of the heat pipes and extended surfaces reduced the airflow through the windcatcher but did not impede the flow even at low outdoor wind speeds, this was further confirmed by the smoke visualisation tests. Analysis of pollutant concentration in the building model showed that the proposed windcatcher configuration was capable of delivering fresh air at a sufficient rate to lower CO2 concentration levels below the recommended guidelines for air quality. The thermal comfort analysis was conducted and it was observed that for the present design, equal distribution of thermal comfort was not achieved due to combination of high air movement, colder temperature and high humidity below the windcatcher which resulted in higher thermal discomfort in this area. Further work is required to develop a suitable control strategy in the form of volume control dampers which would enable the supply flowrate to be monitored and altered as required and optimized the distribution in the occupied space. Field tests data was used to validate the numerical modelling, showing good agreement between both methods. © 2019 Elsevier Ltd. All rights reserved.

Keywords: Built environment Computational modelling Extended surfaces Heat pipes Passive cooling

1. Introduction According to the World Business Council for Sustainable Development, major cities in Brazil, China, EU, India, Japan and USA, which represent the world's six largest economic regions are responsible for up to 46% of the entire energy use [1]. As a means to counter the rising emissions, a strong drive for more sustainable and energy-efficient technologies has gained popularity amongst users and consumers. In addition, governments around the world have committed to cut greenhouse gas emissions from pre-1990

* Corresponding author. ** Corresponding author. E-mail addresses: [email protected], (J.K. Calautit). https://doi.org/10.1016/j.renene.2019.06.040 0960-1481/© 2019 Elsevier Ltd. All rights reserved.

[email protected]

levels by 80% by the year 2050 [1]. Other organisations such as the IEA analysed future scenarios to devote greater attention towards energy access issues [2]. Through the developments of the ‘Sustainable Development Scenario’ and the ‘450 Scenario’ within [3], plausible path to achieve climate targets was demonstrated, ensuring sustainable and modern energy services will be achieved by 2030 through use of the available technology solutions and considerations of the impacts on human health [4]. This demonstrates the strong drive for more sustainable and energy efficient technologies. To achieve this ambitious reduction in emissions is not unfeasible but requires a societal movement away from energy intensive processes and a move toward low energy and vase development in zero energy technology. The building sector is one of the largest consumers of energy (40% of the world energy use and over 33% of the global greenhouse gas emissions) and increasing at unprecedented scales which presents a huge

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opportunity to significantly cut the energy demand and CO2 emissions. A main issue in the design and planning of more sustainable cities is the usage of new technologies to lessen the energy demand of buildings. The rapid growth of countries in the Middle East such as Qatar, Kuwait and UAE placed them at the top of the World Bank's CO2 emissions per capita list [5]. A recent study in the Abu Dhabi city highlighted that air-conditioning loads represents up to 70% of the total electricity demand [6]. The extreme hot climate, the availability of cheap power in these countries and the increasing standards/requirements for comfort have led to the usage of airconditioning in nearly all buildings, where most of these units are operated all year round. Heating, Ventilation and Air-conditioning (HVAC) is responsible for up to 60% of the energy consumed by buildings in the domestic sector [7] and even higher during summer months when outdoor temperature can reach up to 40
C [8]. Clearly, minimising or eliminating the usage of air-conditioning and its energy requirement has the potential to substantially lower the country's emissions and allow to achieve its sustainability goals. However, it is also important that thermally comfortable and safe indoor spaces are provided to the occupants [9]. Though developments have been made to mechanical ventilation systems to improve efficiency, substantial energy requirements are still necessary for operation. Passive ventilation used in conjunction with low energy cooling technologies provides the opportunity to mitigate the energy requirements of mechanical HVAC systems through the use of wind and buoyancy driven flows in the supply of air [10]. An example of such technology which uses both forces of wind and buoyancy to provide airflow is a wind tower [11]. The windcatcher is an architectural component of traditional Middle Eastern buildings [12] which captures the outdoor wind at high elevation and directs it to the interior [13]. In temperate/mild climates such as in the UK, modern version of the windcatcher are incorporated in schools and offices [14] to achieve the essential levels of indoor air quality. The ventilation strategy used in windcatchers is particularly effective for night time cooling as this is when there is the greatest differential between indoor and outdoor temperatures. Night time cooling can be used to dissipate the heat stored in the building fabric thereby moderating internal daytime temperatures [15]. However, unlike mechanical cooling, windcatchers are ineffective at reducing the temperature of supply air. Since it primarily relies on the outdoor temperature for cooling, it presents limitations in the application of existing designs in hot regions and could cause further discomfort to occupants during the summer if the air is not pre-cooled before entering into occupied spaces. Previous studies have attempted to address this by integrating evaporative cooling and heat transfer devices. Using a windcatcher with evaporative cooling for a particularly dry climatic condition (10e35% relative humidity), enabled the reduction in air temperature up to 15
C [16]. However, there are several drawbacks associated with evaporative cooling such as high operation and maintenance cost, effectiveness in warm, moderate to high humid conditions [17]. In addition, evaporative coolers use a substantial amount of water to run which could be considered wasteful in arid climates. The study [18] incorporated heat transfer devices into the internal channel of the windcatcher to reduce the supply airflow temperature (see Fig. 1). 1.1. Significance and research objectives The device functions by capturing hot outdoor airflow at roof level and redirecting it through a series of cylindrical heat pipes. This ensures the heat from the incoming hot air to evaporate the working fluid in the sealed heat pipes. The vapour then flows to the

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other end and when it condenses; it gives up the heat to the fluid that circulates within the inside of the closed-loop cool sink. Consequently, this will maintain the operating conditions and repeat the cyclic operation of the heat pipe. Dampers are mounted at the bottom of the unit to control the delivery rate of air, as fluctuations in external wind speed greatly affect the air movement rate. The cooled air is supplied to the room below the channel via the ceiling diffusers. In the previous studies [19], a low energy cooling system that integrates a natural ventilation windcatcher and arrangement of heat pipes was designed and manufactured to a full-scale prototype for field tests. It was envisioned that the system will replace or complement energy-intensive air-conditioning units in warm-hot conditions. Like the modern commercial windcatchers in the UK, the device is designed to be compact and easy to retrofit to existing buildings. The prototype cooling system comprised of a roof mounted passive ventilation component consisting of a windcatcher combined with a series of heat pipes that can deliver fresh air that is 4e12
C cooler than ambient temperature, depending on wind conditions [19]. It operates by capturing hot outdoor airflow at higher level (roof) and redirecting it through its channel where cylindrical heat pipes are located. The supply airflow temperature is reduced by the heat pipes by absorbing the heat from the incoming airflow and transferring it to the fluid in the cool sink as shown in Fig. 2. Results of previous research showed promising reduction of supply air temperature but there was limited analysis of the indoor environment quality in the space ventilated by the device. Hence, the present study aims to address the research gap by demonstrating its capabilities in providing good indoor air quality and adequate comfort. The effect of hear transfer enhancement on the outside surfaces of heat pipes will be investigated. Numerical 3D CFD modelling, validated with full scale field measurements, will be used to carry out the simulation and assessment of the natural ventilation and cooling potential of the proposed windcatcher device under hot climatic conditions. For more accurate representation of the wind, an atmospheric boundary layer (ABL) flow profile was implemented with the computational domain set based on best practice guidelines urban wind simulations. The analysis of thermal comfort in the ventilated space will be carried out. The simplified exhalation model will be used for the study of the influence of the proposed windcatcher on the distributions of indoor CO2 levels. 2. Literature review For many centuries, natural ventilation systems such as windcatchers have been used in Middle East to deliver fresh and cooler air into buildings [11]. The design of windcatcher has been customarily based on the location, climate, architects’ experience and status of the inhabitants. Different designs of wind tower can be seen across the Middle East i.e. varied height, size and form, number of openings, construction and mounting location [11]. Developments to the windcatcher have been explored to increase airflow through the windcatcher along with adjustments to improve the cooling operation. Nejat et al. [20] evaluated the integration of a two-sided windcatcher and wing wall using both experimental scaled wind tunnel testing and CFD simulation for the areas with low wind speeds. Their results implied that the windcatcher with a wind wall angle of 30
can achieve a 50% higher ventilation performance than a conventional windcatcher while the length of the wind wall has an inconsiderable effect on it. Bahadori [21] modified the tower head to prevent the airflow from short-circuiting or leaving the other openings, adding a column for increasing the heat transfer area and integrating evaporative

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Fig. 1. (a) A roof mounted windcatcher with heat transfer devices (b) schematic diagram of the windcatcher operation.

cooling. A theoretical model was used to analyse the proposed system under various environmental settings. Bansal [22] coupled a wind tower and a solar updraft tower to increase ventilation in a building. The analytical model predicted that the effect of the solar chimney is much greater for lower winds as compared to the wind tower. Grosso and Ahmadi [23] conducted parametrical analysis of a uni-directional windcatcher model located in Italy using SPERAvent dynamic thermal simulation tool. The study concluded that the windcatcher can save 43e61% of the annual cooling energy

requirements. Satwiko and Tuhari [24] carried out numerical and wind tunnel analysis of a longitudinal hybrid windcatcher for ventilating a basement in warm-humid climates. Haghighi et al. [25] proposed a windcatcher integrated with a solar-driven adsorption chiller to cool a two-story building. Numerical results indicated that 1800W cooling demand for each 50m3 room is required to achieve adequate comfort levels. Kang and Strand [26] carried out parametric analysis of an evaporative cooling windcatcher under a wide range of climatic conditions using CFD

Fig. 2. (a) Full scale prototype of the passive cooling windcatcher. (b) 3D schematic showing the interior of the system (c) proposed windcatcher systems incorporating horizontal heat transfer devices with extended surfaces.

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modelling. Commercial windcatchers are generally compact, light-weight and relatively smaller in size as compared to traditional systems. It is also integrated with volume control systems, sensors and PV panels. The current development of numerous features of modern windcatchers to advance the design and ventilation efficiency has been well studied. Optimum designs of the shape, louvers and dampers have all been explored using a combination of experimental testing and numerical simulation [11]. Several researchers have improved the operation of the windcatcher for different climates. A passive cooling device employing the combination of heat pipes and cool sink has been developed and incorporated into a roof mounted windcatcher for hot climates [18]. Recently, O'Connor et al. [27] explored the potential of integrating a passive (low pressure drop) heat recovery wheel into a windcatcher to recover the heat from the outlet channels and transfer to the inlet or incoming airflow. The proposed modification will enable consistent use during cooler months in temperate climates, improving the year-round capabilities of the windcatcher. Vivian et al. [28] proposed a passive ventilation system with heat recovery (PVHR), which is composed by a roof cowl, a coaxial heat exchanger and a flow splitter. The experimental results indicated that in both heating and cooling seasons the indoor temperature and CO2 concentration can meet the requirements of UK guidelines through its suitable performance. O'Connor et al. [29] also developed a passive rotary dehumidification system for windcatchers and other passive ventilation applications in humid climates. The results highlighted that the device was able to reduce the moisture levels of the supply airstream at a lower regeneration air temperature and with lower pressure drop as compared to existing technologies. Extending the previous research, the present work will investigate the capabilities of a windcatcher integrated with passive cooling technologies to determine the effect on the air temperature and hence the thermal comfort of occupants. No previous work has conducted analysis of heat pipe integrated windcatcher units with a focus on thermal comfort and indoor air quality of the occupied space. This was carried out by measuring the CO2 concentration as an indicator of the flow around the building from the windcatcher. By analysing the windcatcher system in isolation the effect of the modifications cannot be fully assessed. By conducting far-field testing, a better understanding of the influence of the windcatcher can be made. Furthermore, there is no study which enhanced the cooling performance by modifying the design of heat transfer devices in the proposed windcatcher system. 3. Materials and methods This section introduces the numerical and experimental methods used to evaluate the performance of the cooling windcatcher system. The numerical modelling section introduces the theory, modelling of the geometry and domain, generation of the grid and verification and setting up of the boundary conditions. Section 3.2 details the setup of the lab scale experiment which utilised a low speed wind tunnel and scaled 3D printed models. Then, section 3.3 details the manufacturing of the 1:1 prototype and the field test carried out in a hot climate city. 3.1. Numerical modelling The numerical simulations were completed using ANSYS commercial software FLUENT. The assumptions for the RANS simulation included a 3D, fully-turbulent and incompressible flow. The modelling of the turbulent nature of the flow was carried out using the standard keε model, which is well recognised in the field of indoor ventilation and windcatcher research [11]. The Finite

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Volume Method approach and Semi-Implicit Method for PressureLinked Equations velocity-pressure coupling algorithm with the 2nd order upwind discretisation were employed by the CFD tools. The governing equations for the 1-continuity, 2-momentum, 3energy, 4-turbulent kinetic energy, 5-energy dissipation rate are summarised below:

vr þ V  ðruÞ ¼ 0 vt

(1)

where r is density, t is time and u is fluid velocity vector.

vðruÞ þ V  ðruuÞ ¼ Vp þ rg þ V  ðmVuÞ  V  tt vt

(2)

where p is the pressure, g is vector of gravitational acceleration, m is molecular dynamic viscosity and tt is the divergence of the turbulence stresses which accounts for auxiliary stresses due to velocity fluctuations.


 X vðreÞ þ V  ðreuÞ ¼ V  keff VT  V  ð hi ji Þ vt i

(3)

where e is the specific internal energy, keff is the effective heat conductivity, T is the air temperature, hi is the specific enthalpy of fluid and ji is the mass flux.

h i vðrkÞ þ V  ðrkuÞ ¼ V  ak meff Vk þ Gk þ Gb  rε vt

(4)

h i vðrεÞ ε ε2 þ V  ðrεuÞ ¼ V  aε meff Vε þ C1ε ðGk þ C3ε Gb Þ  C2ε r vt k k (5) where Gk is the source of TKE due to average velocity gradient, Gb is the source of TKE due to buoyancy force, ak and aε are turbulent Prandtls numbers, C1ε , C2ε and C3ε are empirical model constants. The modelling work was conducted on a workstation with dual processor Intel Xeon 3.3 GHz and 24 GB Fully-buffered DDR2.

3.1.1. Computational domain The surrounding domain and set boundary conditions are shown in Fig. 3. The domain consisted of a macro-climate or enclosure to simulate the outdoor airflow around the windcatcher model mounted on top of a micro-climate or small room with the height H. The windcatcher is mounted on the windward side of the building to maximise the crossflow ventilation across the openings. It should be noted that the positioning of the windcatcher on the roof can significantly affect its ventilation performance along with other external factors such as the shape of the roof and building [30], other wind catchers [31] and other buildings which should be taken into account when selecting the position. The present work won't cover the effect of these parameters/factors on the passive cooling windcatcher. Sizing of the enclosure was conducted following the best practice procedures for urban wind flow simulations for single-building model [32]. According to the guideline, the spacing for the inlet boundary, top and side walls should be 5 x H (height of the target building) while the space between the outlet wall and the model should be at least 10 x H as shown in Fig. 4. The indoor domain or room used for the analysis of the windcatcher had a 25 m2 floor area with a height of 3 m (H). The dimensions were selected in order to have the same dimensions as in the experimental study. Maintaining the same dimensions ensures accuracy between the two analysis methods. The windcatcher model had a dimension of 0.33H  0.33H x 0.4H. The louvers at the

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inlet of the windcatcher were angled at 45
and spaced at regular intervals. Fig. 4 shows a 3D diagram of the windcatcher, the spacing between the heat pipes (in mm) and extended surfaces. The horizontal and vertical spacing between the heat transfer devices were 0.05 m and 0.02 m [19]. To assess the impact of the windcatcher on the indoor air quality, indoor carbon dioxide CO2 concentration modelling was conducted. For this purpose, 12 occupants which were equally distributed in the space were located inside the micro-climate as shown in Fig. 4. For simplification, the occupants were modelled as cuboid shape with a height of 1.8 m [33]. The mouth opening area was equal to 0.013  0.010 m2. The exhalation model proposed by Ref. [33] was employed. The average value of exhaled air was assumed as 6 l/min for the simulation [31]. The CO2 present in the atmospheric air was assumed to have a concentration of 382 ppm, typical outdoor CO2 range between 300 and 500 [31]. The air was modelled containing of the following species: oxygen (O2), carbon dioxide (CO2) and water vapour (H2O).

    v rYO2 þ V, rYO2 ¼ V,jO2 vt

(6)

    v rYCO2 þ V, rYCO2 ¼ V,jCO2 vt

(7)

    v rYH2 O þ V, rYH2 O ¼ V,jH2 O vt

(8)

where Yi is mass fraction of i-th air constituent.

3.1.2. Computational mesh The numerical domain was discretised using unstructured mesh techniques with refinements in essential areas, maintaining larger size elements where possible, reducing the total number of elements and calculation times. The grid was refined around the windcatcher, openings and heat transfer devices in order to ensure that the flow field in these regions can be captured accurately. Fig. 5a displays the mesh generated around the surfaces of the model. Mesh verification was used to verify the computational model and mesh independency. The verification of the mesh starts with a coarse mesh and gradually refines it until the variation observed between the results were smaller than the predefined acceptable error. The boundary conditions remained fixed throughout the simulation process to ascertain precise comparison of the results. The verification of the mesh was carried out by running the same simulation setup for three sets of element sizes from 2.2 million (coarse) to 4.34 million (fine) and investigating the

Fig. 3. Computational domain and boundary conditions.

Fig. 4. Model of the windcatcher, heat pipes, extended surfaces and building with cuboids.

effect of the mesh sizing on the results. An example of this is shown in Fig. 5b which compares the air velocity results along the airstream from the outlet of the windcatcher to floor for the coarse, medium and fine mesh. The mean error between the medium and fine mesh results was 3.7% (max 8% which is equivalent to ±0.016 m/s). The medium size mesh was employed for the simulations to have a balance between the speed and accuracy.

3.1.3. Boundary conditions The boundary conditions were set based on best practice guidelines [34] for the numerical modelling of wind flows in the urban environment. The urban wind profile power law exponent was 0.25, corresponding to a suburban terrain [31]. For the kepsilon model, the values of epsilon were obtained by assuming local equilibrium of Pk ¼ ε. The top and side boundaries were set as symmetry. The pressure outlet was set to atmospheric (0 Pa). The standard wall functions were set to the wall boundaries [34]. Based on [35] for the ground surface, the wall functions were adjusted to replicate the influence of ground roughness by utilizing the equivalent sand-grain roughness height ks and roughness constant Cs. Table 1 details the set boundary conditions. Fig. 6 shows the weather data of the surrounding area of Ras-Al Khaimah, UAE where the far-field testing took place. The dominant wind direction is an important consideration in the placement of the windcatcher, particularly if a uni-direction windcatcher is used. As observed in Fig. 6a, the prevailing wind direction was between NW and N, hence the windcatcher opening was oriented to the prevailing wind (NNW). The monthly average wind speed varies between 3.6 and 4.6 m/s, measured at 20 m elevation [36,37]. The highest average temperature was during the months of July and August and lowest during January. The relative humidity data (Fig. 6b) shows lower humidity levels during summer when there is higher cooling demand and higher humidity levels during the winter. Table 2 shows the boundary conditions for the numerical model relating to the CO2 analysis. The recommendations for the CO2 concentration can be found in Ref. [33].

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Table 1 Summary of the CFD model boundary conditions for urban flow model. Velocity inlet (m/s) Pressure outlet Temperature inlet  (C) Relative humidity (%) Gravity (m/s2) Walls (all) Roughness height KS (103 m) Roughness constant CKS

Fig. 5. (a) Generated high-resolution computational grid around surfaces of test room with windcatcher (b) Computational mesh verification of the supply airstream.

3.2. Experimental analysis Two types of experimental tests were conducted: scaled wind tunnel and full-scale field testing. For the wind tunnel tests, a scaled model (1:10) of the windcatcher with heat pipes and extended surfaces was utilised. Fig. 7a shows the setup of the model in the wind tunnel test section which had a cross sectional area of 0.5  0.5 m2. Due to the limitation in size of the test section, only the windcatcher was exposed to the wind tunnel flow while the 0.5 m  0.5 m x 0.3 m test room was located below the test section. It should be noted that for the purpose of validation, a CFD model replicating the wind tunnel setup was also simulated. The blockage ratio produced by the windcatcher model was below 5% and therefore the measurement required no further corrections. As shown in Fig. 7b, all components of the windcatcher were 3D printed except for a side wall and top wall which was constructed using perspex glass for visualising the flow during smoke test. Cylindrical copper tubes were used to model the heat pipes and arranged in the windcatcher. A window is located on the leeward side with the 0.1 m x 0.1 dimension. The room was also constructed using perspex glass for smoke test flow visualisation. The windcatcher can be completely rotated in the test section to simulate varying wind directions. There are several openings in the test room model which allowed for measurements of the airflow velocity using a hot-wire anemometer (uncertainty of ±1.0% of reading). The full-scale 1 m  1 m prototype of the windcatcher was constructed and tested in Ras-Al-Khaimah, UAE during a typical summer month (September). The test site was located in RAK Research and Innovation Center 25
400 04.300 N 55
460 43.900 E. The climate of UAE is generally very hot and sunny with very high temperatures (avg. 35e40
C) from June to September as observed in Fig. 8. The test was conducted during the month of September

ABL flow Atmospheric 21e40 34e67 9.81 No-slip Macro-micro climate walls: 0.001 All walls: 0.5

(17e18, 2014). The time of the test was from 11:00 to 16:00 in order to investigate the performance of the system during peak periods. The wind speed in RAK ranged between 3.6 and 4.6 m/s on average however the wind speeds during the rest reached up to 6 m/s. The predominant wind direction in RAK is from the N-NW direction but during the test, the direction of the wind was completely opposite in the morning (SSE) and changed to NW in the afternoon. It should be noted that the test room and windcatcher openings was oriented towards the predominant wind angle. Similar to the numerical model, the test room had a volume of 27 m3 and was completely empty except for the measuring equipment. For the full-scale tests, vacuum sealed heat pipes charged with water (working fluids) were arranged (staggered grid) in the channel of the windcatcher prototype as detailed in Fig. 4. The dimension of a single heat pipe was 20 mm in diameter and 800 mm in length (equally split in the air channel and cool sink). The working pressures were set to saturation and at an operating temperature of 20
C. In order to maintain the operating conditions of the device, the cold sink was fed with water at around 20
C. The fabric of the test room was fully insulated and an outlet was located at the leeward side. For the measurements of the air temperature, PICO Type K thermocouples (exposed wire type, tip diameter of 1.5 mm and a tip temperature range between 75
C to 250
C) was used. A PICO data logger (uncertainty ±0.6
C at 50
C) collected the thermocouple data and links to a PC for real time monitoring of the data. Temperature was recorded at an interval of 1 s. The RAK Research and Innovation Center weather station was used as a reference and provided other climatic data. 4. Results and discussions The results from the CFD simulations of the uni-direction windcatcher integrated with low-energy cooling technology are presented. The primary conditions that were assessed during the testing were indoor air quality, as a measurement of air supply rate and CO2 concentration, thermal performance, relative to the internal air temperature and moisture content, and the thermal comfort experienced by occupants. According to the ASHRAE Standards 62.1, indoor air quality (IAQ) criteria are fulfilled if the ventilation is adequate to keep the indoor carbon dioxide CO2 level under 1000 ppm. CO2 at low levels of concentration is typically used as an indicator for IAQ and for ventilation requirements, reflecting the loading of pollutant from occupant exhalation. Hence, it is important that enough fresh air is supplied into the space, for example according to the CIBSE Guide A, the recommended fresh air requirement for a typical classroom is 10L/s per occupant. Standard windcatchers in general are able supply the recommended ventilation rates even at low wind speeds (1 m/s). However, the incorporation of heat pipes and extended surfaces in the windcatcher channel presents a pressure drop to the airflow, thus lowering the supply air speed and potentially impede the fresh air supply. This can be observed in Fig. 9a which shows that the airflow speed was reduced by up to 0.40 m/s after moving across the heat pipe arrangement, limiting the windcatcher's

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Fig. 6. (a) Wind data for Ras-Al Khaimah, UAE (b) mean temperature  (C) and relative humidity (%) [36,37].

Table 2 Boundary conditions for the CO2 analysis [33]. Cuboid volume Number of cuboids Cuboid mouth surface area Exhaled air flow rate Outdoor air CO2 concentration Exhaled air CO2 concentration

0.09 m3 12 (equally spaced) 0.00013 m2 6 L/min 382 ppm 36,000 ppm

ventilation rate. The system provided between 79 and 445 L/s (at 1e5 m/s outdoor wind). In contrast a similar size four sided windcatcher without heat pipes can provide between 135 and 722L/s [13]. Furthermore, the cross-sectional plane contours in Fig. 9a show that the airflow movement around the room is limited when occupants (represented by manikins) were located in the room, with low air velocity (<0.1 m/s) in the center of the room and high drafts (up to 0.50 m/s) in one corner of the ventilated space. The column of air moving down vertically from the windcatcher could potentially cause discomfort to the occupants located near the supply area, especially during high outdoor wind speeds. The volume control damper are designed to reduce or eliminate this issue. The subsequent increase in concentration of CO2 above the outdoor levels (250e350 ppm) can be utilised as an approximation of the sufficiency of indoor ventilation. Fig. 9b shows the crosssectional contours of CO2 concentration in the space with equally spaced 12 exhaling occupants. As the fresh air is supplied in to the space, it mixes with the stale air, diluting the CO2 concentration in the room. This lowers the mean CO2 concentration in the space below 1000 ppm as suggested by ASHRAE Standard 62. The results were consistent with the air velocity distribution in Fig. 9a with higher CO2 levels in low air speed zones. A comparison between the average supply air velocity and the

Fig. 7. (a) Wind tunnel setup (b) 3D printed windcatcher model.

concentration of CO2 in the test room as factors of the outdoor wind speed is shown in Fig. 10. As the incoming supply air is at a lower concentration than the air in the test room, this effect is expected as less polluted air is introduced. The graph shows that above an outdoor wind velocity of 1 m/s, resulting in an average supply velocity of 0.16 m/s, the recommended concentration of CO2 for good indoor air quality below 1000 ppm is achieved.

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Fig. 8. Experimental field test location in RAK.

The average supply air velocity increases linearly with the increasing outdoor wind velocity, given that the supply air velocity is directly proportional to the outdoor wind velocity, this is expected. The carbon dioxide concentration decreases in a logarithmic pattern as outdoor wind velocity. Given that sources of CO2 exist in the test room, the average concentration cannot decrease to the same value as the incoming concentration of 382 ppm and will be above this value.

Fig. 9. Cross-sectional contours of (a) air velocity and (b) carbon dioxide distribution in the modelled space with exhaling occupants, at UH ¼ 2.2 m/s.

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As previously shown in Fig. 6, the average outdoor wind velocity for Ras-Al- Khaimah is between UH ¼ 3.6e4.6 m/s throughout the year. This suggests that the average CO2 concentration will be lower than the recommended concentration a significant percentage of the year through the use of the passive cooling windcatcher. The inclusion of the cooling heat pipes do not reduce the supply air velocity significantly to prevent good indoor air quality. During very low wind speed days, integration of a solar powered fan could guarantee fresh air ventilation. Hughes and Ghani [38] investigated the potential of incorporating a solar powered fan into the channel of the wind catcher. The study showed that a low-induced pressure of 20 Pa is enough to satisfy the fresh air requirements in periods of low external wind velocities (1 m/s). Fig. 11 shows the influence of the addition of the extended surface on the average indoor velocity at various outdoor wind speeds. The comparison between two configurations; (1) windcatcher with heat pipes and (2) windcatcher with heat pipes and extended surfaces showed that the addition of extended surface reduced the average indoor velocity by 18e25%. The effect of the heat pipes on the incoming air temperature and relative humidity were simulated for each of the months using the average weather data shown in Fig. 6. Using this data for the input values at the boundary conditions, the effect of the heat pipes was determined. Fig. 12a shows the static air temperature contours when the outdoor air temperature was at 40
C, taken as the average temperature in July. July was used as the reference month for summer as the average temperature was one of the highest measured. The heat pipes had a clear effect on the incoming air temperature, reducing it to 25
C and the average indoor air temperature was 30e32
C showing up to 10
C decrease in air temperature. The results for the indoor air temperature in the winter month of January (inlet temperature set to 21
C) is shown in Fig. 12b. January was used as the reference month for winter as the average temperature was the lowest measured. The heat pipes are less effective as compared to summer due to the low temperature difference between supply air and the cold sink. The heat pipes operate most effectively when the temperature difference between the outdoor air and heat pipe fluid is greatest. A temperature drop of 1e2
C is observed in the winter months. Thermal comfort analysis should be carried out to assess the effect of other environmental and personal factors. The relative humidity content of air is dependent on the air temperature and in turn has a significant impact on the thermal comfort of occupants as sweat cannot evaporate from the skin's surface, preventing evaporative cooling. For thermal comfort, an accepted range of relative humidity is 40e70%. Though the mass fraction of water vapour in the air may not change, increases or

Fig. 10. Outdoor wind speed effect on CO2 and supply air velocity.

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decreases in temperature affect the relative humidity value. This effect can be seen clearly in Fig. 13a and to a lesser degree in Fig. 13b, which show contours for relative humidity in the test room for the summer and winter months respectively. The two figures directly relate to Fig. 12a and b which show the air temperature in the test room in the two different climate conditions. Fig. 13a shows the contours for relative humidity in the test room for a summer month with a low relative humidity of 50% at the inlet. As the air temperature decreases from the inlet to the space below the windcatcher, the relative humidity is shown to increase in the same regions. Though the mass fraction of water vapour in the air was not affected by the interaction with the heat pipes, the relative humidity increases as the air temperature drops as the air is less able to store a total volume of water vapour until the air is completely saturated and condensation forms. The average relative humidity in the test room in the summer months is approximately 65% based on the contour plots. This is within the range generally accepted for relative humidity with regards to occupant comfort and wellbeing. However, the relative humidity below the heat pipes reaches a maximum of 100%, meaning the air is fully saturated. This means that condensation will form, which may result in mould and bacteria growth if not dealt with appropriately. Fig. 13b shows a similar set of contours for the relative humidity during the winter months. Though the inlet air enters at a higher relative humidity value, the small reduction in air temperature results in an 8e10% increase. The average relative humidity is similar to that in the summer months, approximately 65e70%. This suggests that the relative humidity of the test room was within the acceptable range for the majority of the year given the two extremes of inlet values were assessed. However, it should be noted that the data is based on mean outdoor humidity levels and there could be the possibility of condensation formation during times of the day when humidity is at higher levels. During these conditions, the windcatcher should be closed off to prevent condensation or a suitable dehumidification system can be incorporated. It should be noted that existing dehumidification technologies would not be suitable due to significant pressure drop. The pressure drop would require additional fans to be used to supply the necessary flow rate for adequate ventilation. The incurred energy costs may offset the energy saved by using the windcatcher and low-energy cooling system. However [29], provides an example of a new design of dehumidification technology that features lower pressure drop across the device. This suggests that the device could be integrated into passive ventilation technologies as no fans would be required to maintain the ventilation supply rates. Integration of this technology would be beneficial by providing a method to continuously operate

Fig. 11. Impact of extended surface on average indoor speed.

Fig. 12. Contours of air temperature in the test room during the (a) summer month (July) and (b) winter month (Dec).

the windcatcher with low energy cooling without the concern of condensate build up. The values of average indoor air temperature and average relative humidity in the test room, based on the weather data, over the course of the year are shown in Fig. 14. The average indoor air temperature measurements follow the expected trend, increasing from a minimum in January to a maximum in the summer months of July and August before falling again as the seasons move back to winter. This mirrors the outdoor air temperatures shown. The average relative humidity measurements in the test room do not follow a clear trend as the air temperature measurements. Though the lowest outdoor relative humidity was measured in May, as seen in Fig. 6 and corresponds to the lowest value in Fig. 14, the indoor highest relative humidity measurement was recorded in September despite this not being the highest outdoor relative humidity measurement. The maximum and minimum air temperatures measured in the indoor test room were 31
C and 21
C respectively. The range of relative humidity was 48e75%. The relative humidity was within the acceptable range throughout the year, however it is likely that the maximum indoor air temperature of 31
C will cause discomfort to occupants despite the lower relative humidity. The thermal comfort levels were calculated using the ASHRAE PMV method (Standard 55e2017), expressed by the thermal sensation scale ranging between þ3 hot to 0 neutral and to 3 cold.

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753

Fig. 14. Annual indoor temperature and humidity.

performance was not assessed in this study but is investigated in previous works by Hughes and Ghani [39] and Elmualim [40]. Control of this system would be managed either by automated actuated dampers responding to sensor data collecting internal air temperature, or by a manual system of operation by occupants. This would improve the PMV scores and so the thermal comfort in the occupied space. Therefore, a control strategy must be developed and implemented to optimise the comfort levels in all areas.

Fig. 13. Contours of relative humidity in the test room during (a) summer month (Jul) and (b) winter month (Dec).

In this work, the numerical results for air velocity, temperature and relative humidity with defined clothing levels and metabolic rate were utilised to determine the comfort indices. It should be noted that we did not intend to accurately model or predict the thermal comfort levels in the space but rather use the tool to assess the comfort distribution in the space ventilated with the passive cooling windcatcher. Fig. 15a displays the horizontal cross section contours of PMV in the room with a passive cooling windcatcher during a typical summer month. As observed for the left corner area below the windcatcher, the PMV values were in the slightly-cool 0.96 to neutral range 0.12 while the other areas were in the slightly warm range 0.36e0.60. The combination of high air movement, colder temperature and high humidity below the windcatcher resulted in discomfort in this area (42% PPD avg.) while the other zones were in thermal comfort zone (6.7% PPD avg.). The windcatcher with low energy cooling technology was tested fully open to the external environment, with little control method for flow rate into the building below. A control strategy in the form of dampers below the heat pipes would enable the flowrate of supply to be monitored and altered as needs required. Changing the angle of the dampers between fully open, 0
, and fully closed, 90
, would affect the cooling rate of the incoming air, along with the ventilation rate. The impact of the dampers on windcatcher

Fig. 15. Contours of Predicted Mean Vote (PMV) thermal comfort distribution in the room during the (a) summer (Sep) and (b) winter (Dec).

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Fig. 17. Measurements temperature for outdoor air, supply air and heat pipe surface during a 5-h measurement window.

Fig. 16. (a) Smoke flow visualisation in the space ventilated by a windcatcher with heat pipes and extended surface (b) volumetric airflow through the wind tower supply channel at various wind direction.

Fig. 15b displays the PMV contours during a typical winter month (December) which shows a similar trend for the supply area having a different level of comfort than the other locations. Fig. 16a displays the visualised flow distribution inside the space ventilated by the windcatcher with heat pipes and extended surfaces. Fig. 16b shows the results of the airflow rate through the unidirectional windcatcher channel at various wind angles. In this figure, the flow direction in the supply channel are recognised by positive and negative values of airflow correspondingly. An airflow rate of 0.53 m3/s was achieved through the supply channel at 0
wind angle at an external wind speed of 3 m/s. At wind angle of 90
, the channel functions as an exhaust and the flow rate was not affected by the addition of the heat pipes and extended surfaces. Fig. 17a displays the results of the field tests for the supply air temperature at 3 positions (equally spaced points), heat pipe surface temperature and outdoor air temperature. The test was started at 11:00 when the outdoor wind was from the W direction and the windcatcher was functioning as an exhaust during this period. From 11:30 onwards the windcatcher started to supply airflow into the space and the temperature reduction ranged between 3 and 4
C. From 13:00 to 16:00, the wind began to blow consistently within the ±40
angle and the temperature drop ranged between 3 and 11.5
C during this period. Fig. 17b shows a zoomed in view of the measurements of temperature from 15:00 to 16:00 compared with the numerical

modelling results at various periods. The numerical results showed that in most cases it under-predicted the supply air temperature, however a comparable trend between both methods was observed. The average error between the numerical and experimental results was 3.15%. This is deemed an acceptable level of error between the measurements of the far-field testing and CFD simulations given the accuracy of the air temperature measurement device. 5. Conclusions and future works The thermal comfort and indoor air quality of a test room passively ventilated by a windcatcher integrated with heat pipes enhanced with extended surface was evaluated through numerical CFD simulations, wind tunnel and far-field testing in the UAE during a summer month. Results of the scaled wind tunnel tests showed that the addition of the heat pipes and extended surfaces reduced the airflow through the windcatcher but did not impede the flow even at low outdoor wind speeds, this was further confirmed by the smoke visualisation tests. Field tests data was used to validate the numerical modelling, showing good agreement between both methods. The investigation showed that the thermal performance of the heat pipe integrated windcatcher system can be further enhanced by using extended surfaces but presents additional pressure drop to the airflow. The simulation results showed that air temperature reduction up to 10
C was achieved using the windcatcher and heat pipe arrangement during summer months. The thermal comfort analysis was conducted using the ASHRAE PMV Standard 55e2017 method. It should be noted that we did not intend to accurately model or predict the thermal comfort levels in the space but rather use the tool to assess the comfort distribution

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in the space ventilated with the passive cooling windcatcher. This helped highlight the importance of other factors (humidity, air speed, metabolic rate, clothing level) when assessing the performance of the windcatcher system i.e. instead of just focusing on temperature reduction. It was observed that for the present design, equal distribution of thermal comfort was not achieved due to combination of high air movement, colder temperature and high humidity below the windcatcher which resulted in higher thermal discomfort in this area. Further work is required to develop a suitable control strategy in the form of dampers which would enable the supply flowrate to be monitored and altered as required and optimized the distribution in the occupied space. The CO2 concentration analysis showed that the system was capable of delivering fresh air inside the space and lowering the CO2 levels. The results showed that above an outdoor wind velocity of 1 m/s, the recommended concentration of CO2 for good indoor air quality below 1000 ppm was achieved. Although the proposed windcatcher can provide good indoor air and thermal comfort by reducing the incoming air temperature, more testing and development is required. Although the present windcatcher design is capable of reducing the supply air temperature more work is necessary to provide satisfactory indoor conditions especially during extreme summer conditions. In addition, the reduction in air temperature causes high relative humidity which may lead to and condensation formation and bacterial and mould growth, dehumidification technologies are required to address this. Further analysis of the spatial distribution of temperature and carbon dioxide in a room with different geometry and room furniture should also be performed. The effect of internal heat gains should also be taken into account in future works. Declarations of interest None. Acknowledgement The support by the University of Sheffield is gratefully acknowledged. The statements made herein are solely the responsibility of the authors. The technology presented here is subject to a UK patent (WO2015/087305). References [1] WBCSD, Transforming the Market: Energy Efficiency in Buildings, World Business Council for Sustainable Development, Geneva, 2009. [2] OECD/IEA, World energy outlook model [online] Available at: https://www. iea.org/weo/weomodel/, 2018. (Accessed 19 November 2018). [3] OECD/IEA, World energy outlook 2017, Exec. Summ. (2017) [online] Available at: https://www.iea.org/Textbase/npsum/weo2017SUM.pdf. [4] OECD/IEA, World energy outlook 2016. Chapter 1: introduction and scope [online] Available at: https://www.iea.org/media/publications/weo/ WEO2016Chapter1.pdf, 2016. [5] P. Sofotasiou, B.R. Hughes, J.K. Calautit, Qatar 2022: facing the FIFA World Cup climatic and legacy challenges, Sustain. Cities Soc. 14 (2015) 16e30. [6] A. Afshari, L.A. Friedrich, A proposal to introduce tradable energy savings certificates in the emirate of Abu Dhabi, Renew. Sustain. Energy Rev. 55 (2016) 1342e1351. [7] Y.H. Yau, S.K. Lee, Feasibility study of an ice slurry-cooling coil for HVAC and R systems in a tropical building, Appl. Energy 87 (2010) 2699e2711. [8] B.R. Hughes, H.N. Chaudhry, J.K. Calautit, Passive energy recovery from natural ventilation air streams, Appl. Energy 113 (2014) 127e140. [9] C.W. Chen, C.W. Lee, Y.W. Lin, Air conditioning - optimizing performance by reducing energy consumption, Energy Environ. 25 (2014) 1019e1024. [10] J.K. Calautit, B.R. Hughes, D. O'Connor, S.S. Shahzad, Numerical and experimental analysis of a multi-directional wind tower integrated with verticallyarranged heat transfer devices (VHTD), Appl. Energy 2815 (2017) 1120e1135. [11] B.R. Hughes, J.K. Calautit, S.A. Ghani, The development of commercial wind towers for natural ventilation: a review, Appl. Energy 92 (2012) 606e627. [12] S.H. Hosseini, E. Shokry, A.J. Ahmadian Hosseini, G. Ahmadi, J.K. Calautit, Energy for Sustainable Development, vol 35, 2016, pp. 7e24. https://doi.org/10.

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Nomenclature U: Velocity magnitude (m/s)

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T: Air temperature  (C) X, Y, Z: Cartesian co-ordinates (m) r: Air density (kg/m3) m: Kinematic viscosity (m2/s) g: Gravitational acceleration (m/s2) A: Cross-sectional area (m2) P: Pressure (Pa) L: Length (m) W: Width (m) H: Height (m) t: Time e: Specific internal energy (J/kg) keff: Effective heat conductivity (W/mK) hi: Specific enthalpy of fluid ji: Mass flux (kg s1 m2) Gb : Source of turbulent kinetic energy due to buoyancy force ak : Turbulent Prandtl numbers k: Turbulence kinetic energy (m2/s2)

ε: Turbulence dissipation rate (m2/s3)

a: Power law coefficient

ABL: Atmospheric Boundary Layer CFD: Computational Fluid Dynamics CIBSE: Chartered Institution of Building Services Engineers FVM: Finite Volume Method HVAC: Heating, Ventilation and Air-conditioning HTD: Heat Transfer Device PMV: Predicted Mean Vote PPD: Predicted Percentage Dissatisfied RAK: Ras-Al-Khaimah RANS: Reynolds averaged NaviereStokes RH: Relative Humidity SIMPLE: Semi-Implicit method for Pressure-linked Equations TKE: Turbulence Kinetic Energy UAE: United Arab Emirates UK: United Kingdom