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Air heating system design for a sub-Arctic climate using a CFD technique
Building and Environment 160 (2019) 106164
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Air heating system design for a sub-Arctic climate using a CFD technique ∗
T
Petter Lundqvist , Mikael Risberg, Lars Westerlund Division of Energy Science, Department of Engineering Sciences and Mathematics, Lulea University of Technology, Lulea, Sweden
A R T I C LE I N FO
A B S T R A C T
Keywords: Air heating CFD Cold climate PPD Sub-Arctic climate Thermal comfort
The thermal comfort in a residential building equipped with an air heating system and located in a sub-Arctic region was investigated with computational fluid dynamics (CFD) software. The predicted percentage of dissatisfied (PPD) was used to identify flaws with the heating system during winter conditions. New scenarios were simulated and compared to each other to see potential improvements of the thermal indoor climate. Comparison was done by combining the discomfort spaces inside rooms, the level of the discomfort and the time spent in these spaces. The discomfort covered 8–38% of the interior volume depending on the test case. The results provide the necessary means to create a satisfactory thermal indoor climate if an air heating system is to be utilized in sub-Arctic regions during the winter. The correct heat demand for each floor and appropriate placement of the supply devices are required. Adding air transfer units or grilles in rooms from which exhaust air is removed further improves the comfort. The results also show the strength of using CFD technique when investigating the indoor discomfort with PPD, and how a fair assessment can be done by combining the PPD with time.
1. Introduction As a result of low energy buildings becoming standardized as residential buildings in cold climate regions, the required heat demand is reduced. It is established that an improved building envelope changes the air flow in buildings due to less infiltration and convection flows. Mechanical ventilation is therefore commonly required when natural ventilation is inadequate for supplying air, which in turn also affects the thermal indoor climate. Mechanical ventilation gives the possibility of heating the air and using it as a heat source, which is a feasible approach that has been utilized in passive houses in Europe [1]. Usage of heated air instead of radiators or other hydronic heating systems in residential buildings during the cold months in southern Sweden has been studied, and although risk of overheating has been observed, the approach has been considered feasible [2]. Risberg et al. [3] investigated different heating systems in the subArctic climate, and found that an air heating (AH) could create a good thermal comfort. The AH system in that study was similar to the heating system investigated by Gustafsson et al. [4], who considered it to be the best option for a well-insulated house in a cold climate. Risberg et al. [5] continued to investigated the AH system in a pilot passive house in the sub-Arctic region of northern Sweden, and showed that the AH was feasible from an energy point of view. If an AH system is utilized in a
passive house, there is no need for the installation of pipes and radiators, and thus an AH system can be less expensive compared to underfloor heating or a hydronic radiator system [3], which are integral parts of the traditional heating system in Sweden. However, the findings so far regarding the AH system in sub-Arctic region point out problems with the thermal comfort, which was not satisfactory due to mainly overheating in the cold winter months. The phenomenon of overheating during winter has been seen in other passive buildings in climates ranging from cold to Arctic climates [6–9]. Ji et al. [10] noted overheating in a passive house during the winter in a cold region of China when performing research with an outdoor temperature of −23.4 °C. The indoor measurements exhibited temperatures above 24 °C, which were caused by added solar gains, the occupants and indoor apparatus. This implies the heating system was not properly designed. Horikiri et al. [11] showed that occupants can have a significant impact on the thermal indoor climate, however, the heating system should be designed to withstand cold days with no internal gains such as the occupants. Both internal and solar gains can be left out completely when designing the AH system in passive houses [8]. The system should be able to regulate itself when necessary, i.e. when occupants are present. In the same study by Horikiri et al. [11], it was also concluded that furniture had an influence on the local air velocities around
∗ Corresponding author. Luleå University of Technology, Division of Energy Science, Department of Engineering Sciences and Mathematics, SE-971 87, Luleå, Sweden. E-mail addresses: [email protected] (P. Lundqvist), [email protected] (M. Risberg), [email protected] (L. Westerlund).
https://doi.org/10.1016/j.buildenv.2019.106164 Received 4 April 2019; Received in revised form 13 May 2019; Accepted 24 May 2019 Available online 25 May 2019 0360-1323/ © 2019 Elsevier Ltd. All rights reserved.
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possible visualize detailed air flows, temperature gradients and also investigate the Predicted mean Vote (PMV) and Predicted Percentage Dissatisfied (PPD), as described by Fanger [21]. However, while the PMV and PPD has been visualized with CFD software, most studies focus on planes as a method of investigation and presentation. One of the strengths with CFD software is to simulate a whole building interior, which gives the possibility to investigate the PMV and PPD in the form of a volume so that it gives a better understanding of the whole thermal comfort simultaneously, rather than a plane. The PMV and PPD consider air humidity, air temperature, air velocity, mean radiant temperature, clothing thermal insulation and metabolic rate. In order to make a more correct evaluation of the thermal indoor climate, two other aspects could also be considered, especially if a whole building interior is investigated. The first aspect is where occupants dwell, in other words, the occupied zone. Definitions of the occupied zone can vary but are in general the zone between 0.05 and 1.80 m on horizontal level, 1.00 m from windows and doors, and 0.50 from both internal and external walls [22]. According to Swedish regulations the occupied zone is the zone between 0.10 m and 2.00 m above the floor level, 1.00 from windows and doors, and 0.6 m from the external walls [23]. The second aspect is the amount of time people spend in each room, which should be considered and weighed in. A lower performance of the thermal indoor climate in the laundry room, for example, can have a significant impact on the total PMV and PPD if the whole apartment is being considered. In reality, however, this is not the case, as the occupants will spend more time in bedrooms and living rooms, as shown by Khajehzadeh et al. [24]. Based on the same pilot passive house as in the previous study, the main objective of this study was to establish that it is possible to achieve a satisfactory thermal indoor climate using an AH system in a sub-Arctic climate. This type of heating system can reduce the construction cost by removing the need to install a hydronic system. A secondary objective is to show how PPD can be combined with time to make a fairer assessment of the thermal indoor climate. The investigation of the thermal indoor climate was performed with CFD software. The current settings of the AH system were analyzed by rendering the PPD as volumes in the building model which was validated with experimental data. Using the PPD volumes, it was possible to identify the flaws associated with the current settings, placement of the supply and exhaust devices and the air transfer inside the building. New solutions were proposed, simulated and compared with each other to see changes in the thermal comfort. The results obtained in this study can in turn provide guidance for the future design and use of AH systems in sub-Arctic regions, as well as provide directions of how to
Abbreviations AH semi DWOT UDFs COM
Air heating Semi-detached house Design winter outdoor temperature User-defined functions Comparison number
the furniture, but the buoyancy strength was not affected significantly. Heat generated by electric appliances such as TVs can be neglected when the thermal indoor climate is simulated [11]. In the sub-Arctic regions of Sweden snow is present during the cold winter months. The snow surrounding a passive house has an insulation effect on the foundation, but for a passive house with a thick insulation, the added effect from snow is not significant enough to be considered during simulations and design of an AH system [12]. Georges et al. [8] investigated AH systems in a cold and sub-Arctic climate in Norway. They used the TRNSYS software and concluded that if AH was to be used, there would be some limitations and certain conditions that would have to be addressed for sub-Arctic climates. Something else to consider is that the amount of heat air also must meet the minimum requirement for supplied air. Although it has been shown that an air flow of 0.2 l/(s m2) of floor area can be sufficient [13], the Swedish requirement for the air change rate is generally 0.35 l/(s m2) of floor area [14], which gives a lower boundary for the amount of supplied air. As mentioned, in the previous study by Risberg et al. [5] from the northern Sweden, the thermal indoor climate was not satisfactory in the whole building. One of the flaws of the AH system in the building was the excessively high mass flow of air on the first floor, which led to the overheating of part of the building, while the bathrooms and laundry room were too cold. The uncertainty associated with the design, installation and operation of an AH system in sub-Arctic regions remains. Risberg et al. [5] therefore suggested that the AH system should be studied further to detect problems and suggested new settings to improve the thermal indoor climate. While software such as TRNSYS can give good estimations of the thermal comfort in a building, it does not give a full understanding of the thermal indoor climate, in contrast to computational fluid dynamics (CFD) simulations, which can provide detailed results for the indoor climate throughout the whole building. CFD as a tool for thermal indoor climate simulations has been shown to be able to provide a detailed view of the thermal indoor climate [10,11,15–20]. With CFD it is
Fig. 1. (a) Top-view of the building location showing the north-south direction. The bottom half of the building is the southern semi, while the top half is the northern semi. (b) Front view of the building, with the southern semi to the left. 2
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compare different thermal indoor climates using PPD.
Table 1 U-values for different parts of the building.
2. Methods 2.1. The building The building is located in Kiruna just south of latitude 68°N, built to withstand polar nights throughout the winter, along with the possibility of severe cold between −30 °C and −40 °C [25]. The building was built as a pilot case test building to test the AH system among other low energy techniques. The building is made up of two semi-detached houses (semi) and has a concrete foundation; the front of the building faces roughly east-south-east, as seen in Fig. 1 (a). A front view of the building is provided in Fig. 1 (b) to show the two semis. The total area for living which has to maintain a satisfactory indoor climate in each semi is 134 m2. The layout of each semi is a mirror image of that of the other semi. The ground floor contains one bedroom, living room, kitchen, bathroom and laundry room, as well as an entrance hall with a staircase to the first floor. The first floor contains three bedrooms, a family room, a bathroom and a storage room. Due to the semis mirror each other, it is only necessary to perform the investigation on one of them. The semi focused on in this study is the southern one, and a layout of this house can be seen in Fig. 2. The cardinal direction did not have to be considered since the investigation was performed in a period of polar night and sun did not affect the building. The heated air is supplied to the bedrooms, living room and family room via circular supply devices which are 125 mm in diameter and are mounted in the ceiling. The supply devices will henceforth be referred to as inlets. For the exhaust air, the kitchen, bathrooms and laundry room have circular outlets which are 100 mm in diameter and are located in the ceiling. These exhaust air devices will henceforth be referred to as outlets. The placement of the inlets and outlets on the two floors is illustrated in Fig. 2, where the inlets are denoted by circles and the outlets by crosses. Each bathroom is equipped with an electric towel heater as a heat source, a 50 W towel heater for the ground floor and a 102 W towel heater for the first floor, and these heaters should in theory be able to maintain a comfortable room temperature. The towel heaters are denoted by rectangles in the bathrooms in Fig. 2.
Section
U-value [W/(m2∙K)]
Envelope area [m2]
Roof External walls, ground floor External walls, first floor Foundation Windows, ground floor Windows, first floor Front door
0.035 0.083 0.083 0.139 0.650 0.650 0.700
67.1 53.2 46.1 67.3 11.0 12.5 2.10
first floor because this floor has more rooms with supply air [5]. To sustain the thermal indoor climate during winter conditions, a supply air temperature of 45 °C was needed, and this is the temperature used in this study. Based on the heating demand, new mass flows for each inlet were calculated separately for the two floors. EN ISO 13789:2007 and the Swedish regulation FEBY18 were used to determine the heat loss from the building [26,27]. Concerning the envelope area, the U-values for the external walls, roof, foundation, windows and front door were based on the blueprints and are presented in Table 1. The total infiltration was measured to be 87 l/s at a pressure of 50 Pa between the indoor and outdoor conditions, done accordingly to standard EN-13829 [28] with a blower door. In normal conditions this correspond to approximately 0.02 l/(s m2) interior envelope area. A wind coefficient of 0.7 was used and this value corresponds to a moderate shielding from the wind, such as that typical of a suburban environment. It was assumed that the infiltration rate was the same on both floors. The heat transfer coefficient for the external walls was recalculated and changed from 0.079 to 0.083 W/(m2∙K) to take the energy loss from infiltration into consideration, and the latter value was used as the overall heat transfer coefficient for the walls. The mass loss of air due to the infiltration corresponded to less than a permille of the indoor air, and therefore assumed to not have a significant impact on the indoor air, and therefore not included in the CFD model. Based on SS-EN ISO 15927-5, a seven-day value for the design winter outdoor temperature (DWOT) was used, giving an outdoor air temperature of −25.7 °C, and the ground temperature beneath the building was set to −7.3 °C [29]. For the calculations of heat demand, an indoor temperature of 21 °C was used in all the rooms except the bathrooms, where 25 °C was used. Other data used had been gathered by NCC AB Sweden through investigation of the building [28]. The total heat demand for the ground floor in these conditions was manually calculated to be 897 W, and the corresponding demand for the first
2.2. Heating demand and required air flows The previous study of the building has shown that the original design of the AH system allocates an excessive amount of air flow to the
Fig. 2. Location of the inlets (circles), outlets (crosses) and towel heaters in bathrooms (rectangles – not according to scale). The dashed line represents a plane that is used later for presentation of results. 3
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floor was likewise calculated to be 796 W. The supplied air for each floor was calculated as sensible heat. The rooms with supply air will have higher air flows than those actually needed, because the heat is not supplied to all the rooms, and rooms with no supply air relied on warm air from other rooms. Table 2 presents the air flows for each room, both according to the original design and according to the heat demand on each floor. The newly calculated supply air for the whole semi corresponds to an air change rate of 0.43 l/(s m2) of floor area. The new total supply air mass flow was 0.064 kg/s for the semi, with 55% distributed to the ground floor and 45% distributed to the first floor. In the original design, a total supply of 0.075 kg/s was used, with 37% distributed to the ground floor and 63% distributed to the first floor [5]. Since the total supply air flow was 15% lower than that in the original design case, the exhaust flow was lowered by 15% for each outlet. In the original design case, no towel heaters were implemented. For comparison the original design case was simulated with the towel heaters included. To match the heat demand, the temperature was set to 44 °C.
applied. The criteria used to reach a solution were the stabilized mass, heat transfer and turbulence residuals, along with the stable temperatures in the center of each room at a height of 1.1 m. The external surfaces were assumed to have a uniformly heat loss corresponding to their respectively heat transfer coefficients in Table 1 as boundary conditions. As previously stated, the energy loss due to infiltration was included for the external walls. Mass transport of the air through the walls was not included in the CFD model since it was assumed that it did not have any significant impact on the indoor air as previously stated. The internal walls and intermediate floors in the model were set as solids, with a thermal conductivity of 0.05 W/(m · K). The internal wall shared with the northern semi was set as adiabatic. Supply devices were created as inlets and exhaust devices as outlets. Each inlet and outlet were given the mass flow according to Table 2. To be able to investigate PPD with CFX-Post, a PPD variable was created that was connected to an expression that calculated the PPD [5]. The PPD comfort parameter was calculated according to ASHRAE Standard 55–2010 [32]. The cloth factor was set to 1 Clo, while the metabolic rate was set to 58.2 W/m2 and humans were considered to be 1.7 m2 to correspond to the previous study [5]. The pressure for the humidity was set to 600 Pa to correspond to measured relative humidity during the validation period. Draught rate was also created as a variable, linked to an expression, based on the ISO 7730 standard [33].
2.3. Computational fluid dynamics model The ANSYS CFX 19.2 software was used to solve the numerical model. The post-processing of the results was performed in CFX-Post. The geometry used contained the whole interior of the southern semi as one volume. The glazed porch at the front door of the real building was excluded, as this section does not contain any heating. The temperature in this part was considered to be the same as the outdoor temperature. The original validated CFD [5] model of the building was used, including the mesh that was produced accordingly to a previous grid size study, which also established the computational setup for the model [16]. The model was simplified by removing the staircase and appliances in the bathrooms and the kitchen. The settings used for the mesh include having the grid size maximum cell face length of 0.1 m, with inflation layers applied to all the zones delimiting the room with a growth rate of 1.5. For air transfer units or grilles between rooms the elemental size was set to 0.01 m. For smaller volumes between closed doors and floors (slits), the element size was set to 0.01 m on horizontal surfaces and 0.001 m on vertical surfaces. The total amount of elements in the model varied from 4.5 million to 4.9 million, because of different features being added later on. Because a worst-case scenario was used, based on SS-EN ISO 159275 [29], and thermal comfort by using PMV/PPD was to be investigated, only steady state solution was calculated. The simulation was set up according to the previously validated simulation [5], which was validated during polar night period of December 16th. This resulted in 0.00 W/m2 from solar radiation, confirmed by measurement data from the Swedish Meteorological and Hydrological Institute for the validated time period. Additional heat transfer due to sky radiation and longwave radiation from nearby buildings was calculated but did not have any significant impact on the simulations for this case. A second order upwind discretization scheme with RMS residuals set to 10−6 was used for the simulations. The fluid model used a total energy model and a standard k-epsilon turbulence model, and the discrete transfer model was used for the thermal radiation. The thermal radiation for internal surfaces had an emissivity of 0.9, while the windows had an emissivity set to 0.83. The boundary conditions for internal convection were defined with user-defined functions (UDFs), which were written based on the correlations from Awbi and Hatton [30] and were created based on the fact that heat was distributed along the ceiling in this case. The use of UDFs with the k-epsilon model have provided reliable results with relative fast simulation times [31]. The initial indoor air temperature of the model was set to 21 °C, while the gauge pressure and momentum in all directions were set to 0 Pa and 0 m/s, respectively. Due to CFX using a pseudo time step for solving steady state simulations a timescale is needed, and in order to reach convergence faster with CFX, a physical timescale of 1000 s was
2.4. Improvements of the indoor climate In order to improve the thermal indoor climate, new scenarios were simulated. The scenario according to the original design and how the house was built was named Case 1.0. Another four cases were tested and are presented in Table 3, to give an overview of the cases. Case 1.0. had been validated previously through continuous measurements of the air temperature and relative humidity in each room, supplemented with manual measurements of the air velocities [5]. It should be noted, however, that the previous study did not include towel heaters in the bathrooms. Along with the towel heaters and the recalculated heat demand for the seven-day DWOT, the original case required a supply air temperature of 44 °C, which was used for Case 1.0 to compare the original design with new scenarios. Case 1.1. introduced the rearranged mass flow of supplied air on each floor according to the heat demand for the respective floors. It was expected that the adjusted heat demand would not be sufficient to create a satisfactory thermal indoor climate. The previous study showed problems with heat concentrating in some regions with subsequent overheating. At the same time, some areas had a deficit of heat. On the ground floor, all the inlets are placed in one side of the semi, which creates an uneven distribution of heat. The placements of the inlets on the first floor were also expected to create an uneven distribution. Table 2 Supply/exhaust air flows for each room.
4
Location
Floor
Mass flow Supply/exhaust Original design [g/s]
Mass flow Supply/exhaust Corrected for heat demand for each floor [g/s]
Bedroom 1 Living room Bedroom 2 Bedroom 3 Bedroom 4 Family room Bathroom 1 Kitchen Laundry room Bathroom 2 Total
Ground floor Ground floor First floor First floor First floor First floor Ground floor Ground floor Ground floor First floor
12.5/15.0/12.5/12.5/14.0/8.50/-/15.0 Open -/15.0 -/25.0 75.0
8.15/13.5/6.71/3.81/5.88/12.4/-/12.8 Open -/12.8 -/21.3 63.9
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Table 3 Scenarios simulated.
Table 4 Time spent in each room.
Case
Settings
1.0 1.1 2.1 3.1 3.2
Original design Changed air mass flow/floor Case 1.1 + New placement of inlets Case 2.1 + Horizontal units Case 2.1 + Vertical units
Bedroom 1 Bedroom 2 Bedroom 3 Bedroom 4 Kitchen Living room Family room Entrance hall Bathroom Laundry room Storage room
Case 2.1. was created from Case 1.1 to avoid the heat concentrating in some areas by moving two inlets. Case 2.1 kept the same mass flow rate of air through each inlet. One of the inlets in the living room was moved to the entrance hall and the inlet in the family room was moved to a position above the stairs, according to Fig. 3. Because of the doors to the bathrooms and the laundry room remain closed in the simulations, which also is the case in practice most of the time, the air has difficulty in flowing into these rooms and can only enter through small slits under the door. Two different types of air transfer units, or transfer grilles, were added to Case 2.1 and were placed in the internal walls of the bathrooms and laundry room, creating Case 3.1 and Case 3.2, respectively.
Hours
% of total time
2.27 2.27 2.27 2.27 2.89 2.89 0.77 0.28 0.78 0.11 0.01
14 14 14 14 17 17 4.6 0.8 4.6 0.7 0.06
investigations, a distance of 0.6 m from the external walls was used for the windows and doors, and a distance of 0.1 m was added for the internal walls. The volumes were colored according to a PPD scale from 10% to 100%, to clarify further how severe the problems were. To make a fairer assessment, the time per day spent in each room was included as well. It is more important to have a good climate in a room where the occupants spend a large amount of time compared to a room where the occupants only spend a short time. In order to compare the different air distribution arrangements for the different cases, a comparison number (COM) was created, calculated as follows,
Case 3.1. used the more traditional approach with horizontal transfer units with the dimensions (W*H) 700 × 100 mm and centered 150 mm above the door frame. Case 3.2 was provided with vertical transfer units along the door frame. The same cross-section area was kept for these units, but the dimensions were changed to (W*H) 39 × 1800 mm and the units were placed 40 mm from each door opening and 100 mm above the floor level. By keeping the same cross-section area, the two types could be compared. A high narrow opening along the vertical door frame was expected to increase the air exchange between the rooms due to convection flows. All of these units were made as part of the interior volume between the rooms to give an overview of the flow behavior and to determine whether the thermal comfort could be increased with the help of the transfer units.
COMj =
∑
(PPDj (%)⋅VOLj (%)⋅Timej (%)).
For each room in the semi, the average PPD value above 10% in the occupied zone (PPDj(%)) was multiplied by the volume where the PPD value was above 10% (VOLj(%)) and the amount of time spent in that room (Timej (%) ), based on the numbers given in Table 4. The results for all the rooms were then summed up to create the COM value for each case. The comparison number does not include the fraction which each room volume constitutes of the total volume of the semi, since the time spent in the respective rooms is included. The COM value is only used as a reference to make it possible to compare the improvement of one case with that of another. The numbers in Table 4 was based on a study by Khajehzadeh et al. [24]. Although this study was performed in New Zealand, similar results can be expected in Sweden. It was assumed in the present study that, in total, 16.8 h per day were spent at home on average (calculated on an annual basis) for the setup consisting of a separated living room, kitchen and dining room. The family room was considered as a room used for a combination of play, study and games. Moreover, it was
2.5. Evaluation of the scenarios According to ASHRAE Standard 55–2010, an acceptable indoor climate when using PPD is a value up to 10% [34]. To detect problems in the semi more easily, the PPD variable was plotted for each element and presented as volumes when the PPD value was greater than 10% within the occupied zone. To simplify the occupied zone for PPD
Fig. 3. New position of two inlets. 5
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room, but had an average value of 23 °C. The radiant temperature reached 28 °C in the living room, and together the air temperature and the radiant temperature created dissatisfaction in almost all the occupied zone. This shows that the amount of air supplied to the living room was excessive due to the placement of the inlets. Out of the 897 W needed for the ground floor, 651 W were supplied to the living room, representing 73% of the ground-floor heat demand. Case 1.1 clearly shows that air was not properly supplied to the semi. The heat demand is the same throughout the semi, but on the ground floor the air was supplied in one side of the semi, which created a thermal imbalance. In Case 2.1, two of the inlets were moved to new positions, decreasing the COM value to 2.7, which is a significant improvement compared to Case 1.1. The problematic volumes can be seen in Fig. 6, where the large volume in the living room has almost disappeared. In Case 2.1 the volumes with PPD values above 10% covered in total 22% of the occupied zone, with an average PPD value of 16%. A comparison of the results for Case 1.1 and Case 2.1 clearly shows that the heat was spread more evenly in the latter case. The average temperature difference between each floor in Case 1.1 was 1.1 °C, with the ground floor being warmer. In Case 2.1 the corresponding difference was 0.2 °C, with the ground-floor temperature decreasing and the firstfloor temperature increasing. After one of the inlets was moved from the living room, the heat was no longer concentrated in that part of the ground floor. However, the inlets in the entrance hall and above the staircase created dissatisfaction instead. The air supplied at the groundfloor entrance moved up to the first floor and the heat accumulated together with the air supplied on the first floor. The plume of hot air moving up to the first floor contributed to a flow of air going down along the staircase. Consequently, there was a significant increase in the PPD value of the occupied zone at the stairs. While this affected the results, this specific area was mainly on the first floor above the groundfloor staircase part, which in reality would not affect the thermal comfort. The problem in the laundry room remained, since this room showed no improvement, even though an inlet was located outside the room. The local velocities along the floor remained high, and local velocities of 0.5 m/s appeared in all the exhaust air rooms. However, the use of transfer units in Case 3.1 and 3.2 resulted in improvements for these rooms. The COM value decreased to 0.8 and 1.3 in Case 3.1 and 3.2, respectively, as seen in Fig. 4. In Case 3.1, the PPD volume covered 8% of the occupied zone, with an average PPD value of 15%. In Case 3.2, the PPD value covered 10% of the occupied zone, with an average PPD value of 15%. In order to see the effect of the transfer units, especially the vertical one in Case 3.2, a plane was created 1.68 m from the entrance door, visible as a dashed line in Fig. 2. The PPD value, air temperature and velocity were investigated separately, and are presented in Fig. 7. From Fig. 7 (a), one can observe that while the PPD values were at low levels of 5–10% in most of the plane, the laundry room still had some problems, with the average PPD value remaining high at 39% and 40% for Case 3.1 and Case 3.2, respectively. The increased velocity does not increase the PPD which is due to an increase of the temperature with fresh air from the inlet. By comparing the laundry room with the first-floor bathroom, which both have the same layout and external area, it is possible to see the benefit of the towel heater, which is helping the AH system to create a better thermal comfort. Both of the
assumed that the amount of time spent in bedrooms was 9 h, and this time was divided by the number of bedrooms because the comparison was based on a summation of the time spent in all the rooms in the semi. The total time spent in the kitchen and living room was divided into two equal parts. 3. Results and discussion The COM value result for each case is presented in Fig. 4. A lower COM value indicates better thermal comfort. With the original design in Case 1.0, the COM value was 30.4. This value was decreased in the following cases when the air flow was adjusted to match the heat demand and there was a better distribution of the supplied air on each floor, implying a better thermal comfort. The results regarding Case 1.0 are not discussed in detail in this paper as this case was covered by Risberg et al. [5]. All the other cases were investigated separately in detail, which includes pressure fields, PPD, draught and temperatures. For each case, a visualization was created of the volumes inside the occupied zone where the PPD value was above 10%. This was done by plot the created PPD variable for each element within the occupied zone. An example of such a visualization is provided in Fig. 5, which shows the volumes for Case 1.1, where the original design of the inlets was kept and there were no added features to help distribute the air flow. The volumes visualized in the blue-to-red scale have a PPD value above 10%. Even if the result was not satisfactory, the COM value showed a remarkable improvement in the thermal indoor climate compared to the original design, Case 1.0, in that the COM value was decreased almost by a factor of four. In Case 1.1 the PPD volumes covered in total 38% of the occupied zone and had an average PPD value of 19%. The rooms which can be seen in Fig. 5 (a) to have had major problems in Case 1.1 are the exhaust air rooms along with the storage room. In Fig. 5 (b), with the building shown from a reverse angle, one can see that almost all the occupied zone of the living room and bedroom 1 was covered by PPD volumes. This clearly indicates the problem of an uneven distribution of air on the ground floor. Concerning the ground-floor bathroom, dissatisfaction was shown within 75% of the occupied zone, with an average PPD value of 26%. The first-floor bathroom results showed that 65% of the occupied zone was covered by PPD volumes, with an average PPD value of 25%. Both the air and the radiant temperatures were higher inside the PPD volume than outside the PPD volume, while the air velocities remained at similar value. This implies that the main problem in the bathrooms and living room is overheating. The radiant temperature in the bathrooms on both floors reached 26 °C, but, combined with the air temperature and air velocities, the operating temperature was 24 °C. For the bathrooms this was actually not a real problem due to these rooms are designed for an increased temperature. The laundry and storage rooms were too cold. Because the outlets are located in rooms with closed doors and air was entering through the slit below the door, a draught was created along the floor, which also increased the dissatisfaction. The local velocities reached 0.5 m/s in the bathrooms. The first floor had an average temperature of 21 °C, while the ground floor had an average temperature of 22 °C. The laundry and storage rooms had excessively low values both for the air and the radiant temperature, with temperatures of 18 and 19 °C, respectively. The storage room has no outlet and only exchange air flowing through the slit beneath the door. Both these rooms had an unsatisfactory thermal comfort, but this did not necessarily cause problems due to the rooms’ functions. When calculating the COM value, these rooms were basically neglected due to the time factor. The storage room will therefore not be discussed in detail. The laundry room is still of interest as it is one of the rooms which has an outlet but does not contain any heat source. In this room the whole occupied zone is problematic, with an average PPD value of 41%. The highest air temperature was 25 °C and was found in the living
Fig. 4. Comparison numbers for the different cases. 6
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Fig. 5. Volumes with PPD values above 10% in the occupied zone for Case 1.1: (a) view towards the gable and front of the semi, and (b) view towards the wall shared with the other semi and the back wall.
Fig. 6. PPD volumes above 10% in the occupied zone for Case 2.1: (a) view towards the gable and front of the semi, and (b) view towards the back and the section bordering on the adjacent semi.
overall, the velocities were lower in Case 3.2 than in Case 3.1. The transfer units used in the study are only represented as part of the interior volume, whereas in reality some sort of plate or grid are mounted on the units. This will affect the air flow compared to an opening in the wall, and any further study or optimization of this system also have to consider this. In all of the cases the kitchen had small PPD volumes. The largest PPD volume observed in the kitchen, with the original case excluded in the comparison, was for Case 1.1, where the PPD volume covered 9% of the occupied zone. In all the other cases it was lower than 3%. The kitchen has an exhaust air outlet only, but with two door openings the velocities were kept low and did not cause a draught. This type of flow is good for thermal comfort, and an open-plan living area where the kitchen and living room are combined could be preferable with an AH system. This would most likely create a better distribution of the air flow with no interior walls obstructing the flow. Of the other exhaust air rooms, the bathrooms were the ones most used, and these rooms required a higher air temperature than the rest of the semi. In Cases 3.1 and 3.2 in this study, the operating temperature for the bathrooms was around 24 °C, which from a thermal comfort point of view is good. Investigating the PPD in the manner selected for this study, with the Clo-parameter set to a static value of one, is not a perfect approach, and this has to be taken into consideration when analyzing the results. Correct air flows can also reduce energy use, if an excessive amount of air is supplied. In this study, the air mass flow was reduced with 15%
bathrooms are located at the gable corner, where the outdoor temperature will have the largest impact on the building. Without the towel heaters the temperature would drop below acceptable levels, even with the transfer units. This can be regarded as a major flaw of the AH system design, which forces these rooms to rely on an extra heat source. It is not unusual that bathrooms and laundry rooms are placed together at one end of the building, and not directly adjacent to bedrooms or living rooms. The problems connected with this have to be considered when designing an AH system. An AH system will always contribute to a temperature difference between rooms, with the supply air rooms always being warmer than the exhaust air rooms. The difference in temperature between these types of rooms decreases for Case 2.1 and even further for Case 3.1 and 3.2. To decrease this difference even further, the exhaust air rooms should ideally be placed in the middle of the building, so that it would be possible to supply heat to all the rooms with external walls. In the case of the studied building, the bedrooms should be located at the gable, and the bathrooms should share a wall with the adjacent semi. A comparison of the results obtained with the two different transfer units reveals that the horizontal traditional unit gave a better overall result with a lower COM value. Although it is not clearly shown in Fig. 7 (c), the vertical units produced a higher flow of air between the rooms, creating a better distribution of the heat. In the horizontal unit, air entered through almost the whole cross-sectional area of the unit and barely any air in the exhaust air room was exchanged with the adjacent room. The vertical unit was beneficial for the exhaust air rooms and,
7
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Fig. 7. A plane plot showing (a) the PPD value, (b) the air temperature and (c) the air velocities, with Case 3.1 to the left and Case 3.2 to the right.
necessary. While this study is focused on the thermal comfort, the produced heat from occupants have been left out of the simulations in order to simplify the simulations and design for a building without internal gains. The main reason is due to no occupants living in the building during measurement and validation period. Any future work will have to include occupants in some form to better investigate convection flows, and more importantly, the added heat source from the occupants. With an AH system that can regulate itself the air supply temperature and/or mass flow must be lowered in order to not overheat the building when the occupants are present. The study provides guidelines on how to design an AH system in the sub-Arctic climate, and future study to improve and optimize the thermal indoor climate in this type of building is recommended to include the transfer units.
between Case 1.0 and Case 1.1 when it was recalculated. This also reduce the amount of energy required to heat the incoming air with 792 W under the circumstances of the study. This adjusted flow is an improvement for both comfort and energy that can be done to the actual semi without practical implications. Any further improvements can however prove challenging, like relocating inlets. Relocating two of the inlets improved the thermal indoor climate significantly but will require extensive retrofit of the ventilation system. The placement of the supply devices must be considered during the design of these types of buildings, as there can be implications to adjust it in a later stage. During the winter the relative humidity (RH) is decreased when the incoming air is heated. Measurements performed in the studied building show RH values around 25–30% during the winter. There is no recommended RH range, but Li et al. [35] investigated the humidity in different seasons and at different indoor temperatures to find the optimums for human comfort. In a cold-climate winter in China with an outdoor temperature average of 9.4 °C, these authors found that an absolute humidity optimum is between 2.9 and 12.8 g/kg at an indoor temperature between 23.2 °C and 28.6 °C. This corresponds to a span of RH values from 17 to 52%. The RH in the Seventh House should therefore be satisfactory. However, the ways in which different levels of RH affect the thermal comfort in the building have not been investigated in the present study. Sub-Arctic winters have a low absolute humidity, and this could possibly be a problem, but increasing the humidity in the supply air in an AH system is possible if proven
4. Conclusions The results indicate that it is possible to achieve a good thermal indoor climate with an AH system during the cold winter months in the sub-Arctic climate. The original case had discomfort in 98% of the occupied zone, which was reduced to 8.3% in the final test scenario. Besides having to comply with the normal regulations for the ventilation of buildings, the system design has to include the heat demand in order to balance the amount of air in the building to establish an even distribution of the heat. The supply air mass flow was adjusted and 8
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reduced with 15% for the building, which correspond to 0.8 kW in heating the supply air. The placement of inlets also has to be considered when designing the system. Rooms like bathrooms require some sort of heat source, for instance a towel heater, to make it possible to maintain a comfortable air temperature. Since exhaust air rooms have no direct heat source, they should be placed in such a way that the influence of the external wall surfaces is minimized, meaning that they should be placed in the middle of the house or so that they share a wall with other houses in the case of multi-house buildings. Laundry and storage rooms do not necessarily have to keep a perfect thermal indoor climate, since they are typically not used frequently. This is enough to create a satisfactory thermal indoor climate, but by adding transfer units the comfort can be increased further. The installation of air transfer units in exhaust air rooms leads to a better thermal balance. Traditional horizontal units above door frames and vertical units along the vertical door frame give a similar overall result. The vertical unit gives a higher exchange of air but can cause a draught near the transfer unit. Vertical units show slightly worse overall results but are beneficial for the exhaust air rooms. The horizontal unit mainly creates a flow of air into the room, which might not change the temperature in a satisfactory way, but the resulting flow and draught is kept outside the occupied zone. Transfer units or grilles are recommended regardless. The study also shows how to create a fairer assessment of the thermal comfort when using PPD and a whole building interior, when PPD is combined with time spent in the various rooms.
[5] D. Risberg, M. Risberg, L. Westerlund, Investigation of thermal indoor climate for a passive house in a sub-Arctic region using computational fluid dynamics, Indoor Built Environ. (2018), https://doi.org/10.1177/1420326X17753707 1420326X1775370. [6] Z. Wang, Y. Ji, J. Ren, Thermal adaptation in overheated residential buildings in severe cold area in China, Energy Build. 146 (2017) 322–332, https://doi.org/10. 1016/j.enbuild.2017.04.053. [7] M. Berge, H.M. Mathisen, Perceived and measured indoor climate conditions in high-performance residential buildings, Energy Build. 127 (2016) 1057–1073, https://doi.org/10.1016/j.enbuild.2016.06.061. [8] L. Georges, M. Berner, H.M. Mathisen, Air heating of passive houses in cold climates: investigation using detailed dynamic simulations, Build. Environ. 74 (2014) 1–12, https://doi.org/10.1016/j.buildenv.2013.12.020. [9] K.M. Luc, M. Kotol, T. Lading, Energy-efficient building in Greenland: investigation of the energy consumption and indoor climate, Procedia Eng. 146 (2016) 166–173, https://doi.org/10.1016/j.proeng.2016.06.368. [10] Y. Ji, Z. Wang, H. Ning, X. Zhang, Field investigation on indoor thermal environment at a rural passive solar house in severe cold area of China, Procedia Eng. 121 (2015) 596–603, https://doi.org/10.1016/j.proeng.2015.08.1044. [11] K. Horikiri, Y. Yao, J. Yao, Numerical optimisation of thermal comfort improvement for indoor environment with occupants and furniture, Energy Build. 88 (2015) 303–315, https://doi.org/10.1016/j.enbuild.2014.12.015. [12] D. Risberg, M. Risberg, L. Westerlund, The Impact of Snow and Soil Freezing for Commonly Used Foundation Types in a Subarctic Climate, Energy Build, 2018, https://doi.org/10.1016/j.enbuild.2018.05.049. [13] A. Hesaraki, J.A. Myhren, S. Holmberg, Influence of different ventilation levels on indoor air quality and energy savings: a case study of a single-family house, Sustain. Cities Soc. 19 (2015) 165–172, https://doi.org/10.1016/j.scs.2015.08.004. [14] Boverket, Regelsamling För Byggande, BBR, first ed., Karlskrona, 2015 [Swedish Only]. [15] P.V. Nielsen, Fifty years of CFD for room air distribution, Build. Environ. 91 (2015) 78–90, https://doi.org/10.1016/j.buildenv.2015.02.035. [16] D. Risberg, L. Westerlund, J.G.I. Hellström, CFD-simulation of indoor climate in low energy buildings computational setup, Therm. Sci. (2015) 167, https://doi.org/10. 2298/TSCI150604167R 167. [17] V. Vitale, G. Salerno, A Numerical Prediction of the Passive Cooling Effects on Thermal Comfort for a Historical Building in Rome, Energy Build, 2016, pp. 1–10, https://doi.org/10.1016/j.enbuild.2017.06.049. [18] P. Rohdin, A. Molin, B. Moshfegh, Experiences from nine passive houses in Sweden indoor thermal environment and energy use, Build. Environ. 71 (2014) 176–185, https://doi.org/10.1016/j.buildenv.2013.09.017. [19] C.G. Popovici, HVAC system functionality simulation using ANSYS-fluent, Energy Procedia 112 (2017) 360–365, https://doi.org/10.1016/j.egypro.2017.03.1067. [20] D.G. Leo Samuel, S.M.S. Nagendra, M.P. Maiya, Simulation of indoor comfort level in a building cooled by a cooling tower-concrete core cooling system under hotsemiarid climatic conditions, Indoor Built Environ. 26 (2017) 680–693, https://doi. org/10.1177/1420326X16635260. [21] P.O. FANGER, Thermal Comfort. Analysis and Applications in Environmental Engineering, Danish Technical Press., Copenhagen, 1970. [22] CEN, EN 16798-3, (2015). [23] SIS, SS-EN 16798-3:2018 vol. 1, (2018). [24] I. Khajehzadeh, B. Vale, N. Isaacs, Time-use in different rooms of selected New Zealand houses and the influence of plan layout, Indoor Built Environ. 27 (2018) 19–33, https://doi.org/10.1177/1420326X16665161. [25] M. Magnusson, S. Kindell, S. Malmsten, M. Asp, S. Bergström, L. Wern, Lokalklimat Kiruna, (2012) [Swedish Only]. [26] SS-EN ISO 13789:2007, second ed., (2008) Stockholm. [27] FEBY18, Kravspecifikation För Energieffektiva Byggnader Bostäder Och Lokaler, (2018) [Swedish Only]. [28] S. Dehlin, 7nde Huset – Lågenergiparhus I Kiruna, (2017) [Swedish Only]. [29] SS-EN ISO 15927-5, first ed., (2011) Stockholm. [30] H.B. Awbi, A. Hatton, Natural convection from heated room surfaces, Energy Build. 30 (1999) 233–244, https://doi.org/10.1016/S0098-8472(99)00063-5. [31] D. Risberg, M. Risberg, L. Westerlund, CFD modelling of radiators in buildings with user-defined wall functions, Appl. Therm. Eng. 94 (2016) 266–273, https://doi.org/ 10.1016/j.applthermaleng.2015.10.134. [32] G. Paliaga, L.J. Schoen, P.F. Alspach, E. a Arens, R.M. Aynsley, R. Bean, J. Eddy, T.B. Hartman, D. Int-hout, M. a Humphreys, E.E. Khalil, M.P.O. Rourke, P. Simmonds, J.L. Stoops, S.C. Turner, K.W. Cooper, W.F. Walter, D.S. Abramson, C.S. Barnaby, S.J. Emmerich, J.M. Ferguson, R.L. Hall, R.M. Harrold, Thermal environmental conditions for human occupancy, Ashrae. ASHRAE Sta 58 (2013) ISSN 1041-2336. [33] ISO, ISO 7730:2005, (2005). [34] ANSI ASHRAE Standard 55-2013, (2013). [35] C. Li, H. Liu, B. Li, A. Sheng, Seasonal effect of humidity on human comfort in a hot summer/cold winter zone in China, Indoor Built Environ. 0 (2018) 1–14, https:// doi.org/10.1177/1420326X17751594.
Funding This work was supported by the EU program Interreg Nord, Region Norrbotten and Luleå University of Technology. Declaration of interest Declarations of interest: none. Acknowledgement Cooperation has taken place with NCC and our academic partners at Luleå University of Technology, who made this research possible. Appendix A. Supplementary data Supplementary data to this article can be found online at https:// doi.org/10.1016/j.buildenv.2019.106164. References [1] W. Feist, J. Schnieders, V. Dorer, A. Haas, Re-inventing air heating: convenient and comfortable within the frame of the Passive House concept, Energy Build. 37 (2005) 1186–1203, https://doi.org/10.1016/j.enbuild.2005.06.020. [2] C. Isaksson, F. Karlsson, Indoor climate in low-energy houses-an interdisciplinary investigation, Build. Environ. 41 (2006) 1678–1690, https://doi.org/10.1016/j. buildenv.2005.06.022. [3] D. Risberg, M. Vesterlund, L. Westerlund, J. Dahl, CFD simulation and evaluation of different heating systems installed in low energy building located in sub-arctic climate, Build. Environ. 89 (2015) 160–169, https://doi.org/10.1016/j.buildenv. 2015.02.024. [4] M. Gustafsson, G. Dermentzis, J.A. Myhren, C. Bales, F. Ochs, S. Holmberg, W. Feist, Energy performance comparison of three innovative HVAC systems for renovation through dynamic simulation, Energy Build. 82 (2014) 512–519, https://doi.org/10. 1016/j.enbuild.2014.07.059.
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Air heating system design for a sub-Arctic climate using a CFD technique ∗
T
Petter Lundqvist , Mikael Risberg, Lars Westerlund Division of Energy Science, Department of Engineering Sciences and Mathematics, Lulea University of Technology, Lulea, Sweden
A R T I C LE I N FO
A B S T R A C T
Keywords: Air heating CFD Cold climate PPD Sub-Arctic climate Thermal comfort
The thermal comfort in a residential building equipped with an air heating system and located in a sub-Arctic region was investigated with computational fluid dynamics (CFD) software. The predicted percentage of dissatisfied (PPD) was used to identify flaws with the heating system during winter conditions. New scenarios were simulated and compared to each other to see potential improvements of the thermal indoor climate. Comparison was done by combining the discomfort spaces inside rooms, the level of the discomfort and the time spent in these spaces. The discomfort covered 8–38% of the interior volume depending on the test case. The results provide the necessary means to create a satisfactory thermal indoor climate if an air heating system is to be utilized in sub-Arctic regions during the winter. The correct heat demand for each floor and appropriate placement of the supply devices are required. Adding air transfer units or grilles in rooms from which exhaust air is removed further improves the comfort. The results also show the strength of using CFD technique when investigating the indoor discomfort with PPD, and how a fair assessment can be done by combining the PPD with time.
1. Introduction As a result of low energy buildings becoming standardized as residential buildings in cold climate regions, the required heat demand is reduced. It is established that an improved building envelope changes the air flow in buildings due to less infiltration and convection flows. Mechanical ventilation is therefore commonly required when natural ventilation is inadequate for supplying air, which in turn also affects the thermal indoor climate. Mechanical ventilation gives the possibility of heating the air and using it as a heat source, which is a feasible approach that has been utilized in passive houses in Europe [1]. Usage of heated air instead of radiators or other hydronic heating systems in residential buildings during the cold months in southern Sweden has been studied, and although risk of overheating has been observed, the approach has been considered feasible [2]. Risberg et al. [3] investigated different heating systems in the subArctic climate, and found that an air heating (AH) could create a good thermal comfort. The AH system in that study was similar to the heating system investigated by Gustafsson et al. [4], who considered it to be the best option for a well-insulated house in a cold climate. Risberg et al. [5] continued to investigated the AH system in a pilot passive house in the sub-Arctic region of northern Sweden, and showed that the AH was feasible from an energy point of view. If an AH system is utilized in a
passive house, there is no need for the installation of pipes and radiators, and thus an AH system can be less expensive compared to underfloor heating or a hydronic radiator system [3], which are integral parts of the traditional heating system in Sweden. However, the findings so far regarding the AH system in sub-Arctic region point out problems with the thermal comfort, which was not satisfactory due to mainly overheating in the cold winter months. The phenomenon of overheating during winter has been seen in other passive buildings in climates ranging from cold to Arctic climates [6–9]. Ji et al. [10] noted overheating in a passive house during the winter in a cold region of China when performing research with an outdoor temperature of −23.4 °C. The indoor measurements exhibited temperatures above 24 °C, which were caused by added solar gains, the occupants and indoor apparatus. This implies the heating system was not properly designed. Horikiri et al. [11] showed that occupants can have a significant impact on the thermal indoor climate, however, the heating system should be designed to withstand cold days with no internal gains such as the occupants. Both internal and solar gains can be left out completely when designing the AH system in passive houses [8]. The system should be able to regulate itself when necessary, i.e. when occupants are present. In the same study by Horikiri et al. [11], it was also concluded that furniture had an influence on the local air velocities around
∗ Corresponding author. Luleå University of Technology, Division of Energy Science, Department of Engineering Sciences and Mathematics, SE-971 87, Luleå, Sweden. E-mail addresses: [email protected] (P. Lundqvist), [email protected] (M. Risberg), [email protected] (L. Westerlund).
https://doi.org/10.1016/j.buildenv.2019.106164 Received 4 April 2019; Received in revised form 13 May 2019; Accepted 24 May 2019 Available online 25 May 2019 0360-1323/ © 2019 Elsevier Ltd. All rights reserved.
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possible visualize detailed air flows, temperature gradients and also investigate the Predicted mean Vote (PMV) and Predicted Percentage Dissatisfied (PPD), as described by Fanger [21]. However, while the PMV and PPD has been visualized with CFD software, most studies focus on planes as a method of investigation and presentation. One of the strengths with CFD software is to simulate a whole building interior, which gives the possibility to investigate the PMV and PPD in the form of a volume so that it gives a better understanding of the whole thermal comfort simultaneously, rather than a plane. The PMV and PPD consider air humidity, air temperature, air velocity, mean radiant temperature, clothing thermal insulation and metabolic rate. In order to make a more correct evaluation of the thermal indoor climate, two other aspects could also be considered, especially if a whole building interior is investigated. The first aspect is where occupants dwell, in other words, the occupied zone. Definitions of the occupied zone can vary but are in general the zone between 0.05 and 1.80 m on horizontal level, 1.00 m from windows and doors, and 0.50 from both internal and external walls [22]. According to Swedish regulations the occupied zone is the zone between 0.10 m and 2.00 m above the floor level, 1.00 from windows and doors, and 0.6 m from the external walls [23]. The second aspect is the amount of time people spend in each room, which should be considered and weighed in. A lower performance of the thermal indoor climate in the laundry room, for example, can have a significant impact on the total PMV and PPD if the whole apartment is being considered. In reality, however, this is not the case, as the occupants will spend more time in bedrooms and living rooms, as shown by Khajehzadeh et al. [24]. Based on the same pilot passive house as in the previous study, the main objective of this study was to establish that it is possible to achieve a satisfactory thermal indoor climate using an AH system in a sub-Arctic climate. This type of heating system can reduce the construction cost by removing the need to install a hydronic system. A secondary objective is to show how PPD can be combined with time to make a fairer assessment of the thermal indoor climate. The investigation of the thermal indoor climate was performed with CFD software. The current settings of the AH system were analyzed by rendering the PPD as volumes in the building model which was validated with experimental data. Using the PPD volumes, it was possible to identify the flaws associated with the current settings, placement of the supply and exhaust devices and the air transfer inside the building. New solutions were proposed, simulated and compared with each other to see changes in the thermal comfort. The results obtained in this study can in turn provide guidance for the future design and use of AH systems in sub-Arctic regions, as well as provide directions of how to
Abbreviations AH semi DWOT UDFs COM
Air heating Semi-detached house Design winter outdoor temperature User-defined functions Comparison number
the furniture, but the buoyancy strength was not affected significantly. Heat generated by electric appliances such as TVs can be neglected when the thermal indoor climate is simulated [11]. In the sub-Arctic regions of Sweden snow is present during the cold winter months. The snow surrounding a passive house has an insulation effect on the foundation, but for a passive house with a thick insulation, the added effect from snow is not significant enough to be considered during simulations and design of an AH system [12]. Georges et al. [8] investigated AH systems in a cold and sub-Arctic climate in Norway. They used the TRNSYS software and concluded that if AH was to be used, there would be some limitations and certain conditions that would have to be addressed for sub-Arctic climates. Something else to consider is that the amount of heat air also must meet the minimum requirement for supplied air. Although it has been shown that an air flow of 0.2 l/(s m2) of floor area can be sufficient [13], the Swedish requirement for the air change rate is generally 0.35 l/(s m2) of floor area [14], which gives a lower boundary for the amount of supplied air. As mentioned, in the previous study by Risberg et al. [5] from the northern Sweden, the thermal indoor climate was not satisfactory in the whole building. One of the flaws of the AH system in the building was the excessively high mass flow of air on the first floor, which led to the overheating of part of the building, while the bathrooms and laundry room were too cold. The uncertainty associated with the design, installation and operation of an AH system in sub-Arctic regions remains. Risberg et al. [5] therefore suggested that the AH system should be studied further to detect problems and suggested new settings to improve the thermal indoor climate. While software such as TRNSYS can give good estimations of the thermal comfort in a building, it does not give a full understanding of the thermal indoor climate, in contrast to computational fluid dynamics (CFD) simulations, which can provide detailed results for the indoor climate throughout the whole building. CFD as a tool for thermal indoor climate simulations has been shown to be able to provide a detailed view of the thermal indoor climate [10,11,15–20]. With CFD it is
Fig. 1. (a) Top-view of the building location showing the north-south direction. The bottom half of the building is the southern semi, while the top half is the northern semi. (b) Front view of the building, with the southern semi to the left. 2
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compare different thermal indoor climates using PPD.
Table 1 U-values for different parts of the building.
2. Methods 2.1. The building The building is located in Kiruna just south of latitude 68°N, built to withstand polar nights throughout the winter, along with the possibility of severe cold between −30 °C and −40 °C [25]. The building was built as a pilot case test building to test the AH system among other low energy techniques. The building is made up of two semi-detached houses (semi) and has a concrete foundation; the front of the building faces roughly east-south-east, as seen in Fig. 1 (a). A front view of the building is provided in Fig. 1 (b) to show the two semis. The total area for living which has to maintain a satisfactory indoor climate in each semi is 134 m2. The layout of each semi is a mirror image of that of the other semi. The ground floor contains one bedroom, living room, kitchen, bathroom and laundry room, as well as an entrance hall with a staircase to the first floor. The first floor contains three bedrooms, a family room, a bathroom and a storage room. Due to the semis mirror each other, it is only necessary to perform the investigation on one of them. The semi focused on in this study is the southern one, and a layout of this house can be seen in Fig. 2. The cardinal direction did not have to be considered since the investigation was performed in a period of polar night and sun did not affect the building. The heated air is supplied to the bedrooms, living room and family room via circular supply devices which are 125 mm in diameter and are mounted in the ceiling. The supply devices will henceforth be referred to as inlets. For the exhaust air, the kitchen, bathrooms and laundry room have circular outlets which are 100 mm in diameter and are located in the ceiling. These exhaust air devices will henceforth be referred to as outlets. The placement of the inlets and outlets on the two floors is illustrated in Fig. 2, where the inlets are denoted by circles and the outlets by crosses. Each bathroom is equipped with an electric towel heater as a heat source, a 50 W towel heater for the ground floor and a 102 W towel heater for the first floor, and these heaters should in theory be able to maintain a comfortable room temperature. The towel heaters are denoted by rectangles in the bathrooms in Fig. 2.
Section
U-value [W/(m2∙K)]
Envelope area [m2]
Roof External walls, ground floor External walls, first floor Foundation Windows, ground floor Windows, first floor Front door
0.035 0.083 0.083 0.139 0.650 0.650 0.700
67.1 53.2 46.1 67.3 11.0 12.5 2.10
first floor because this floor has more rooms with supply air [5]. To sustain the thermal indoor climate during winter conditions, a supply air temperature of 45 °C was needed, and this is the temperature used in this study. Based on the heating demand, new mass flows for each inlet were calculated separately for the two floors. EN ISO 13789:2007 and the Swedish regulation FEBY18 were used to determine the heat loss from the building [26,27]. Concerning the envelope area, the U-values for the external walls, roof, foundation, windows and front door were based on the blueprints and are presented in Table 1. The total infiltration was measured to be 87 l/s at a pressure of 50 Pa between the indoor and outdoor conditions, done accordingly to standard EN-13829 [28] with a blower door. In normal conditions this correspond to approximately 0.02 l/(s m2) interior envelope area. A wind coefficient of 0.7 was used and this value corresponds to a moderate shielding from the wind, such as that typical of a suburban environment. It was assumed that the infiltration rate was the same on both floors. The heat transfer coefficient for the external walls was recalculated and changed from 0.079 to 0.083 W/(m2∙K) to take the energy loss from infiltration into consideration, and the latter value was used as the overall heat transfer coefficient for the walls. The mass loss of air due to the infiltration corresponded to less than a permille of the indoor air, and therefore assumed to not have a significant impact on the indoor air, and therefore not included in the CFD model. Based on SS-EN ISO 15927-5, a seven-day value for the design winter outdoor temperature (DWOT) was used, giving an outdoor air temperature of −25.7 °C, and the ground temperature beneath the building was set to −7.3 °C [29]. For the calculations of heat demand, an indoor temperature of 21 °C was used in all the rooms except the bathrooms, where 25 °C was used. Other data used had been gathered by NCC AB Sweden through investigation of the building [28]. The total heat demand for the ground floor in these conditions was manually calculated to be 897 W, and the corresponding demand for the first
2.2. Heating demand and required air flows The previous study of the building has shown that the original design of the AH system allocates an excessive amount of air flow to the
Fig. 2. Location of the inlets (circles), outlets (crosses) and towel heaters in bathrooms (rectangles – not according to scale). The dashed line represents a plane that is used later for presentation of results. 3
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floor was likewise calculated to be 796 W. The supplied air for each floor was calculated as sensible heat. The rooms with supply air will have higher air flows than those actually needed, because the heat is not supplied to all the rooms, and rooms with no supply air relied on warm air from other rooms. Table 2 presents the air flows for each room, both according to the original design and according to the heat demand on each floor. The newly calculated supply air for the whole semi corresponds to an air change rate of 0.43 l/(s m2) of floor area. The new total supply air mass flow was 0.064 kg/s for the semi, with 55% distributed to the ground floor and 45% distributed to the first floor. In the original design, a total supply of 0.075 kg/s was used, with 37% distributed to the ground floor and 63% distributed to the first floor [5]. Since the total supply air flow was 15% lower than that in the original design case, the exhaust flow was lowered by 15% for each outlet. In the original design case, no towel heaters were implemented. For comparison the original design case was simulated with the towel heaters included. To match the heat demand, the temperature was set to 44 °C.
applied. The criteria used to reach a solution were the stabilized mass, heat transfer and turbulence residuals, along with the stable temperatures in the center of each room at a height of 1.1 m. The external surfaces were assumed to have a uniformly heat loss corresponding to their respectively heat transfer coefficients in Table 1 as boundary conditions. As previously stated, the energy loss due to infiltration was included for the external walls. Mass transport of the air through the walls was not included in the CFD model since it was assumed that it did not have any significant impact on the indoor air as previously stated. The internal walls and intermediate floors in the model were set as solids, with a thermal conductivity of 0.05 W/(m · K). The internal wall shared with the northern semi was set as adiabatic. Supply devices were created as inlets and exhaust devices as outlets. Each inlet and outlet were given the mass flow according to Table 2. To be able to investigate PPD with CFX-Post, a PPD variable was created that was connected to an expression that calculated the PPD [5]. The PPD comfort parameter was calculated according to ASHRAE Standard 55–2010 [32]. The cloth factor was set to 1 Clo, while the metabolic rate was set to 58.2 W/m2 and humans were considered to be 1.7 m2 to correspond to the previous study [5]. The pressure for the humidity was set to 600 Pa to correspond to measured relative humidity during the validation period. Draught rate was also created as a variable, linked to an expression, based on the ISO 7730 standard [33].
2.3. Computational fluid dynamics model The ANSYS CFX 19.2 software was used to solve the numerical model. The post-processing of the results was performed in CFX-Post. The geometry used contained the whole interior of the southern semi as one volume. The glazed porch at the front door of the real building was excluded, as this section does not contain any heating. The temperature in this part was considered to be the same as the outdoor temperature. The original validated CFD [5] model of the building was used, including the mesh that was produced accordingly to a previous grid size study, which also established the computational setup for the model [16]. The model was simplified by removing the staircase and appliances in the bathrooms and the kitchen. The settings used for the mesh include having the grid size maximum cell face length of 0.1 m, with inflation layers applied to all the zones delimiting the room with a growth rate of 1.5. For air transfer units or grilles between rooms the elemental size was set to 0.01 m. For smaller volumes between closed doors and floors (slits), the element size was set to 0.01 m on horizontal surfaces and 0.001 m on vertical surfaces. The total amount of elements in the model varied from 4.5 million to 4.9 million, because of different features being added later on. Because a worst-case scenario was used, based on SS-EN ISO 159275 [29], and thermal comfort by using PMV/PPD was to be investigated, only steady state solution was calculated. The simulation was set up according to the previously validated simulation [5], which was validated during polar night period of December 16th. This resulted in 0.00 W/m2 from solar radiation, confirmed by measurement data from the Swedish Meteorological and Hydrological Institute for the validated time period. Additional heat transfer due to sky radiation and longwave radiation from nearby buildings was calculated but did not have any significant impact on the simulations for this case. A second order upwind discretization scheme with RMS residuals set to 10−6 was used for the simulations. The fluid model used a total energy model and a standard k-epsilon turbulence model, and the discrete transfer model was used for the thermal radiation. The thermal radiation for internal surfaces had an emissivity of 0.9, while the windows had an emissivity set to 0.83. The boundary conditions for internal convection were defined with user-defined functions (UDFs), which were written based on the correlations from Awbi and Hatton [30] and were created based on the fact that heat was distributed along the ceiling in this case. The use of UDFs with the k-epsilon model have provided reliable results with relative fast simulation times [31]. The initial indoor air temperature of the model was set to 21 °C, while the gauge pressure and momentum in all directions were set to 0 Pa and 0 m/s, respectively. Due to CFX using a pseudo time step for solving steady state simulations a timescale is needed, and in order to reach convergence faster with CFX, a physical timescale of 1000 s was
2.4. Improvements of the indoor climate In order to improve the thermal indoor climate, new scenarios were simulated. The scenario according to the original design and how the house was built was named Case 1.0. Another four cases were tested and are presented in Table 3, to give an overview of the cases. Case 1.0. had been validated previously through continuous measurements of the air temperature and relative humidity in each room, supplemented with manual measurements of the air velocities [5]. It should be noted, however, that the previous study did not include towel heaters in the bathrooms. Along with the towel heaters and the recalculated heat demand for the seven-day DWOT, the original case required a supply air temperature of 44 °C, which was used for Case 1.0 to compare the original design with new scenarios. Case 1.1. introduced the rearranged mass flow of supplied air on each floor according to the heat demand for the respective floors. It was expected that the adjusted heat demand would not be sufficient to create a satisfactory thermal indoor climate. The previous study showed problems with heat concentrating in some regions with subsequent overheating. At the same time, some areas had a deficit of heat. On the ground floor, all the inlets are placed in one side of the semi, which creates an uneven distribution of heat. The placements of the inlets on the first floor were also expected to create an uneven distribution. Table 2 Supply/exhaust air flows for each room.
4
Location
Floor
Mass flow Supply/exhaust Original design [g/s]
Mass flow Supply/exhaust Corrected for heat demand for each floor [g/s]
Bedroom 1 Living room Bedroom 2 Bedroom 3 Bedroom 4 Family room Bathroom 1 Kitchen Laundry room Bathroom 2 Total
Ground floor Ground floor First floor First floor First floor First floor Ground floor Ground floor Ground floor First floor
12.5/15.0/12.5/12.5/14.0/8.50/-/15.0 Open -/15.0 -/25.0 75.0
8.15/13.5/6.71/3.81/5.88/12.4/-/12.8 Open -/12.8 -/21.3 63.9
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Table 3 Scenarios simulated.
Table 4 Time spent in each room.
Case
Settings
1.0 1.1 2.1 3.1 3.2
Original design Changed air mass flow/floor Case 1.1 + New placement of inlets Case 2.1 + Horizontal units Case 2.1 + Vertical units
Bedroom 1 Bedroom 2 Bedroom 3 Bedroom 4 Kitchen Living room Family room Entrance hall Bathroom Laundry room Storage room
Case 2.1. was created from Case 1.1 to avoid the heat concentrating in some areas by moving two inlets. Case 2.1 kept the same mass flow rate of air through each inlet. One of the inlets in the living room was moved to the entrance hall and the inlet in the family room was moved to a position above the stairs, according to Fig. 3. Because of the doors to the bathrooms and the laundry room remain closed in the simulations, which also is the case in practice most of the time, the air has difficulty in flowing into these rooms and can only enter through small slits under the door. Two different types of air transfer units, or transfer grilles, were added to Case 2.1 and were placed in the internal walls of the bathrooms and laundry room, creating Case 3.1 and Case 3.2, respectively.
Hours
% of total time
2.27 2.27 2.27 2.27 2.89 2.89 0.77 0.28 0.78 0.11 0.01
14 14 14 14 17 17 4.6 0.8 4.6 0.7 0.06
investigations, a distance of 0.6 m from the external walls was used for the windows and doors, and a distance of 0.1 m was added for the internal walls. The volumes were colored according to a PPD scale from 10% to 100%, to clarify further how severe the problems were. To make a fairer assessment, the time per day spent in each room was included as well. It is more important to have a good climate in a room where the occupants spend a large amount of time compared to a room where the occupants only spend a short time. In order to compare the different air distribution arrangements for the different cases, a comparison number (COM) was created, calculated as follows,
Case 3.1. used the more traditional approach with horizontal transfer units with the dimensions (W*H) 700 × 100 mm and centered 150 mm above the door frame. Case 3.2 was provided with vertical transfer units along the door frame. The same cross-section area was kept for these units, but the dimensions were changed to (W*H) 39 × 1800 mm and the units were placed 40 mm from each door opening and 100 mm above the floor level. By keeping the same cross-section area, the two types could be compared. A high narrow opening along the vertical door frame was expected to increase the air exchange between the rooms due to convection flows. All of these units were made as part of the interior volume between the rooms to give an overview of the flow behavior and to determine whether the thermal comfort could be increased with the help of the transfer units.
COMj =
∑
(PPDj (%)⋅VOLj (%)⋅Timej (%)).
For each room in the semi, the average PPD value above 10% in the occupied zone (PPDj(%)) was multiplied by the volume where the PPD value was above 10% (VOLj(%)) and the amount of time spent in that room (Timej (%) ), based on the numbers given in Table 4. The results for all the rooms were then summed up to create the COM value for each case. The comparison number does not include the fraction which each room volume constitutes of the total volume of the semi, since the time spent in the respective rooms is included. The COM value is only used as a reference to make it possible to compare the improvement of one case with that of another. The numbers in Table 4 was based on a study by Khajehzadeh et al. [24]. Although this study was performed in New Zealand, similar results can be expected in Sweden. It was assumed in the present study that, in total, 16.8 h per day were spent at home on average (calculated on an annual basis) for the setup consisting of a separated living room, kitchen and dining room. The family room was considered as a room used for a combination of play, study and games. Moreover, it was
2.5. Evaluation of the scenarios According to ASHRAE Standard 55–2010, an acceptable indoor climate when using PPD is a value up to 10% [34]. To detect problems in the semi more easily, the PPD variable was plotted for each element and presented as volumes when the PPD value was greater than 10% within the occupied zone. To simplify the occupied zone for PPD
Fig. 3. New position of two inlets. 5
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room, but had an average value of 23 °C. The radiant temperature reached 28 °C in the living room, and together the air temperature and the radiant temperature created dissatisfaction in almost all the occupied zone. This shows that the amount of air supplied to the living room was excessive due to the placement of the inlets. Out of the 897 W needed for the ground floor, 651 W were supplied to the living room, representing 73% of the ground-floor heat demand. Case 1.1 clearly shows that air was not properly supplied to the semi. The heat demand is the same throughout the semi, but on the ground floor the air was supplied in one side of the semi, which created a thermal imbalance. In Case 2.1, two of the inlets were moved to new positions, decreasing the COM value to 2.7, which is a significant improvement compared to Case 1.1. The problematic volumes can be seen in Fig. 6, where the large volume in the living room has almost disappeared. In Case 2.1 the volumes with PPD values above 10% covered in total 22% of the occupied zone, with an average PPD value of 16%. A comparison of the results for Case 1.1 and Case 2.1 clearly shows that the heat was spread more evenly in the latter case. The average temperature difference between each floor in Case 1.1 was 1.1 °C, with the ground floor being warmer. In Case 2.1 the corresponding difference was 0.2 °C, with the ground-floor temperature decreasing and the firstfloor temperature increasing. After one of the inlets was moved from the living room, the heat was no longer concentrated in that part of the ground floor. However, the inlets in the entrance hall and above the staircase created dissatisfaction instead. The air supplied at the groundfloor entrance moved up to the first floor and the heat accumulated together with the air supplied on the first floor. The plume of hot air moving up to the first floor contributed to a flow of air going down along the staircase. Consequently, there was a significant increase in the PPD value of the occupied zone at the stairs. While this affected the results, this specific area was mainly on the first floor above the groundfloor staircase part, which in reality would not affect the thermal comfort. The problem in the laundry room remained, since this room showed no improvement, even though an inlet was located outside the room. The local velocities along the floor remained high, and local velocities of 0.5 m/s appeared in all the exhaust air rooms. However, the use of transfer units in Case 3.1 and 3.2 resulted in improvements for these rooms. The COM value decreased to 0.8 and 1.3 in Case 3.1 and 3.2, respectively, as seen in Fig. 4. In Case 3.1, the PPD volume covered 8% of the occupied zone, with an average PPD value of 15%. In Case 3.2, the PPD value covered 10% of the occupied zone, with an average PPD value of 15%. In order to see the effect of the transfer units, especially the vertical one in Case 3.2, a plane was created 1.68 m from the entrance door, visible as a dashed line in Fig. 2. The PPD value, air temperature and velocity were investigated separately, and are presented in Fig. 7. From Fig. 7 (a), one can observe that while the PPD values were at low levels of 5–10% in most of the plane, the laundry room still had some problems, with the average PPD value remaining high at 39% and 40% for Case 3.1 and Case 3.2, respectively. The increased velocity does not increase the PPD which is due to an increase of the temperature with fresh air from the inlet. By comparing the laundry room with the first-floor bathroom, which both have the same layout and external area, it is possible to see the benefit of the towel heater, which is helping the AH system to create a better thermal comfort. Both of the
assumed that the amount of time spent in bedrooms was 9 h, and this time was divided by the number of bedrooms because the comparison was based on a summation of the time spent in all the rooms in the semi. The total time spent in the kitchen and living room was divided into two equal parts. 3. Results and discussion The COM value result for each case is presented in Fig. 4. A lower COM value indicates better thermal comfort. With the original design in Case 1.0, the COM value was 30.4. This value was decreased in the following cases when the air flow was adjusted to match the heat demand and there was a better distribution of the supplied air on each floor, implying a better thermal comfort. The results regarding Case 1.0 are not discussed in detail in this paper as this case was covered by Risberg et al. [5]. All the other cases were investigated separately in detail, which includes pressure fields, PPD, draught and temperatures. For each case, a visualization was created of the volumes inside the occupied zone where the PPD value was above 10%. This was done by plot the created PPD variable for each element within the occupied zone. An example of such a visualization is provided in Fig. 5, which shows the volumes for Case 1.1, where the original design of the inlets was kept and there were no added features to help distribute the air flow. The volumes visualized in the blue-to-red scale have a PPD value above 10%. Even if the result was not satisfactory, the COM value showed a remarkable improvement in the thermal indoor climate compared to the original design, Case 1.0, in that the COM value was decreased almost by a factor of four. In Case 1.1 the PPD volumes covered in total 38% of the occupied zone and had an average PPD value of 19%. The rooms which can be seen in Fig. 5 (a) to have had major problems in Case 1.1 are the exhaust air rooms along with the storage room. In Fig. 5 (b), with the building shown from a reverse angle, one can see that almost all the occupied zone of the living room and bedroom 1 was covered by PPD volumes. This clearly indicates the problem of an uneven distribution of air on the ground floor. Concerning the ground-floor bathroom, dissatisfaction was shown within 75% of the occupied zone, with an average PPD value of 26%. The first-floor bathroom results showed that 65% of the occupied zone was covered by PPD volumes, with an average PPD value of 25%. Both the air and the radiant temperatures were higher inside the PPD volume than outside the PPD volume, while the air velocities remained at similar value. This implies that the main problem in the bathrooms and living room is overheating. The radiant temperature in the bathrooms on both floors reached 26 °C, but, combined with the air temperature and air velocities, the operating temperature was 24 °C. For the bathrooms this was actually not a real problem due to these rooms are designed for an increased temperature. The laundry and storage rooms were too cold. Because the outlets are located in rooms with closed doors and air was entering through the slit below the door, a draught was created along the floor, which also increased the dissatisfaction. The local velocities reached 0.5 m/s in the bathrooms. The first floor had an average temperature of 21 °C, while the ground floor had an average temperature of 22 °C. The laundry and storage rooms had excessively low values both for the air and the radiant temperature, with temperatures of 18 and 19 °C, respectively. The storage room has no outlet and only exchange air flowing through the slit beneath the door. Both these rooms had an unsatisfactory thermal comfort, but this did not necessarily cause problems due to the rooms’ functions. When calculating the COM value, these rooms were basically neglected due to the time factor. The storage room will therefore not be discussed in detail. The laundry room is still of interest as it is one of the rooms which has an outlet but does not contain any heat source. In this room the whole occupied zone is problematic, with an average PPD value of 41%. The highest air temperature was 25 °C and was found in the living
Fig. 4. Comparison numbers for the different cases. 6
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Fig. 5. Volumes with PPD values above 10% in the occupied zone for Case 1.1: (a) view towards the gable and front of the semi, and (b) view towards the wall shared with the other semi and the back wall.
Fig. 6. PPD volumes above 10% in the occupied zone for Case 2.1: (a) view towards the gable and front of the semi, and (b) view towards the back and the section bordering on the adjacent semi.
overall, the velocities were lower in Case 3.2 than in Case 3.1. The transfer units used in the study are only represented as part of the interior volume, whereas in reality some sort of plate or grid are mounted on the units. This will affect the air flow compared to an opening in the wall, and any further study or optimization of this system also have to consider this. In all of the cases the kitchen had small PPD volumes. The largest PPD volume observed in the kitchen, with the original case excluded in the comparison, was for Case 1.1, where the PPD volume covered 9% of the occupied zone. In all the other cases it was lower than 3%. The kitchen has an exhaust air outlet only, but with two door openings the velocities were kept low and did not cause a draught. This type of flow is good for thermal comfort, and an open-plan living area where the kitchen and living room are combined could be preferable with an AH system. This would most likely create a better distribution of the air flow with no interior walls obstructing the flow. Of the other exhaust air rooms, the bathrooms were the ones most used, and these rooms required a higher air temperature than the rest of the semi. In Cases 3.1 and 3.2 in this study, the operating temperature for the bathrooms was around 24 °C, which from a thermal comfort point of view is good. Investigating the PPD in the manner selected for this study, with the Clo-parameter set to a static value of one, is not a perfect approach, and this has to be taken into consideration when analyzing the results. Correct air flows can also reduce energy use, if an excessive amount of air is supplied. In this study, the air mass flow was reduced with 15%
bathrooms are located at the gable corner, where the outdoor temperature will have the largest impact on the building. Without the towel heaters the temperature would drop below acceptable levels, even with the transfer units. This can be regarded as a major flaw of the AH system design, which forces these rooms to rely on an extra heat source. It is not unusual that bathrooms and laundry rooms are placed together at one end of the building, and not directly adjacent to bedrooms or living rooms. The problems connected with this have to be considered when designing an AH system. An AH system will always contribute to a temperature difference between rooms, with the supply air rooms always being warmer than the exhaust air rooms. The difference in temperature between these types of rooms decreases for Case 2.1 and even further for Case 3.1 and 3.2. To decrease this difference even further, the exhaust air rooms should ideally be placed in the middle of the building, so that it would be possible to supply heat to all the rooms with external walls. In the case of the studied building, the bedrooms should be located at the gable, and the bathrooms should share a wall with the adjacent semi. A comparison of the results obtained with the two different transfer units reveals that the horizontal traditional unit gave a better overall result with a lower COM value. Although it is not clearly shown in Fig. 7 (c), the vertical units produced a higher flow of air between the rooms, creating a better distribution of the heat. In the horizontal unit, air entered through almost the whole cross-sectional area of the unit and barely any air in the exhaust air room was exchanged with the adjacent room. The vertical unit was beneficial for the exhaust air rooms and,
7
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Fig. 7. A plane plot showing (a) the PPD value, (b) the air temperature and (c) the air velocities, with Case 3.1 to the left and Case 3.2 to the right.
necessary. While this study is focused on the thermal comfort, the produced heat from occupants have been left out of the simulations in order to simplify the simulations and design for a building without internal gains. The main reason is due to no occupants living in the building during measurement and validation period. Any future work will have to include occupants in some form to better investigate convection flows, and more importantly, the added heat source from the occupants. With an AH system that can regulate itself the air supply temperature and/or mass flow must be lowered in order to not overheat the building when the occupants are present. The study provides guidelines on how to design an AH system in the sub-Arctic climate, and future study to improve and optimize the thermal indoor climate in this type of building is recommended to include the transfer units.
between Case 1.0 and Case 1.1 when it was recalculated. This also reduce the amount of energy required to heat the incoming air with 792 W under the circumstances of the study. This adjusted flow is an improvement for both comfort and energy that can be done to the actual semi without practical implications. Any further improvements can however prove challenging, like relocating inlets. Relocating two of the inlets improved the thermal indoor climate significantly but will require extensive retrofit of the ventilation system. The placement of the supply devices must be considered during the design of these types of buildings, as there can be implications to adjust it in a later stage. During the winter the relative humidity (RH) is decreased when the incoming air is heated. Measurements performed in the studied building show RH values around 25–30% during the winter. There is no recommended RH range, but Li et al. [35] investigated the humidity in different seasons and at different indoor temperatures to find the optimums for human comfort. In a cold-climate winter in China with an outdoor temperature average of 9.4 °C, these authors found that an absolute humidity optimum is between 2.9 and 12.8 g/kg at an indoor temperature between 23.2 °C and 28.6 °C. This corresponds to a span of RH values from 17 to 52%. The RH in the Seventh House should therefore be satisfactory. However, the ways in which different levels of RH affect the thermal comfort in the building have not been investigated in the present study. Sub-Arctic winters have a low absolute humidity, and this could possibly be a problem, but increasing the humidity in the supply air in an AH system is possible if proven
4. Conclusions The results indicate that it is possible to achieve a good thermal indoor climate with an AH system during the cold winter months in the sub-Arctic climate. The original case had discomfort in 98% of the occupied zone, which was reduced to 8.3% in the final test scenario. Besides having to comply with the normal regulations for the ventilation of buildings, the system design has to include the heat demand in order to balance the amount of air in the building to establish an even distribution of the heat. The supply air mass flow was adjusted and 8
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reduced with 15% for the building, which correspond to 0.8 kW in heating the supply air. The placement of inlets also has to be considered when designing the system. Rooms like bathrooms require some sort of heat source, for instance a towel heater, to make it possible to maintain a comfortable air temperature. Since exhaust air rooms have no direct heat source, they should be placed in such a way that the influence of the external wall surfaces is minimized, meaning that they should be placed in the middle of the house or so that they share a wall with other houses in the case of multi-house buildings. Laundry and storage rooms do not necessarily have to keep a perfect thermal indoor climate, since they are typically not used frequently. This is enough to create a satisfactory thermal indoor climate, but by adding transfer units the comfort can be increased further. The installation of air transfer units in exhaust air rooms leads to a better thermal balance. Traditional horizontal units above door frames and vertical units along the vertical door frame give a similar overall result. The vertical unit gives a higher exchange of air but can cause a draught near the transfer unit. Vertical units show slightly worse overall results but are beneficial for the exhaust air rooms. The horizontal unit mainly creates a flow of air into the room, which might not change the temperature in a satisfactory way, but the resulting flow and draught is kept outside the occupied zone. Transfer units or grilles are recommended regardless. The study also shows how to create a fairer assessment of the thermal comfort when using PPD and a whole building interior, when PPD is combined with time spent in the various rooms.
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Funding This work was supported by the EU program Interreg Nord, Region Norrbotten and Luleå University of Technology. Declaration of interest Declarations of interest: none. Acknowledgement Cooperation has taken place with NCC and our academic partners at Luleå University of Technology, who made this research possible. Appendix A. Supplementary data Supplementary data to this article can be found online at https:// doi.org/10.1016/j.buildenv.2019.106164. References [1] W. Feist, J. Schnieders, V. Dorer, A. Haas, Re-inventing air heating: convenient and comfortable within the frame of the Passive House concept, Energy Build. 37 (2005) 1186–1203, https://doi.org/10.1016/j.enbuild.2005.06.020. [2] C. Isaksson, F. Karlsson, Indoor climate in low-energy houses-an interdisciplinary investigation, Build. Environ. 41 (2006) 1678–1690, https://doi.org/10.1016/j. buildenv.2005.06.022. [3] D. Risberg, M. Vesterlund, L. Westerlund, J. Dahl, CFD simulation and evaluation of different heating systems installed in low energy building located in sub-arctic climate, Build. Environ. 89 (2015) 160–169, https://doi.org/10.1016/j.buildenv. 2015.02.024. [4] M. Gustafsson, G. Dermentzis, J.A. Myhren, C. Bales, F. Ochs, S. Holmberg, W. Feist, Energy performance comparison of three innovative HVAC systems for renovation through dynamic simulation, Energy Build. 82 (2014) 512–519, https://doi.org/10. 1016/j.enbuild.2014.07.059.
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