Research Article

Indoor/Outdoor A

irflow

and Air Q

uality

E-mail: will.turner@tcd.ie

Residential hybrid ventilation: Airflow and heat transfer optimisation of a convector using computational fluid dynamics

William J. N. Turner1,2 (), Hazim B. Awbi1

1. The University of Reading, Whitenights Campus, Reading, Berkshire, UK 2. Trinity College Dublin, Museum Building, College Green, Dublin 2, Ireland Abstract Hybrid ventilation systems suitable for residential applications are being developed to reduce the energy demand of the housing sector. This paper describes the development and validation of a computational fluid dynamics (CFD) model of a convector unit that is a component of an existing residential hybrid system. The system incorporates a wall-mounted convector unit that controls ventilation airflow rate and air temperature. Airflow is provided by natural driving forces; a mechanical exhaust fan is used at times of low natural driving forces. The CFD model was used to study the aerodynamics and heat transfer processes of the convector unit with the aim of optimising system performance. Based on the modelling results, alterations to the geometry of a set of louvre blades inside the convector unit are suggested. The new louvre geometry prevents the formation of an airflow separation zone inside the convector unit. This improvement reduces the energy requirements of the system by reducing the convector air resistance by 20% and by increasing the thermal effectiveness of its heat exchanger.

Keywords hybrid ventilation,

residential,

aerodynamics,

energy,

computational fluid dynamics (CFD) Article History Received: 9 April 2014

Revised: 30 June 2014

Accepted: 7 July 2014 © Tsinghua University Press and

Springer-Verlag Berlin Heidelberg

2014

1 Introduction

Over 25% of EU final energy consumption is attributable to residential buildings (European Commission 2014b). This energy use is dominated by heating and cooling indoor air, of which between one-quarter and one-third is attributable to ventilation (Sherman and Matson 1997). Therefore, minimising the energy use of residential ventilation tech-nologies is a key measure to reduce CO2 emissions and meet EU energy targets for 2030 (European Commission 2014a).

The primary function of ventilation is to provide fresh air to ensure a healthy indoor environment for building occupants. It is important that indoor air quality (IAQ) is not sacrificed in the interest of reducing ventilation energy use. With estimates suggesting that people spend between 80% and 90% of their time indoors, either at home, working or commuting (ASHRAE 2004; Bower 1995; Sundell et al. 2011; Szalai 1972), adequate ventilation is necessary to prevent incidences of ailments related to poor IAQ such as Sick Building Syndrome (Seppänen and Fisk 2002).

Ventilation approaches traditionally fall into two categories: mechanical ventilation and natural ventilation. Mechanical ventilation relies on the use of electrically- powered fans and blowers to provide fresh air to the indoor environment. Air temperature and moisture concentration inside mechanically-ventilated buildings are usually controlled with air conditioning. Natural ventilation takes advantage of natural forces such as wind and buoyancy-driven pressure differences to move the air. Buildings that have natural ventilation systems generally tend to avoid the use of air conditioning and instead utilise cool night-time air to cool the thermal mass of the building. Natural ventilation systems have very low or zero energy demand, but are prone to under- and over-ventilation due to the variable nature of the natural driving forces, and occupants can expect to experience a wider range of thermal comfort conditions. Under-ventilation has adverse health implications for the building occupants, and over-ventilation increases heating and cooling energy use and can create draughts. Mechanical ventilation systems offer consistent and controllable

BUILD SIMUL (2015) 8: 65–72 DOI 10.1007/s12273-014-0192-5

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ventilation, but have an associated energy penalty from using powered mechanical components and relying on air conditioning.

A third ventilation approach has recently been the subject of research—hybrid ventilation. Also known as mixed-mode ventilation, this new approach combines both mechanical and natural ventilation processes to minimise ventilation- related energy use while still providing good IAQ (Heiselberg et al. 2001; Van Heemst 2001). Hybrid ventilation systems have both mechanical and natural components that may be used in conjunction with each other or separately at different times of the day as required. The natural component reduces the overall energy demand of the system, while the mechanical component ensures that the building is not over- or under-ventilated. Hybrid ventilation is intended to be a low-energy alternative to energy-intensive, mechanically- ventilated and air-conditioned buildings (Emmerich 2006) or naturally-ventilated buildings with low levels of control and potentially inferior occupant comfort.

Hybrid ventilation applications have recently become more prevalent in non-residential buildings e.g., Brohus et al. (2003). In 2001, a study by Heiselberg et al. (2001) reported on 22 commercial buildings with hybrid ventilation systems. At the University of California, Berkeley, Brager and Lehrer (2013) oversee the Mixed Mode project—a database of commercial buildings that use hybrid ventilation. At the time of writing the database contains information on over 150 buildings. While experiments have shown that hybrid ventilation is a feasible approach to ventilate residential buildings (Jreijiry et al. 2007; Kim and Hwang 2009; Niachou et al. 2008; Turner and Walker 2013; Yoshino et al. 2003), there are still very few successful implementations when compared with commercial buildings.

This study attempts to help redress the lack of studies on residential hybrid ventilation, by examining a residential hybrid system using numerical techniques. A computational fluid dynamics (CFD) model was developed to analyse airflow and heat transfer mechanisms inside a convector unit—the main component of a residential hybrid ventilation system—with the aim of improving overall system performance.

2 Research objectives

The subject of this study is a hybrid system designed by Lennart Wetterstad of Wetterstad Consulting AB, Löddeköpinge (formerly Minergi AB). It has been in development for several years, and the subject of academic research (see Elmualim et al. (2003), Wetterstad (2000), and Op’t Veld (2004)). The system uses a wall-mounted convector unit to temper and filter outside ventilation air. Usually, air is driven through the convector unit by natural

pressure differences. When these pressure differences are too low to provide natural ventilation, the system uses a mechanical exhaust fan to depressurise the room and promote airflow through the convector unit. Ventilation systems that rely on natural pressure differences to provide ventilation air require very low air resistances to operate effectively. Improving the aerodynamics of the convector unit to decrease its air resistance is an important part of improving the energy efficiency of the system. Decreasing the air resistance of the unit decreases the amount of time that the system will be reliant on the mechanical fan to provide adequate ventilation.

This study has three objectives: 1. analyse the internal aerodynamics of the convector

component of the system, 2. analyse the internal heat transfer mechanisms within the

convector, and 3. offer recommendations to optimise or improve the system

performance. These objectives will be addressed by developing a com-putation fluid dynamics (CFD) model to study the hybrid ventilation system that is described in the next section.

3 The hybrid ventilation system

The system comprises of a wall-mounted convector unit, an extract fan located within the installation room (Fig. 1), and a hot water source. Airflow through the convector system is driven by pressure differences between the indoor and outdoor environments. These pressure differences occur naturally due to wind pressure against the unit, internal and external air temperature differences, buoyancy within the room and also mechanically due to the extract fan. The extract fan acts to depressurise the room, drawing outside air in through the convector. When the natural pressure differences are too low the ventilation rate may be increased using the mechanical fan, hence hybrid ventilation.

A transparent cross-section of the convector unit is shown in Fig. 2. Outdoor air enters the unit at the inlet. A

Fig. 1 The hybrid ventilation system (IKM 2008)

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Fig. 2 Cross-section of the wall-mounted convector unit

louvre (a set of angled blades fixed at regular intervals to allow air to pass through) prevents the ingress of leaves and small creatures. An aerodynamic damper regulates the airflow through the unit and eliminates drafts due to sudden gusts of wind. A water-to-air heat exchanger inside the convector unit is plumbed into the hot water central heating system of the building. Hot water flows under pressure through a coil inside the heat exchanger. As ventilation air flows over the heat exchanger, the air is warmed. The tempered air then passes through a filter before it leaves the unit and enters the room. The ventilation air mixes with the internal air in the room via convection processes. Stale indoor air is exhausted through the extract fan while it is operating, or through alternative airflow paths such as cracks in the building envelope, chimneys, passive stacks etc.

4 Methodology

A computational fluid dynamics (CFD) model was developed to investigate the internal aerodynamics and heat transfer mechanisms of the convector using ANSYS CFX v.11 (ANSYS 2008). Below will be described the procedure through which the model was developed and validated, including experiments conducted to characterise the convector so that it could be accurately simulated. The CFD model was then used to perform simulations to address the three research objectives outlined above.

4.1 Model development

Three properties of the convector needed to be determined experimentally and then included in the CFD model empirically: 1. the air pressure drop across the heat exchanger and the

air filter as a function of airflow rate, 2. the heat transfer rate in the water-air heat exchanger, 3. the angle of the aerodynamic damper blades as a function

of the airflow rate. Due to the intricate and complex geometry of the water–air heat exchanger and the filter, they were modelled as porous

materials using an isotropic loss model based on Darcy’s Law (Darcy 1856). Experiments were conducted in order to ascertain the correct pressure–airflow relationships for the two components. Pressure taps were drilled into the top of the convector either side of the heat exchanger (see insert of Fig. 3). A manometer (accuracy ± 0.1 Pa) was used to measure the pressure drop across the heat exchanger, across the filter and across the heat exchanger and filter combined, for varying airflow rates. The pressure–flow relationships were then plotted (Fig. 3) and expressions were obtained for the pressure drop across the heat exchanger HeatEx )(ΔP and air filter Filter )(ΔP [Pa] as functions of airflow rate (QA) [L/s]:

1.5664HeatEx AΔ 0.0076P Q= ⋅ (1)

1.1997Filter AΔ 0.0302P Q= ⋅ (2)

As the geometry of the heat exchanger was not modelled explicitly, the water-to-air heat transfer relationship was determined experimentally. Experiments were carried out to determine the air temperature difference (TA) before and after the heat exchanger for different water temperature differences (TW) into and out of the heat exchanger coil (Fig. 4). Water temperatures were measured using thermally- insulated Type-K thermocouples (± 0.25 K) attached to the supply and return pipes of the heat exchanger coil. Air temperatures were measured using the same Type-K ther-mocouples in multiple locations before and after the heat exchanger air inlet and outlet so that a face-averaged air temperature difference could be obtained. Water flow rate through the heat exchanger coil (QW) was maintained at

Fig. 3 Pressure differences relative to the internal ambient pressure were obtained by measuring the pressure drops (P) across the heat exchanger, air filter and the combined heat exchanger and filter, for different airflow rates (QA) through the convector unit. The locations of the pressure taps are shown (insert). Room pressure was also measured at the convector outlet to obtain the pressure drop across the filter, and the filter and the heat exchanger combined. Error bars were calculated using standard error propagation methods (see Turner (2009))

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Fig. 4 Water temperature difference (TW) required to produce varying air temperature differences (TA) in the convector unit heat exchanger

4.0 L/min. The energy given to the air and the power of the heat exchanger (P) can then be calculated using:

= p AΔP mC T (3)

where m is the mass flow rate of air ( Am ρQ= ) [kg/s] over the heat exchanger, Cp is the specific heat capacity of air [kJ/(kg·K)], and TA is the change in air temperature [K].

The aerodynamic damper inside the convector unit consists of an array of six pivoting, light-weight, self- regulating damper blades (see Fig. 2, Fig. 5 and Fig. 6). At zero airflow rate the damper blades are in the neutral position (horizontal). As the airflow rate through the convector unit increases, so does the angle of the damper blades and the airflow resistance of the damper, thus regulating the ventilation rate. The dampers will fully close (to the vertical position) and cut off airflow through the convector unit entirely when the pressure difference across the convector unit exceeds 25 Pa. The 25 Pa pressure difference can be caused by either a steady airflow rate exceeding 140 L/s, or a sudden gust of wind. The angle of the damper blades relative to their neutral position was measured for different airflow rates using a small charged-couple device (CCD) camera installed inside the convector with a low-energy LED light. Figure 5 shows the damper angle-airflow rate relationship, while the insert in Fig. 5 shows the view from the CCD camera inside the convector. At times the damper blades would behave erratically, flipping to unanticipated positions. Turbulent airflow through the damper, coupled with the light weight of the damper blades is the most likely cause for this behaviour. It is plausible that there could be more than one damper angle and pressure drop for any given airflow rate, but Fig. 5 shows the best steady-state positions of the dampers that the authors could achieve.

Mesh-independency tests were performed using three separate meshes with default body spacings of 5, 10 and

Fig. 5 Angle of the aerodynamic damper blades relative to horizontal (0°) as a function of airflow rate (QA) through the convector unit. The damper blades would suddenly flip to the fully closed position (90°) when the pressure difference across the convector unit exceeded 25 Pa, caused by an airflow rate exceeding 140 L/s or a sudden gust of wind

Fig. 6 CFD mesh detail for the main body of the convector and heat exchanger (a) and the air filter (b). Note the inflation layers around the louvre and the damper blades, and from the air– material and material–material interfaces of the convector, heat exchanger and filter

20 mm, producing meshes with 1.5 × 106 (fine), 0.7 × 106 (medium) and 0.5 × 106 (coarse) elements respectively. Each mesh was used with three different turbulence models: κ-epsilon, RNG κ-epsilon and SSG (Speziale-Sarkar-Gatski) (Jones and Launder 1972; Speziale et al. 1991; Yakhot and Orszag 1986). Results showed that the coarse mesh with the RNG κ-epsilon turbulence model produced simulation results within the limits of experimental accuracy and in the shortest time, so the combination was used for the simulations. For more detailed results on the different meshes and turbulence models see Turner (2009). Figure 6 shows detail of the CFD mesh used—note that the line of symmetry along the length of the convector unit was used to simplify the CFD model by only simulating half of the convector, and then mirroring the results in the line of symmetry. Inflation layers were used to increase the mesh density at the air–material and material–material interfaces. The CFD domain was bounded by the internal geometry of the convector unit. The upstream end at the convector inlet was extended by 150 mm to prevent a simulated pressure

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build-up from forming at the top louvre blade. The pressure build-up was a modelling artefact of the way CFX handles airflow into a fluid domain and not representative of the physical airflow.

4.2 Model validation

In order to validate the CFD model, two experiments were conducted so that their results could be used for comparison with results from the CFD model. The experiments measured the face-averaged ventilation air velocity from the convector outlet, and the warm ventilation air jet path from the convector outlet.

During these time-averaged (5 minutes) experiments, air velocity measurements were taken using two hot wire anemometers (± 0.1 m/s) at 16 locations evenly-distributed over the face of the convector outlet (the filter) for a range of airflow rates. The time-averaged air velocities were then averaged over the face of the convector and compared with CFD results (Fig. 7). The results show good agreement between experiment and simulation.

A thermal imaging camera was used to capture the warm air jet emanating from the convector outlet for airflow rates of 40 and 80 L/s and air outlet temperature of 26℃. Room ambient air temperature was measured at 21.0℃ ± 0.5℃. A large, thin cardboard sheet (black in colour) was placed vertically inside the air jet and perpendicular to the outlet so that the air would warm the sheet to a temperature close to that of the jet. The same conditions were then recreated using the CFD model. For these simulations only, the CFD domain was extended at the convector outlet to include a cuboid room with dimensions 6.0 m × 6.0 m × 2.4 m. Figure 8 shows that the CFD results give a good representation of the size, angle and momentum of the experimental jets.

Fig. 7 Comparison between experimental face-averaged outlet air velocities (v) for a range of volumetric airflow rates (QA) and CFD predictions using the RNG κ-epsilon turbulence model with the coarse mesh

Fig. 8 Results showing the warm air ventilation jets. Experimental results for airflow rates of 40 L/s (a) and 80 L/s (b). CFD results for 40 L/s (c) and 80 L/s (d)

4.3 Simulations

Simulations were performed using the CFD model to analyse the internal aerodynamics and heat transfer mechanisms of the convector unit. An airflow rate of 40 L/s was used; a suitable ventilation rate for residential applications based on the 2006 UK building regulations (GB Office of the Deputy Prime Minister 2006). Since this work was undertaken, these have now been updated with the publication of the 2010 regulations. An output power of 0.5 kW was used for the heat exchanger, uniformly distributed throughout its volume. Buoyancy and gravity (9.81 m/s2) were switched on for all simulations. An initial outlet static pressure of 0 Pa was used with a reference pressure of 1.01325 × 105 Pa (1 atm). The CFD routine would then calculate the actual pressures inside the convector based on iterative calculations. Air was modelled as a constant property gas due to low simulated air velocities and density variations driven by small temperature differences.

5 Results and discussion

Initial simulation results showed the presence of a large separation zone inside the main body of the convector due to the geometry of the louvre (Fig. 9). Subsequent simulations were then conducted with different louvre geometries and configurations to ascertain the best way to alleviate the separation zone and its effects.

5.1 Original louvre configuration

Streamlines showing the paths of individual air particles through the convector from inlet to outlet were plotted in the

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Fig. 9 Streamlines through the centre of the convector along the z-axis. Note the separation zone in the low-y half of the main body, stemming from the lowermost louvre blade. Note: an airflow rate of 40 L/s is used for all of the CFD results shown

yz-plane to show the airflow pattern through the convector (Fig. 9). Ventilation air comes in through the inlet at the low-z end of the convector, flows over the louvre, then the damper, crosses the main body of the convector before transiting the heat exchanger and then finally exiting through the outlet of the filter at the high z-end. Flow over the top three damper blades is fairly consistent with the large bulk of air flowing in the positive z-direction. Flow over the bottom three damper blades is greatly affected by the lowermost louvre blade splitting the airflow. Most of the air flows upwards but a small volume travels between the underside of the damper blade and the lower wall of the convector causing a separation zone (vortex) to form in the main body of the convector where the air re-circulates. The airflow in the lower half of the convector body reaches the heat exchanger at which point the boundary interface created by the resistance of the heat exchanger, coupled with the low pressure area in the separation zone causes the air to re-circulate.

Figure 10 shows the total pressure (static pressure plus flow kinetic pressure) inside the convector. The scale of –0.5 to 0.5 Pa was chosen to accentuate the high and low pressure regions (note: the actual pressures at the top of the range are higher than 0.5 Pa). Areas of higher pressure are closer to the inlet before the louvre, around the upper two damper blades, and at the high y-end of the heat exchanger/filter interface. Low pressure areas are beneath the high z-end of the louvre where the airflow is diverted upwards by the natural angle of the louvre blades, beneath the trailing tip of the top three damper blades, and at the filter outlet. The high pressure areas arise as a result of the air resistance of the louvre.

Figure 11 shows the heat transfer between the heat exchanger and the ventilation air. Cooler air is coloured blue and warmer air is coloured red. Air enters the convector inlet at 20℃, traverses the main body of the convector and

Fig. 10 Total pressure plot of the inside of the convector with reduced scale to highlight areas of high and low pressure

Fig. 11 Temperature plot showing heat transfer between the heat exchanger and the ventilation air inside the convector

then passes through the heat exchanger where it is warmed. The separation zone in the main body can be seen where the ventilation air is re-circulating inside the main body of the convector. Air that is warmed by the heat exchanger flows back into the main body of the convector in the negative z-direction rather than leaving the convector at the filter. The air leaving the convector is warmer at the bottom of the filter than at the top where the air velocity is greater. This temperature disparity is because the slower moving air is spending a longer time in thermal contact with the heat exchanger. The separation zone stemming from the lowermost louvre blade is causing non-uniform air velocities at the convector outlet, and also causing a non-uniform distribution of heat transfer from the water in the heat exchanger to the air.

5.2 New louvre configuration

In an attempt to prevent the formation of the separation zone inside the convector, nine different louvre configurations were tested to see which would alleviate the problem most effectively. The angle, shape and position of the louvre blades, and the position of the damper were all investigated. For brevity, the most effective configuration will be discussed

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here. For a complete description of the other configurations see Turner (2009). The same boundary conditions as described in Section 4.3 were used for the alternative louvre simulations.

Out of the alternative louvre configurations considered, smaller and more numerous, elliptical louvre blades proved most effective at preventing the formation of the separation zone. The number of blades was increased from seven to nine, their length was decreased from 50 mm to 40 mm, and their shape was made elliptical rather than flat. A streamline image of the airflow through the convector using the small, elliptical louvre blades shows the near disappearance of the separation zone (Fig. 12). There is a small zone of circulating air between the lowermost blade and the damper, then another thin zone extending from the damper. However, by the time the airflow has reached the interface between the main body and the heat exchanger the separation zone has almost completely diminished.

The smaller size and the elliptical shape of the louvre blades meant that the ventilation air was not diverted upwards to such a large extent as with the original louvre blades. The elliptical shape of the blades helped the air to flow smoothly over their surfaces. Use of the new blades meant that most of the surface area of the heat exchanger was exposed uniformly to incident air. A more uniform distribution of air and air velocities across the heat exchanger will improve its effectiveness. Figure 13 shows the filter outlet air velocity profiles for both the original louvre design and the small, elliptical lourvre design. The new louvre design has a smaller velocity gradient across the outlet of the filter.

Figure 14 shows the total pressure inside the convector with the new louvre configuration. The air resistance of the convector was seen to decrease. Results show that the total pressure difference between the inlet and the outlet for the original louvre was 1.14 Pa for QA= 40 L/s. For the small elliptical louvre blade configuration the total pressure difference decreased to 0.91 Pa, a difference of approximately 20%. Low air resistance is particularly important for

Fig. 12 Streamlines for the new louvre configuration, using nine smaller elliptical blades

Fig. 13 The filter outlet air velocity profile for the original louvre design (a) and the small, elliptical louvre design (b)

Fig. 14 Pressure plot of the inside of the convector with the new louvre configuration. The reduced scale is again used to highlight areas of high and low pressure

ventilation systems that rely on natural pressure differences. For this particular hybrid system, reducing its air resistance will extend the range of natural pressure differences that will be suitable for ventilation airflow, and reduce the time that the mechanical exhaust fan will be required, thereby saving energy.

Finally, a temperature plot for the convector with the new louvre configuration shows that the more uniform airflow through the convector causes a more uniform heat transfer from the heat exchanger to the ventilation air (Fig. 15). The new louvre configuration improves the effectiveness of the heat exchanger and consequently the total system performance.

Fig. 15 Temperature plot of the convector with the new louvre configuration

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6 Conclusions

A detailed CFD model was developed and validated to simulate the physical airflow and heat transfer processes of a wall-mounted convector unit. The convector unit forms part of a hybrid ventilation system suitable for residential applications. Simulations were performed using the CFD model, with the objective of analysing and optimising the aerodynamics and heat transfer mechanisms within the convector unit. Simulation results show that a separation zone forms in the airflow inside the convector unit, caused by the geometry of a set of louvre blades at the unit inlet. An alternative design for the louvre is suggested to prevent the formation of the separation zone. This improvement has two positive implications. First, it reduces the air resistance of the convector geometry by 20%, thus reducing the natural pressure differences required for airflow through the system and reducing the time that mechanical airflow would be required. Second, it improves the uniformity of airflow across the convector heat exchanger, resulting in the increased thermal effectiveness of the heat exchanger.

Acknowledgements

This work was supported by the UK Engineering and Physical Sciences Research Council (EPSRC) and Monodraught Ltd. The authors would like to thank Trevor Orpin and Dirk Keeley for their invaluable contribution towards the experimental components of this research.

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