Ventilation efficiency in an L-shaped room

Transcription

1 Eindhoven University of technology Department of the built environment Master thesis in Building physics and services Ventilation efficiency in an L-shaped room Author: Ing. H.A. Dijkstra Studentnr: Supervisors: Prof. Dr. Ir. J.L.M. Hensen Dr. Ir. M.G.L.C. Loomans Ir. B.E. Cremers (Zehnder) 18 June 2015

2 Index Abstract Introduction Method Performance indicators CFD model Boundary conditions Axial diffuser Radial diffuser Case studies Grid dependency Results Diffuser placement case Diffuser type case Heating and cooling cases Hot patch case Discussion Conclusion Acknowledgements References Appendix A measurements of velocity profiles Measurement protocol Results Smoke test Velocity measurements Validation Appendix B CFD Model Base case Geometry Boundary conditions Models and discretization Case studies Envelope Diffuser Heating/Cooling Hot Patch

3 The effects of geometry on ventilation efficiency Author: Ing. H.A. Dijkstra Eindhoven University of technology Abstract Attempts to decrease energy use in buildings have caused an increase in application of balanced ventilation systems with heat recovery and better insulation in buildings. Ventilation efficiency is often assumed to be adequate in residential buildings. While this is investigated and proven to be the case for simple geometries, more complex geometries could cause some problems. This study investigates the ventilation efficiency of balanced ventilation systems in an L-shaped geometry, where isothermal and non-isothermal cases are considered, as well as axial and radial diffusers. A case study is done using computational fluid dynamics (CFD). A model is built to investigate velocity, temperature and age of air in the cases. Boundary conditions are retrieved from a measurement study. Cases are investigated on air quality, heat removal and draft. The study concludes that also for L-shaped geometries ventilation efficiency is adequate in the investigated cases. However, radial diffusers perform better than axial diffusers in air refreshment, heat removal and draft complaints for these geometries. 1. Introduction In the development of new buildings, energy consumption is becoming an increasingly important subject. Ventilation is accountable for a large part of the heat losses in a building. Balanced ventilation systems were introduced to tackle this problem. Balanced ventilation systems make a building more energy efficient by reusing the heat from extracted air. Compared to other ventilation systems like natural ventilation and mechanical extraction of air, balanced ventilation provides the largest decrease in the energy use. (Hekmat, Feustel, & Modera, 1986) Not having enough fresh air in a building has negative effects on its residents. Relations between insufficient ventilation and negative health effects have been investigated and proven numerous times (Sundell et al., 2011). Balanced ventilation systems have been in the news for causing noise and an unhealthy indoor air quality. These problems can be accounted for by improper installation and improper use of the balanced ventilation systems (Verkade, Duim, Merckx, Meijer, & ten Bolscher, 2009). When balanced systems are properly installed residents are generally content with their ventilation system (de Jong, 2014). In new well insulated residential buildings with balanced ventilation systems the airflow could show almost isothermal behavior, in contrast with older buildings. Leaks and cold/warm surfaces in the building provide free convection in older building. Better insulation means less free convection. Airflow is more dependent on the momentum force generated by diffusers in the room. In winter situations fresh air will often have temperatures near the room air temperature, because little heating is needed in the well-insulated houses, approximating isothermal behaviour. Guidelines for the installation of ventilation systems in general and for balanced ventilation systems are available. The Dutch building code( Bouwbesluit 2012, 2011) requires a minimum installed amount of airflow in residential buildings, as well as a maximum air speed in the living zone. Also requirements on controllability of the system are given. The minimum required airflow should be at least a setting in a system, and the lowest setting should be at least 10% of the minimum setting. The NPR 1088 is a Dutch guideline for ventilation systems in residential buildings. Information is mainly provided on the outdoor parts of a ventilation system. On indoor air the guideline only provides advice on air speed (NPR, 2004). Internationally the ASHRAE guidelines provide more grip. In these guidelines procedures can be found for determination of the ventilation rate. Also a maximum indoor concentration level is specified for specific substances, which makes it more complete than the Dutch guidelines (ASHRAE 62.2 Ventilation and accaptable indoor air quality in lowrise residential buildings, 2013). Most ventilation guidelines give information on the amount of air to ventilate for a good indoor air quality. Little information is given on correct placement of the diffusers in the building. In the Dutch ISSO 62 publication information can be found on where to place the supply diffuser in relation to the walls in the room. Diffusers should be placed at least some distance from the wall. Exact distances can be 3

4 supplied by the manufacturer. Installation at a high point in the room is advised. Also diffusers shouldn t be installed too close together (ISSO, 2010). These guidelines don t give information on the most optimal place to install diffusers. In case of more unusual geometries placement of the diffusers can become more important. Interviews in practice indicate that there are still a lot of questions on how ventilation systems perform when the geometry of a room is different from rectangular. In the Dutch building code, the effectiveness of a ventilation system is only assessed by CO 2 levels, and velocity of the air in a room. These indicators do not describe the air distribution in the room. It could be that while the air flow in the room is sufficient, not every place in the room is ventilated well enough, especially with less free convection and infiltration in new buildings. Indicators like the local air change index (Mundt, Mathisen, Nielsen, & Moser, 2004) describe the air distribution in the room and are more suited to assess how well a balanced ventilation system performs with respect to the ventilation efficiency. The local air change index (LACI) is an index that compares the local mean age of air in the room with the nominal time constant. The nominal time constant is the average age of air at the extract and is by definition the inverse of the air change rate. More information about this indicator and other indicators can be found in the method section. In residential buildings mixing ventilation is applied most of the time. In perfect mixed situations the local mean age of air in the room is homogenously distributed in the room, and equal to the nominal time constant. Thus the LACI is 100% all over the room. For smaller, rectangular room shapes the ventilation efficiency is already investigated for balanced ventilation system and mixing ventilation. In a rectangular climate chamber a balanced ventilation system proved to be mixing well, with LACI values close to 100% in the center of the room (de Coo, 2011). Buoyancy driven forces have little effect on the LACI in the room. Diffuser type and placement does have a significant effect on the LACI in the room. Furthermore, in rectangular shaped bedrooms, balanced ventilation systems have been proven to be efficient for a specified occupancy level (van der Pluijm, 2010), when the Dutch building code is applied. Measurements have proven that when the required volume flow is applied, that the CO 2 levels in the room are below the required threshold for different room set ups, and that mixing is present in the room. Adjustments to the ventilation system were needed for these results, thus applying a balanced ventilation system does not guarantee that the room is ventilated efficiently. Although balanced systems are proven to be efficient in smaller, rectangular shaped rooms, it is unknown how the ventilated air behaves in more complex geometries. Guidelines advise the distance between diffusers and walls, as well as the placement of diffusers relative to the living area, but don t provide information on how to deal with room geometries like alcoves (ISSO, 2010). Questions from practice indicate it is unclear what the influence of room geometries on ventilation efficiency is. L-shaped geometries and alcoves could influence the airflow in the room, as the spatial layout of equipment in a room can also influence airflow patterns (Huang & Lin, 2014). Most research on balanced ventilation systems has been done using simple room shapes(al-sanea, Zedan, & Al-Harbi, 2012) (Hu, 2003) (Malmström & Ahlgren, 1982). In this study the performance of mixing balanced ventilation systems is investigated when these systems are applied in larger, more complex geometries. For this study a single L-shaped geometry is used. The investigation holds different diffusers placements and types, heating and cooling cases, and the effect of sunlight. The following research questions are answered. How well do balanced ventilation systems perform when they are applied in L-shaped rooms? What is the effect of diffuser placement on ventilation efficiency? What is the effect of diffuser type on ventilation efficiency? What is the effect of heating and cooling? a. In what manner is the mixing in the room affected? b. How well do the diffusers perform in removing heat? c. How well do the diffusers perform considering draft? How does sunlight influence the flow in the room? a. Is the ventilation efficiency influenced by the heat caused by the sun? b. Is the excess temperature removed efficiently? 4

5 2. Method To simulate the influence of geometry on ventilation efficiency for L-shaped rooms a method with four phases is used, as well as a validation of the model (figure 1). FIGURE 1: FOUR PHASE METHOD USED IN THIS PAPER First a model that simulates the airflow, temperature, and the mean age of the air in the room is created. This model is built upon measured data from a climate chamber study on ventilation efficiency (de Coo, 2011). The second phase is a study on the inlet boundary conditions. Radial and axial diffusers are used in this study. The diffusers both require other methods to be implemented correctly into the models. Measurements are done to find the inlet velocities of the diffusers. A validation phase is done next. The mean age of air model as simulated is validated to the measurement study done by de Coo. As the case studies contain cases with axial and radial diffusers, velocity profiles of these diffusers are obtained by means of a measurement study. In the case study models, these velocity profiles are used as inlet boundary conditions. Next to the velocity profiles, validation measurements are also performed near the diffusers. To validate the measured velocity profiles, the measurement room is built in a CFD model. Simulations of the airflow in the measurement room with the measured velocity profiles as inlet boundary conditions are performed. The measured validation velocities are compared with the corresponding velocities in the CFD model When both the CFD model and the inlet boundary conditions are validated, case studies investigating the influence of geometry can be performed. Five case studies are investigated. A base case is defined as an L-shaped room with core sided axial supply diffusers, extract diffusers in the tip of the L shape, and isothermal conditions. For the base case a grid dependency study is performed. The four other case studies vary in thermal conditions, diffuser type, and diffuser placement. The five case studies are compared on the key performance indicators mean age of air, local air change index, velocity, temperature, and air diffusion performance index. 2.1 Performance indicators The mean age of air concept is a useful tool to assess the ventilation efficiency of a system (Mundt et al., 2004). Air in the room can be contaminated by contaminants in the room from the moment it enters the room and can be considered more contaminated when it resides longer in the room. In the age of air concept, the age of air is counted from the moment it enters the room. The air in a point of the room is a mixture of air particles that have spent a different time in the room. The average residence time of an air parcel in a point of the room is the local mean age of air. In situations where the air is fully mixed in the room, the mean age of air is homogenous over the room. When a short circuit occurs between the supply diffuser and extract diffuser, the mean age of air is low in the short circuit zone and high in the other zones. In the exhaust the local mean age of air is by definition equal to the nominal time constant, and thus in a fully mixed room, the local mean age of air in the room is also equal to the nominal time constant. The nominal time constant is by definition the inverse of the ventilation rate, which is often used to quantify ventilation. The nominal time constant is the time it takes for a volume of air equal to the room volume to be supplied to the room. Theoretically the nominal time constant is the fastest time to refresh all the air in the room. The nominal time constant can be calculated with equation 1, in which is the volume of the room and / is the volume flow (Etheridge & Sandberg, 1996). / [1] In the mean age of air concept there are two theoretical cases. The first is a fully mixed room as described above. In fully mixed rooms the local mean age of air is the same in the whole room. The second case is the piston flow case. In piston flow, air moves through the room like a piston. At one side of the room, air enters the room. The local mean age of air is 0s for air parcels on this side of the room. At the other side of the room, air leaves the room. The local mean age of air is equal to the nominal time constant on this side. Because the air moves linearly through the room, the mean age of the air in the room 5

6 is equal to /2. Most residential rooms with balanced ventilation systems have a supply and an extract somewhere in the room and tend to have mixed flow ventilation. In rooms with displacement flow ventilation the mean age air at breathing level is between /2 and. In rooms with short circuits the mean age of air at breathing level can become larger then. The air change efficiency is expressed as. It is a measure for how quickly the air in the room is refreshed by assessing the mean age of air. Air change efficiency can be expressed as the ratio between the shortest possible room mean age of air /2 and the actual room mean age of air (equation 2). (Cao et al., 2014). / 2 100% [2] The local variant of air change efficiency is the local air change index. It is a measure for the conditions at a specific point P in the room. The local air change index can vary a lot for different positions in the room, and can be high for positions near the supply and low in insufficient ventilated areas. The local air change index is defined as the ratio between the nominal time constant and the local mean age of air at a point. Equation 3 describes the local air change index. In a situation with complete mixing the local mean age of air matches the nominal time constant in the whole room. This generates a local air change index of 100%. In a piston flow situation the local air change index changes from infinite at entry to 100% at exhaust. / 100% [3] The local ventilation effectiveness for heat removal can be used to assess how well a ventilation system can remove excess heat from the room (Awbi, 2003). Formula 4 describes this indicator. In this indicator the temperature difference between the supply and the extract is compared to the temperature difference between the supply and a point in the room. Levels above 100% indicate that the temperature at a point is lower than the extract temperature and sufficiently removed from the room. Levels below 100% indicate that the ventilation system can t remove all the heat from this point. The ventilation effectiveness for heat removal is only used in cooling situations. / 100% [4] The air diffusion performance index (ADPI) is a measure that indicates the thermal comfort by assessing the draft levels in the room (ASHRAE Handbook: Fundamentals, 2009). To assess the level of draft an empirical equation is used to calculate the draft temperature. The draft temperature is calculated out of the temperature at a point and the average room temperature, together with the air velocity at a point corrected with empirical factors. The ADPI is then calculated by checking which percentage of points in the room satisfies the draft comfort limit. The draft temperature is calculated with equation [5] In equation 5, is the temperature at point x. is the average temperature in the room and is the air velocity at point p. If the calculated draft temperature is the set point is reached. The ADPI can be calculated with equation 6. In this formula is the amount of points with a satisfactory draft temperature and N is the total amount of measurement points. / 100% [6] 2.2 CFD model The first step in the method is to create a model in CFD that can be used to simulate airflow, temperature and age of air in a room. The model is built upon measurement data from a study on the ventilation efficiency in a climate chamber with balanced ventilation (de Coo, 2011). FIGURE 2: SECTION OF THE TU/E CLIMATE CHAMBER WITH THE MEASUREMENT EQUIPMENT The measurements were performed in the TU/e climate chamber. A section of the climate chamber is shown in figure 2. The climate chamber has the dimensions 5400 x 3600 x 2700 mm. The room has 4 channels for air transport with dimensions 200 x 200 mm. The upper two channels are used for supply and exhaust air. On the channels boxes with the supply and extract diffusers are installed. For supply a round hole was used. A measurement on a steady state, isothermal situation was performed. This measurement is later recreated in the simulation. 6

7 A computational model of the climate chamber test setup with dimensions 5400 x 1800 x 2700mm is made in the commercial CFD package Fluent 15 (Ansys, 2015). Assuming symmetry in the room, only half of the room needs to be modelled (Casey & Wintergerste, 2000). The symmetry boundary condition is placed in the center of the room instead, crossing the supply and extract boundary over the center. A simplification is made by making the supply and extract boundaries square, and keeping the total area of the boundary the same. The inlet diffuser is modelled as a square surface with the same area as half the inlet diffuser, measuring 120 x 47.1mm. The outlet diffuser has the same dimensions. Modeling is performed with a RNG turbulence model, which performs best for indoor air simulations (Chen, 1995). Boundary layers are modelled with enhanced wall treatment. Age of air in the room is calculated by means of a scalar which defines the laminar diffusivity of age of air (Fluent inc., 2006a). The age of air equation is calculated separately from the airflow equation. A grid is built with the pre-processor Gambit Grid refinement is used near the inlet and outlet boundaries, as large gradients are expected there. Also grid refinement is used near the walls in the domain. A grid sensitivity analysis is performed by calculating three different meshes. The resolution of mesh sizes can be assessed with the dimensionless wall distance y+, a factor that makes the distance to the wall dimensionless by adjusting the distance with the friction velocity near the wall(fluent inc., 2006b). Near wall treatment of the models is dependent on the dimensionless wall distance. In enhanced wall treatment low Reynolds number models are used when the y+ value is near 1. A blending function between low Reynolds modelling and wall functions is used for dimensionless wall distances between 3 and 10. Standard wall functions are used for higher y+ values. For an isothermal model y+ values near 1 are not as important as in non-isothermal models. The flow in the boundary layer of an isothermal model contributes less to the total flow in the room compared to a non-isothermal model. In the coarse grid the smallest cell sizes were 40mm, resulting in a grid of hexahedral cells (figure 3). Calculations with the coarse grid result in a value of 28 in the area where the jet hits the ground. The medium grid has a minimum cell size of 20mm, resulting in a grid of cells. Calculations with the medium grid result in a value of 12.8 in the jet area. The fine grid has a minimum cell size of 15mm and a grid of cells. Calculations with the fine grid result in a value of 10.8 in the jet area. FIGURE 3: THE COARSE MESH OF THE CLIMATE CHAMBER. THE GREEN SURFACE IS THE SYMMETRY BOUNDARY. To solve the model a second order discretization scheme is used for pressure discretization and a second order upwind scheme is used for the momentum equation. A SIMLE algorithm (Semi- Implicit Method for Pressure Linked Equations) is used for the pressure-velocity coupling (Ambatipudi, n.d.). Convergence is checked at a point 1 meter below the diffuser. When velocity residuals are stable in this point, and below at least 10, the solution is converged. Fluent s standard convergence checks are also monitored. For transient simulations time steps are converged when the continuity residual is less than 10. Velocity magnitude is also checked on the points, but not always stable due to the transient nature of the flow pattern. The grid sensitivity analysis is performed on velocity as the velocity vectors have a large influence on convection, and thus on the age of air in the room. Velocities at the four measurement points are extracted from the data and compared to investigate how they change under different grid conditions (figure 4). The difference in velocity magnitude between the coarse grid and the medium grid is bigger than 5% on one of the measurement positions, rendering the coarse grid not fine enough for the study. FIGURE 4: THE RESULTS OF A GRID REFINEMENT STUDY ON THE CLIMATE CHAMBER 7

8 The difference between the medium and fine grids is smaller than 5% on all measurement positions, making the medium grid independent of the grid size. The largest deviations in velocity magnitude can be found in measurement position 4 where the velocities are the smallest. In the simulations the inlet boundary is modelled in best agreement with the measured results. No exact measurements were done for the supply diffuser, only a measurement 1.3 meter below the diffuser was performed. A mean velocity normal to the boundary of 0.9m/s is set at the inlet boundary condition. A turbulence intensity of 22% is set at the inlet boundary condition. This turbulence intensity is retrieved from a measurement 1.3 meter below the diffuser. A hydraulic diameter of 0.125m is set, matching the diameter of the inlet duct. The associated turbulent length scale is m, assuming fully developed turbulent pipe flow in the pipe connected to the diffuser. Turbulent energy k and dissipation are calculated by the software. Only turbulent flow is modelled so no energy equations are used. The outlet boundary condition is set as outflow and the wall are set with no slip conditions. The simulation results are compared to measurement data. A comparison is made for the airflow in the room, velocity, room air change efficiency, and local air change index. The volume flow of the inlet boundary condition is the same as the volume flow of the measured data. The chosen inlet velocity over the chosen supply inlet surface, equals the measured volume flow rate. A check on the volume flow over the outlet boundary condition of the model proves that both the model and the measurement have an airflow of 38 /. Because of the lack of available data on the diffuser it is almost impossible to match the velocity magnitude at the position below the diffuser for both simulation and measurements. In the simulation the velocity magnitude and the turbulence intensity are about half of the measured values at the measured position (Table 1). Velocity at 1.3m below diffuser Turbulence intensity Measurement 0.9 m/s 22.2% Simulation 0.41 m/s 11.2% TABLE 1: COMPARISON SIMULATION AND MEASUREMENT The local mean age of air was measured at the four measurement positions, as well as simulated in the CFD model. The local air change index at the four measured positions is higher in the simulations than the measured values (figure 5). Also the measured LACI values are different between the two measurement, ranging 7% in the worst case. The same trends can be seen for the measurements and simulations. Both simulations and measurements show that the airflow is near fully mixed conditions but has a larger local air change index near the inlet diffuser and high in the room. FIGURE 5: LACI IN (%) FOR THE FOUR MEASUREMENT POSITIONS IN THE CLIMATE CHAMBER The method for simulating mean age of air is capable of calculating the local air change index in the same order as the measurements provide. Missing boundary conditions make it impossible to do a more accurate validation study. The inlet boundary condition has been proven to be indispensable in this study. For the case studies a more accurate velocity profile for the inlet boundary condition is needed. 2.3 Boundary conditions The CFD model is capable of calculating velocity magnitudes and the age of air in the room. Compared to the measured data, the CFD model overestimates the local air change index in the room. However, no accurate data of the boundary conditions was available in the climate chamber measurement data. Velocities of the jet stream in the model are not sure, which makes the airflow in the room unknown. Improvements are made on the model by implementing more accurate boundary conditions for the supply diffuser. For implementation of axial and radial diffusers in the case studies both types of diffusers are measured. Boundary conditions are created from the measurement data. A validation study is performed on the created boundary conditions. Model building of complex supply diffusers can be performed using different methods. Different diffuser types require different methods to be measured and simulated. For axial round ceiling diffusers the box method is the best method (Srebric & Chen, 2000). In the box method (figure 6) an imaginary box is drawn over the diffuser. The airflow in the box is not simulated. Air flow over the 8

9 bottom box surface is measured by velocity measurements. Air flows back into the box over the side surfaces of the box. Airflow over the diffuser also is measured. The amount of air flowing back over the side surfaces is equal to the difference of airflow over the diffuser and airflow over the bottom box surface. In the CFD model the bottom box surface is implemented as inlet boundary condition. The side box surfaces are implemented as outflow boundary conditions with a specified mass flux. Measurements of the box are case specific. The bottom box surface should be as big as the perimeter of the diffuser jet. The height of the box is determined by the recirculation zone of the diffuser jet. In the model all air is flowing out of the bottom of the box, so the bottom box surface should be below the recirculation area. the flow pattern of the recirculation zone and the perimeter of the diffuser jet can be extracted from the smoke visualization (figure 7) (Andresova, 2014). Velocity vectors can also be extracted from the smoke visualization. Judging from the smoke visualization, the recirculation zone reaches to 250mm from the ceiling. The bottom box surface is placed at this distance from the ceiling. The jet is 440mm wide at this distance from the ceiling, thus the imaginary box has the dimensions 250 by 440 mm. Measurements are performed on a 5 by 5 grid on the bottom box surface. Because the flow is highly turbulent, velocity measurements are performed for 3 minutes, and averaged, on every grid position. Figure 8 shows the measured velocities on the 5x5 grid. In the appendix more information about the measurement techniques and data can be found. FIGURE 6: MASS FLOWS OVER THE BOUNDARIES OF THE BOX METHOD (SREBRIC & CHEN, 2000) Radial diffusers are best simulated with the momentum method, or when the effective area of the opening in a diffuser and the real area of the opening are the same, the direct description method. In the direct description method the measured velocities are simply set to the surface of the inlet boundary condition. This can only be done when the volume flow over the surface is equal to the measured velocity multiplied with the surface area of the diffuser Axial diffuser FIGURE 7: SMOKE VISUALISATION OF THE AXIAL DIFFUSER Smoke visualization is used to determine the flow patterns of the diffusers. In case of the axial diffuser FIGURE 8: MEASURED VELOCITIES ON A 5X5 GRID 250MM BELOW THE CEILING In the measured velocity data, higher velocities are measured on one side of the box area, just as the smoke visualization shows. In this area velocities between 0.15 and 0.3 m/s are measured. The circle pattern in the measured results matches the axial profile of the diffuser. The velocity vectors in the jet stream vary from 15 degree inward to 5 degree outward, according to the smoke visualization. The supply jet of the measurement study is not symmetrical. This is due to how the ducts are connected to the diffuser in the measurement set up, as the duct comes in at an angle towards the diffuser, while in most studies a long straight duct is connected on top of the diffuser. The way the diffusers are connected does have an influence on how the air is supplied to the room, and could also have influence on the ventilation efficiency. Connecting the ducts under an angle to the diffusers is more often done in buildings. To validate the velocity measurements and the implementation to the CFD model, a second set of 9

10 measurements is done, 30 cm below the first set. In the CFD model velocities on these positions should match the measured velocities. Multiple iterations on the exact jet angles are investigated in the validation study, as the exact jet angles are difficult to extract from the smoke visualization. Figure 9 shows a comparison between the measured validation data and the model results with supply conditions 15 degree inward on one side and 5 degree outward on the other side. These condition give the best results, although they do not entirely match. A sensitivity analysis of the jet angle on the LACI in the room indicates that even with two extreme cases, being 15 degree outward and 15 degree inward angles, the difference between the LACI is small (appendix A). validation study is performed on the measured velocities. A second set of measurements at 1 meter from the diffuser is done, on 1cm and 5cm from the ceiling, to validate the direct description method for this diffuser. For both ceiling distances the simulated data is in the range of the inaccuracy of the measured data (figure 11). FIGURE 9: COMPARED VELOCITIES ON A 5X5 GRID 550MM BELOW THE CEILING Radial diffuser Smoke visualization of the flow pattern of a radial diffuser shows the coanda effect created by the diffusers. Supplied air sticks to the ceiling and disperses over the ceiling surface. Fresh air is supplied at a normal angle to the supply surface. The jet sticks to the ceiling on the top side and has a 7 degree angle on the bottom side. At 31 cm from the supply surface the air jet does not diverge any more, these measurements can be found in the appendix. The smoke visualization makes clear that the supply angle of the diffuser can be modelled normal to the supply area in the model (figure 10). With no recirculation area, and the effective supply area matching the real supply area for this diffuser, the direct description method can be applied for the model. FIGURE 10: SMOKE VISUALIZATION OF THE RADIAL DIFFUSER Velocities are measured at the supply surface of at 8 places around the diffuser. An average velocity of 1.7 m/s is measured, with no large deviations. A FIGURE 11: VALIDATION OF THE RADIAL DIFFUSER MEASUREMENT 2.4 Case studies To investigate the effects of room geometry a single geometry is defined. L-shaped rooms are quite common as non-rectangular room shapes. In Holland a lot of typical Dutch terraced houses have L-shape living rooms. The average size of these rooms is 8.5 by 6 meters (Arts, 2011), which is used as size of the simulated model. The simulation is done using an unsteady RNG k- epsilon turbulence model (Chen, 1995) and enhanced wall treatment, with dimensionless wall distance y+ values in the domain as close to 1 as possible (Fluent inc., 2006b). In enhanced wall treatment a low Reynolds variant of the turbulence model is used when the mesh is sufficiently fine (y+=1), which is better at solving non-isothermal flows near walls than the enhanced wall functions used with high y+ values. In grid building mapped and unmapped meshes are used. Mapped meshes offer more control over how the domain is filled with volumes. Nodal points on the edges of the domain are defined. More nodal points can be applied in places with high gradients, to make finer meshes. The domain is filled with hexahedral volumes which match the already set nodal points. In unmapped meshes the domain is filled with computer generated tetrahedral volumes. A method which is faster than mapped meshes, but offer little control. More skewed volumes are present in the meshed domain. Some domains with circular shapes can only be filled with unmapped meshes. The finite volume method simulates better when hexahedral volumes are used (Blocken, 2011). Mapped meshes are used where possible. When unmapped meshes are necessary, the volume 10

11 meshed with this method is kept as small as possible. Five models are made for this investigation. A base case is made first. 4 variants are made of the base case to investigate the effects of the change in the variant (figure 12 and appendix B). The base case model (fig 12.A) has dimensions 8.5 x 6 x 2.8 meter, as are the dimensions of all case studies. The supply diffusers are in the core of the room, as advised by the manufacturer. Distance between the supply diffusers and between the walls is 1 meter. The base case is modelled with axial supply diffusers. The extract diffusers are in the tip of the L-shape. In reality the extracts would probably be positioned in the same place as the kitchen is probably there. The base case is an isothermal model. The mass and momentum equations are solved. In a post-processing step the age of air equation is also solved. The supply diffusers are moved to the envelope wall in the envelope side wall case study (fig 12.B). The distance between the supply diffusers and the envelope wall is 1 meter. This case study is used to show the influence of diffuser positions. In the radial case study the diffusers are placed on exactly the same position as the base case, but a radial diffuser is used instead of the axial diffuser (figure 12.C). This case study investigates the influence of using radial diffusers instead of axial diffusers. The ventilated heating/cooling case is used to investigate how the ventilation efficiency is in a room where the ventilation system is used to heat or cool the room (fig 12.D). In cooling mode supply air has a temperature of K and the walls are modelled with a temperature of K. In heating mode the walls have a temperature of K, the floor has a temperature of K and supply air has a temperature of K. These temperatures match the indoor temperature in the shoulder seasons, where often only little heating or cooling is needed. The hot patch case study investigates the effects of solar irradiation on the floor on ventilation efficiency (fig 12.E). A surface in the room, where in reality solar irradiation occurs, supplies a heat flux into the room. For this model non-isothermal boundary conditions are used. The supply diffusers supply the air with a temperature of K. The walls of the room are modelled as adiabatic walls. The hot patch measures 6.8 and emits a heat flux of 10 /. With these parameters the hot patch heats the room by 3K. A temperature difference of 1.83K is expected between the supply and exhaust diffusers. Simulation of models is performed in two steps. First an initial solution is calculated by a steady simulation. The second step is an unsteady simulation. A simulation converges better in unsteady conditions, and can t find a solution in solely steady conditions. After the initial steady simulation step, the unsteady simulation runs for another 30 minutes of simulation time. Although the nominal time constant of this model is 89.6 minutes, and the volume of air in the model is not refreshed yet, the initial steady simulation ensures that there is a consistent airflow in the model. Monitors are used to monitor temperature and velocity at measurement points in the room. After 30 minutes velocities at the measurement points fluctuate around a mean velocity for that point. The solution is an unsteady airflow in the room. Low velocity airflow near walls probably fluctuates more than high velocity airflow. In non-isothermal cases, buoyancy driven flow near walls will probably increase this effect. All simulations are performed with second order discretization schemes. In the non-isothermal simulations buoyancy is modelled using the Boussinesq approximation. The Boussinesq approximation assumes the temperature difference in a flow is little, and therefore the density varies little. However, buoyancy drives the motion of the flow. FIGURE 12: THE BASE CASE STUDY A, THE ENVELOPE CASE STUDY B, THE RADIAL DIFFUSER CASE STUDY C, THE HEATING/COOLING CASE STUDY D, AND THE HOT PATCH CASE STUDY E. 11

12 The Boussinesq approximation neglects the density difference everywhere, except for the buoyancy term. It is applicable only when the temperature difference is small. In case of air the temperature difference can be 28.6K (Gray & Giorgini, 1976). In the model the constant density of air is modelled as /, and the operating temperature is K. The variation of density is determined by a thermal expansion coefficient. For air this coefficient is /. In the simulations first the mass, momentum, and energy equations are solved to determine flow and temperatures in the model. Mean age of air is calculated afterwards in a separate simulation. Flows and temperatures are not modelled in the second simulation, so the flow field is not changed in the post processing simulation to obtain mean age of air. A A FIGURE 13: POSITION OF THE COMPARISON POINTS The 5 case studies are compared using the key performance indicators LACI, draft temperature, velocities and temperature removal. Per case study only the applicable performance indicators are used. In the geometry 13 points are set for the comparison. These points are in the middle of the room, near walls, at breathing level (1.7m), and at floor level (0.1m). The position of these points is shown in figure 13. At breathing level the local air change index indicates the air quality of the air people breathe in. At floor level the ADPI indicates the risk of draft. For each case study the applicable points are used in the comparison. 2.5 Grid dependency The grid dependency study is performed on both velocity magnitude and local air change index on 11 points in the base case. 4 grid sizes are studied. The smallest grid has 750,000 cells. Grid sizes scale up with a factor 2, to a grid size of 5.2 million cells in the extra fine grid. The y+ value of the grids can be found in table 2. Coarse Med Fine Extra y TABLE 2: Y+ VALUES FOR THE DIFFERENT GRID SIZES FIGURE 14: GRID DEPENDENCY STUDY ON VELOCITY MAGNITUDE FOR THE BASE CASE In figure 14 the velocity magnitude at 11 measurement points is shown for the 4 grid sizes. At lower velocities the velocity magnitude varies a lot between the grid sizes. Even at the extra fine grid, a difference of 56% was simulated compared to the fine grid. At higher velocities the velocity magnitude is more stable. A difference of maximum 6% was simulated in the extra fine grid compared to the medium grid. For the key performance indicators deviations in low velocities don t have a large effect. Velocity magnitudes do also vary because this is a transient case and in this study it is regarded as a stationary case. FIGURE 15: GRID DEPENDENCY STUDY ON LOCAL AIR CHANGE INDEX OF THE BASE CASE When in the same simulations the local air change index is compared (figure 15), the differences between the simulations is a lot smaller. The largest difference simulated between the coarse grid and the medium grid is only 1%, making the coarse grid efficient enough. The rest of the simulations are performed with the medium grid size of 1.5 million cells. Including fluctuations, medium grid size does only vary 6% in velocity from finer grid sizes, while

13 the difference in velocity between the coarse grid and the medium grid is 36.5% near the extract. The medium grid is more consistent in the simulated velocities and therefore should be used to calculate the velocities. have a local air change index of around 100%, or higher in the jet area, in all points. 140% 3. Results In this chapter the results of the 5 case studies are discussed. Comparisons are made on the performance indicators LACI, draft temperature, velocities and temperature removal. In this chapter a couple of cross sections are shown. These cross sections are taken over the centre line of the room as shown in figure 13 (section A-A ). 3.1 Diffuser placement case 100% 90% FIGURE 18: LACI ON SECTION A A OF THE BASE CASE. HIGH VENTILATION EFFICIENCY IS PRESENT IN THE JET AREA IN THE CENTRE OF THE ROOM. 140% 100% 90% FIGURE 19: LACI ON SECTION A A OF THE ENVELOPE CASE. HIGH VENTILATION EFFICIENCY IS PRESENT IN THE JET AREA ON THE ENVELOPE SIDE OF THE ROOM FIGURE 16: COMPARISON OF VELOCITY BETWEEN THE BASE AND THE ENVELOPE CASE The diffuser placement has a large influence on the velocity magnitude (figure 16). For the base case the highest velocity is measured on position D1.7 (diffuser high), the centre of the room. In the envelope case a higher velocity is measured in position ED1.7 (envelope diffuser high). At position EXT (extract) high velocities are measured for both diffusers. The rest of the simulated velocities are low on both diffusers, although the base case has higher velocities on most simulated points. 3.2 Diffuser type case Axial diffusers have a downward jet area (figure 20). Fresh air will flow directly towards the centre of the room. Air will be induced from the centre of the room. Velocities in the middle of the room will be different, depending on which diffuser is chosen m/s 0 m/s FIGURE 20: VELOCITY VECTORS OF AN AXIAL DIFFUSER. THE JET IS DIRECTED TO THE CENTRE OF THE ROOM. By using another diffuser type the jets in the room are directed differently. Radial diffusers make use of the coanda effect (figure 21). Induction will take place in the upper part of the room. Air will travel across the ceiling when these diffusers are used. 1.7 m/s FIGURE 17: COMPARISON OF LACI BETWEEN THE BASE AND THE ENVELOPECASE Local air change index (figure 17) follows the same trends as velocity. In the respective jet area of the diffusers higher local air change indices are simulated (figure 18 and 19). Overall both diffusers 0 m/s FIGURE 21: VELOCITY PROFILE OF A RADIAL DIFFUSER. THE JET RADIATES OUTWARDS FROM THE DIFFUSER, STICKING CLOSE TO THE CEILING. 13

14 In figure 22 velocities of the axial base case and a radial case are compared. For almost all simulated points the velocity magnitude is a lot smaller in the radial case, except for the extract point (EXT) and the wall near the diffusers (DW). In the living area no high velocities were found for a radial diffuser. By using the radial diffuser the local air change index is well above 100% in all simulated points (figure 24). Only at point D0.1, below the diffuser the local air change index is higher in the axial case (figure 23). FIGURE 22: COMPARISON OF VELOCITY BETWEEN THE BASE AND THE RADIAL CASE centre, near the diffuser jet. Towards the envelope of the room, the temperature drops gradually. Because the temperature of the floor is warmer than the temperature of the wall, no stratification is present in the room. (Figure 25) 293.6K K FIGURE 25: TEMPERATURES ON SECTION A A OF THE HEATING CASE. When a cooling case is considered, the cold air from the diffuser jet is heavier than the warm air in the room, and thus drops faster. An increased velocity magnitude at position D1.7, diffuser high, is simulated. The velocity magnitude of the cooling case in the other points is fluctuating around the simulated levels of the base case. The cooling case matches the base case more than the heating case, but with increased velocities near the walls due to rising air caused by heat transfer at the walls. In the cooling case stratification does occur (Figure 26) K K FIGURE 26: TEMPERATURES ON SECTION A A OF THE COOLING CASE. FIGURE 23: COMPARISON OF LACI BETWEEN THE BASE AND THE RADIAL CASE 140% Although the local air change index is not as high in points close to the jet (figure 27), the heating and cooling cases have a higher local air change index in more remote points in the room. The lower central part of the room (point C0.1) and the kitchen (points K0.1 and K1.7) are ventilated more efficiently in the non-isothermal cases. 100% 90% FIGURE 24: LACI ON SECTION A A OF THE RADIAL CASE. THE VENTILATION EFFICIENCY IS HIGHER IN THE LARGEST PART OF THE ROOM, COMPARED WITH THE BASE CASE. 3.3 Heating and cooling cases When compared to the axial isothermal base case, the heated axial case has higher velocities on all positions, except for position D1.7, the high position below the diffuser. In the heating case the diffuser jet is slowed down by the cooler and denser air in the room. Temperature in the room is higher in the FIGURE 27: COMPARISON OF LACI BETWEEN THE BASE CASE, AXIAL HEATING, AND AXIAL COOLING. 14

15 The local air change index is closer to 100% in all points. Buoyancy driven flow increases velocities near walls, which increases mixing in these parts. Heating and cooling both cause improved mixing in the room. The axial diffuser has an effective draft temperature over 1K at a lot of points in the domain, suggesting that the temperature can rise too much at these points (figure 30). At position D1.7, below the diffuser, the effective draft temperature reaches to 0.92, which is still between the requirements of ASHRAE, but suggests that some draft problems can occur there. A percentage draft calculation (ASHRAE Handbook: Fundamentals, 2009) indicates a draft risk of 16% at position D1.7. The draft risk at the same position with the radial diffuser is only 6%. The effective draft temperatures of the radial case are better than in the axial case, being between -1.5K and 1K in most points. FIGURE 28: COMPARISON OF VELOCITY BETWEEN THE BASE CASE, AXIAL HEATING, AND AXIAL COOLING. In a cooling case the axial diffuser has an even higher velocity than in an isothermal case (figure 28). Especially with cooling, draft can be an issue. Applying a radial diffuser for cooling decreases the velocity magnitude at position D1.7, just below the diffuser by a large amount (figure 29). The velocities in other parts of the room are in the same order of magnitude for both diffusers, although the radial diffuser generates lower air velocities in most points. Just like the isothermal cases, in the cooling cases radial diffusers have a higher local air change index in most points compared to axial diffusers. To further investigate cooling effects in a room, the radial diffuser is also investigated. If the isothermal radial case and the radial cooling case are compared, the LACI of the cooling case is more evenly distributed over the room. The isothermal case does have a couple of points where the LACI is higher than in the cooling case, but also points where the LACI is lower. Cooling increases mixing in the radial case as well. FIGURE 30: COMPARISON OF EFFECTIVE DRAFT TEMPERATURES BETWEEN AXIAL AND RADIAL COOLING. The warm sensation which a high effective draft temperature suggests can also be seen in the effective temperature removal of the axial case (figure 31). In most points not all the generated heat can be removed. Only near the diffusers heat is effectively removed, but draft is caused in the process. In most other points the temperature is higher than the temperature in the extract. The radial diffusers perform better in heat removal. In all points heat is removed. FIGURE 31: COMPARISON OF EFFECTIVE TEMPERATURE REMOVAL BETWEEN AXIAL AND RADIAL COOLING. FIGURE 29: COMPARISON OF VELOCITY BETWEEN THE AXIAL COOLING AND THE RADIAL COOLING CASES 15

16 3.4 Hot patch case The partially heated floor causes increased air velocities in the heated part of the room, as air rises from the heated area. The increased flow in this part of the room causes this part of the room to be ventilated better. This happens at the expense of the air quality in the other part of the room, which drops to 81% at the wall (figure 32). Fresh air from the diffuser is drawn into the heated part of the room. The increased flow does have a positive effect on the temperature removal in this part of the room. At 0.1m from the floor an effective temperature removal ratio of 60% is achieved, which is not very effective. In other parts of the room the effective temperature removal ratio is well above 100%, even above the heated floor at breathing level. FIGURE 32: LACI ON SECTION A A OF THE SUN HEATED CASE. MORE AIRFLOW IN THE HEATED PART OF THE ROOM CAUSES AN INCREASED VENTILATION EFFICIENCY IN THAT ZONE. 4. Discussion In the comparison between the base case and the envelope case, the jet area has the largest influence on both velocity and local air change index. Outside of the jet area velocities are low, and LACI are between 90% and 110% for both cases. Inside the jet area of both diffusers high velocities are simulated, and the local air change index is well above 100%. Both cases ventilate efficiently. According to ASHRAE, the local air change index for these systems should be at least 80% in heating cases and 100% in cooling cases ( ASHRAE 62.1: Ventilation for acceptable indoor air quality, 2013).The choice for where to place the diffuser could be based on other factors, as for ventilation efficiency, placement has no influence. The high jet velocities could cause draft below the diffusers. The diffuser type has a large influence on both velocity and local air change index. Axial diffusers have a jet pointing into the centre of the room, while radial diffusers have a jet area along the ceiling. For axial diffusers high velocities are present in the centre of the room, while for the radial diffuser no high velocities were present in the living zone. Radial diffusers have a better ventilation efficiency in almost all points averaging at 115%, compared to 100% for axial diffusers. Both diffusers are highly inducing but induce in different places. Axial diffusers induce in the centre of the room, while radial diffusers induce in the top side of the room. When the room is heated via an axial diffuser the velocities in the room increase, except for the jet velocity which decreases. In a cooling case the jet velocity is increased while the other velocities stay approximately the same. When heating or cooling is applied mixing in the room is increased. The high local air change index near the jet area of the axial diffusers is lowered, while the other points in this room reach closer to 100%. Cooling with the radial diffuser equalizes the local air change index over the simulated points, averaging at 113%. Heat removal is not effective on all points when an axial diffuser is used. The radial diffuser is capable of removing excess heat in all points. The axial diffuser causes a 16% predicted draft rating below the diffuser in the simulated conditions, while the radial diffuser only has a 6% draft rating at the same position and lower everywhere else. Heating from sunlight does affect the local air change index in the non-heated part of the room negatively by causing an increased flow in the heated part of the room. The local air change index dropped 18% in the non-heated part in this case study. In this study the CFD models were not validated, because no suitable measurement room was available. The results cannot be assumed to match reality exactly but can be used for a comparative investigation between the case studies. Models for flow, turbulence, buoyancy, age of air, energy are identical over the different case studies, as well as the non-case specific conditions. A comparison between changes of conditions can thus be made. For future research a validated CFD model of an L- shaped room should be made. In a measurement study, the LACI and contaminant concentrations of an L-shaped room can be investigated. Validation can improve LACI calculations in the model and the contaminant concentration measurement can be used to investigate the actual air quality in the room. The models are investigated as steady state rooms, where no movement and no extra internal heat sources are applied. In reality people will be walking in the rooms, generating heat and movement, furniture will be placed, and heat generating appliances are used. These factors have a large influence on the airflow and ventilation efficiency in the room. Also in the non-isothermal models a single temperature is set over the walls. In reality less heat 16