Experimental study on thermal environment in a simulated classroom with different air distribution methods

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Introduction
The classroom provides a major indoor environment for children away or apart from their home as they may spend substantial time at school.The quality of the indoor climate in schools could be poor due to the lack of adequate outdoor air flow rate [1,2].Associations have been shown between poor IAQ and adverse health effects including asthma, absenteeism, and impaired performance on standardized tests [3].Poor indoor air quality causes sick building syndrome symptoms and distracts occupant's concentration.However, good air quality can enhance children's learning performance and teachers' work productivity [4].
Several studies have shown that increasing the classroom ventilation rate enhanced the performance of schoolwork of children [5,6].The effects of the indoor environment on the pupils' performance were studied widely.Wargocki and Wyon [7] summarized how classroom conditions affect the performance of schoolwork by children, motivated by the fact that the thermal and air quality conditions in school classrooms are now almost universally worse than the relevant standards and building codes stipulate that they should be.They showed low supply airflow rate and high CO 2 concentration can reduce children's performance of schoolwork by as much as 30%.
Therefore, insufficient ventilation has impact on learning outcomes in the classroom [8].Significant associations were observed between percentages of students scoring satisfactory in mathematics and reading tests and both indoor air quality and ventilation rate, which was estimated based on CO 2 levels [9].Johnson et al. [10] characterized IAQ in elementary school classrooms and estimated average effective fresh air ventilation rates under cold, mild, and warm season conditions using a transient mass balance modeling approach.The results showed that the ventilation adequacy varied within the schools across seasons and lack of adequate fresh air ventilation in many cases was found.
To achieve a good indoor climate, adequate airflow should be provided and the functionality of air distribution should be analyzed to guarantee the performance designed.To fulfill this target, the recommend minimum ventilation rate is 8 l/s/person for classrooms according to ASHRAE standard 66 [11].Because the occupancy density is high (1.8-2.4 m 2 /person) in school classrooms, compared, e.g., with offices, reasonably high airflow rates are required to supply in classrooms.
Most school classrooms worldwide experience raised air temperatures during increased heat gains, e.g., in warm weather.Other studies conducted in cold climates have also associated low ventilation along with high indoor temperature with decreased air quality [12,13].High heat gain increases convection flow that has significant influence on air distribution, and it makes challenging to fulfill the local thermal requirement [14].
The temperature and velocity conditions with typical air distribution methods have been studied in six classrooms by a field study without the internal heat gains [15].However, the results illustrated that high air velocities were found in the occupied zone without high internal heat gains.The previous studies have proved that the heat gains have significant effect on the air distribution [16,17].In practice, the spatial and temporal air distribution could be varied in different heat gain conditions and seasons.
Some studies proved that the distribution and strength of the heat gains have a significant effect on the airflow patterns.Lestinen et al. [18,19] measured the effect of heat load on indoor airflow characteristics.The results indicate a cause-effect relationship of the heat load strength and the heat load distribution on airflow characteristics that may also have effects on air distribution.It has generally been acknowledged that the convection flows caused by thermal loads may significantly affect the airflow pattern [20].The performance of air distribution is a combination of the influence of the cold and warm surfaces, the location of air diffusion, and thermal plumes of heat loads.As a result, the performance of air distribution should always be analyzed with different heat gain conditions to provide an acceptable thermal comfort when the air distribution strategy is designed.
The previous studies mostly focused on the indoor air quality and CO 2 concentration in a classroom and its impact on the pupils' learning performance.However, there is not much concern about the thermal comfort and airflow pattern in the classroom which also influences the students' perception.The existed studies about the thermal environment in the classroom are limited in the tropical or sub-tropical region [21,22].Therefore, the thermal environment in the Nordic condition should be paid more attention due to the high air flow rate set in the building code and high variety of the heat gain in summer and winter seasons.
The air distribution and thermal comfort need to be evaluated in detail to provide an excellent indoor climate for children.In this study, the performance of four typical air distribution methods was studied in a mock-up classroom with both cooling and heating situations.The main objective was to get a generic view of the performance of the differences of typical air distribution methods in both cooling and heating conditions with different occupancy ratios.The novelty of the study is to compare the thermal environment and airflow pattern in the classroom with a corridor-wall grille, a ceiling diffuser, a perforated-duct diffuser, and displacement ventilation, respectively.The test conditions included cooling mode, heating mode, full occupancy, and half occupancy were designed in a mock-up classroom.

Set-up of test chamber
To analyze thermal comfort conditions in a mock-up classroom, the physical measurements were conducted in the laboratory conditions at the Halton facilities.The dimenssion of classroom is 6.0 m × 4.4 m x 3.3 m (H), which is equivalent to half of the actual classroom (normal area 6.0 m × 10.0 m).The test conditions in the classroom were designed in three different load situations: (1) summer conditions with full occupancy (heat gain of 54 W/m 2 ); (2) summer conditions with partial occupancy (heat gain of 40 W/m 2 ); (3) winter conditions with partial occupancy (heating demand of 38 W/m 2 ).In the winter case of the full occupancy, there is no need any space heating.So, winter case of the full occupancy was not measured in this study.Heat gains and heat losses in the measurement cases are presented in Table 1.
The room air temperatures were set to correspond to Category A in partial occupancy with both cooling and heating conditions and Category B in full occupancy with the cooling condition [23].Category A means the high level of expectation to indoor environment and is recommended for spaces occupied by very sensitive and fragile persons with special requirements while Category B means the normal level of expectation.According to the Standard EN 16798 [23], the room air temperatures were 26 • C and 24 • C in the cooling case with full occupancy and partial occupancy.The set temperature was 24 • C in the cooling mode.However, the cooling power was designed so that the maximum room air temperature was 26 • C in the design condition (summer conditions with full occupancy).In winter conditions, the room air temperature was set to be 21 • C. The ventilation airflow rate was kept constant 90 l/s (3.4 l/s/m 2 ) in all cases.The airflow rate of 6 l/s/person is the minimum required ventilation rate according to The National Building Code of Finland [24].The supply air temperatures were 17 • C and 18 • C in the cooling and heating cases.Therefore, the cooling load supplied with ventilation was varied between 244 W and 455 W in the heating and cooling cases.The black ball temperature and room temperature were also measured in the middle of the room at the height of 1.3 m as the reference temperature.
Simulated window surface temperature was controlled at 30 • C and 11 • C in the cooling case and heating case such that the target heat gain in summer conditions or heat loss in winter conditions was achieved.The surface temperature was provided by conducting water pipes inside the window panel.The average heat transfer coefficient of simulated window is calculated as 8.0 W/(m 2 ⋅K) based on the power of water side.In laboratory conditions, heat losses were supplied by heat losses through structures, if necessary, to attain the required room air temperature.In the winter conditions, an underneath radiator (250 W) was utilized to prevent the draft risk of the cold window surface.
Occupants were simulated with heated dummies of the diameter of 0.4 m [25] and the height of 1.1 m including 0.1 m high legs against the vertical air distribution.The dummy was heated electrically with the power of 58 W, which describes the sensible heat gain of a child when seated and slight activity [26].Moreover, the dummy used in the

Table 1
The breakdown of the total heat gains and heat losses with the three cases analyzed.experiments originally had three large holes on its flank, near the top surface.In tests, the holes were blocked, since the warm air blew horizontally out of them when there was interaction between convection flow and jet.

Air distribution methods
The performance of four typical air distribution methods was studied with the three indoor conditions.The supply units were selected based on the thrown pattern in cooling and heating conditions.The measured air distribution schemes are: a corridor-wall grille (WTS-450-100), two displacement ventilation units (AFQ-125) on the floor in the corners of the classroom, a ceiling diffuser in the middle of the ceiling (TCV-160/A R4), and a perforated duct diffuser in the middle of the ceiling (HPS-160), as shown in Fig. 1.Air distribution units are shown in Fig. 2. The wall supply air grille is located in the wall opposite to the window with horizontal direction, two displacement ventilation units in the corners, and ceiling diffuser and perforated duct diffuser in the middle of the room perpendicular to the grilles.Two exhaust air valves are located equally in the ceiling near the wall-grille.
The utilized equipment were summarized in Table 2.All the measurement equipment were calibrated before the measurements conducted.In the mock-up classroom, indoor air temperatures and velocity were measured at seven horizontal planes (0.1 m, 0.5 m, 0.9 m, 1.3 m, 1.8 m, 2.4 m, and 3.1 m from the floor) with 24 evenly distributed locations in each plane (altogether 168 points).In this study, there are total 24 measurement locations which covered well the whole room area.The distance between two measurement points varied between 0.5 m and 1.2 m.Readings were 3 min average values.The air distribution with different methods was visualized with marker smoke.The measured section of the classroom, measurement locations, and sensor heights are shown in Fig. 1.

Evaluation indices
The local thermal discomfort was evaluated by the index of draft risk.The draft represents the unwanted local cooling of the human body caused by air movements.The draft rate (DR) is as given by Eq. ( 1) ) 0.62 ( 0.37 ⋅ u a,l ⋅ Tu + 3.14 where t a,l is the local mean air temperature, u a,l is the local mean air speed from 0.05 m/s to 0.5 m/s, and Tu is the local turbulence intensity in percent from 10% to 60%.If u a,l <0.05, use u a,l = 0.05; if DR> 100%, use DR = 100%.According to ISO 7730 [27], DR<10% meets the Category A of thermal environment; DR<20% meets the Category B. Category A means the high level of expectation to the indoor environment and is recommended for spaces occupied by very sensitive and fragile persons with special requirements while Category B means the normal level of expectation.Turbulence intensity (Tu) is defined by Eq. ( 2), as where u SD means the standard deviation of fluctuating velocity and umeans the mean air velocity.Heat removal efficiency (HRE) [28] is proposed to measure the effectiveness of heat removal from space as Eq. ( 3) where T ave(0.1−1.3) ( • C) means the mean air temperature from the height of 0.1 m-1.3 m, T ex means the exhaust air temperature and T su means the supply air temperature.
The air diffusion performance index (ADPI) concept combines local air velocity and local dry-bulb temperature into an effective draft temperature (Eq.( 4)) [29] https://www.sciencedirect.com/science/article/pii/S0378778816313809 -bib0005.The effective draft temperature (EDT) reflects the subjective sensation of coldness in terms of air motion and skin temperature deviation from surroundings [30,31].The effective draft temperature was a reliability and simplicity index to evaluate and predict the performance of mixing ventilation.The thermal comfort is improving while the EDT is approaching zero [32].
The ADPI is calculated as the percentage of an occupied zone within the ranges of acceptable air speeds (≤0.35 m/s) and effective draft temperatures (−1.7 to 1.1 • C).The ADPI of 100% implies the measurements taken at all the sampling points within the occupied zone conform to the given criteria above, and therefore the thermal condition in that space is expected to be at an acceptable level.Most air distribution systems are designed to achieve an ADPI of 80% or greater [32].The higher the ADPI, the higher the percentage of building occupants likely satisfied with the thermal conditions in the occupied zone.
where t i is air temperature at a test point, i; t a is spatial average air temperature in the entire occupied zone and u i is local air velocity.
where N θ is the number of measured points in the occupied space that EDT falls within −1.7< <+1.1 • C, N is the total number of points measured in the occupied space.

Supply airflow visualization with smoke
Smoke visualization of four types of air supply units in full occupancy cooling case is shown in Fig. 3. Two of air distribution methods are based on high momentum flux (corridor-wall grille and ceiling diffuser).Two of strategies are based on low momentum flux (displacement ventilation unit and perforated-duct diffuser).The supply air velocities were 2.9 m/ s and 3.5 m/s with corridor-wall grille and ceiling diffuser.With displacement ventilation unit and perforated-duct diffuser, the initial air velocities were below 0.2 m/s.The momentum flux of a wall-grille was   high enough to reach the other side of the room in Fig. 3a, even if wall grilles had vanes to direct supply air jet horizontally more evenly.Therefore, thermal plumes of occupancy and the heated window have minor effect on the performance of a wall-grille.The supply air spread effectively over the whole occupied zone with the displacement units (Fig. 3b).Moreover, this indicates the thermal load in the classroom did not affect the performance of displacement units.However, Fig. 3c shows that supply air from the ceiling diffuser tends to be carried along thermal plume moving to the opposite unheated wall in the perimeter zone (away from the heated window).This circulating airflow pattern was formed with thermal plumes generated by heated window go upwards then mix with the supply air and go downwards at the opposite side.The momentum flux of a perforated duct diffuser was very small.It had a tendency to create unstable flow conditions and varied loads can cause unexpectedly thrown pattern (Fig. 3d).
The airflow pattern was quite similar with the wall grille, displacement unit, and perforated duct in both cooling and heating situations.However, air distribution with smoke visualization is different with ceiling diffuser in the heating case from the cooling case.Fig. 4 shows that the airflow pattern was more uniform without the heat gain from the window.Therefore, the performance of the ceiling diffuser is vulnerable to the thermal plume from the window to some extent.

Temperature profile
The air and black ball temperature were measured at 1.3 m height in the middle of the classroom (Fig. 1).The room air temperature can be maintained below the designed temperature of 26 • C in the cooling condition with the studied air diffusers except for the wall-grille, as shown in Table 3.This is because of the large-scale circulating airflow pattern created by wall grille.Therefore, the difference from the designed temperature depended on the measurement location.In the heating situation, the room temperature can be kept around 21 • C. The exhaust temperature was lower than the air temperature in the occupied zone for all other supply air diffusers except for the displacement unit.Especially, the exhaust temperature was 0.8 • C higher than the air temperature in the partial occupancy cooling case with the displacement unit.This means that the displacement unit can remove the heat load more effectively than fully mixed concepts.
The measured vertical temperature distribution in the middle of the room (point 12 in Fig. 1) is shown in Fig. 5.With mixing ventilation concepts (the wall-grille, the ceiling diffuser, and the perforated duct diffuser), the air temperature was quite uniform at the different measurement heights with different situations.The maximum vertical temperature difference (0.4 • C) occurred in the full occupancy cooling case with the perforated duct.This indicates that the airflow pattern was fully mixed with the three mixing ventilation concepts in both cooling and heating conditions.However, there exists a temperature gradient in the all cases with the displacement concept because of low supply air temperature at the floor level.
The air temperature at the height of 0.1 m was lower than the designed temperature of 26, 24 and 22 • C, respectively.The air temperature was gradually increased from the height of 0.1 m-1.3 m due to the heat gains in the room.The vertical temperature difference was highest in the full occupancy cooling case and lowest in the partial occupancy winter case.This means that the strength of the heat gain has a significant effect on the vertical air temperature distribution with the displacement ventilation units.In spite of that, the vertical temperature stratification in the occupied zone of the displacement ventilation case was still acceptable (<3 • C) and the local thermal discomfort fulfilled Category B according to Standard ISO 7730:2005 [27].
In Fig. 1, total 24 air temperature measurement points are used in 2D space by contour plot.Algorithm for creating a contour is linear interpolation between the points [33].Fig. 6 shows the horizontal air temperature distribution with all the cases at the height of 0.1 m.In the full occupancy cooling cases, the air temperature was over 26 • C at all measurement locations with the wall-grille diffuser.Because the throw pattern can reach the other side of the room (heated window), the higher temperature did not occur at the window side.The highest temperature is located in the area near the corridor side.This indicates that the circulating airflow pattern from the grille reached the opposite window side along to ceiling zone.The airflow pattern then returned to the wall-grille side along to floor zone across the room.With the displacement unit, the air temperature at 0.1 m height was rather low (22.9 • C) as the supply temperature was 17 • C. In particular, the area near the displacement unit was much colder at 0.1 m height.In real applications, the location of the displacement supply units should be carefully analyzed to prevent near-zone draft.Space constrains should analyzed be project-based when the location and the type of the displacement units are selected.
The horizontal temperature distribution with ceiling diffuser was quite similar to the perforated duct.The air temperature was higher at the heated window side because of the thermal plume and lower at the opposite side.With the four types of the diffuser at 0.1 m, the biggest horizontal temperature difference (1.8 • C) existed with the displacement unit and the smallest difference with the wall-grille (0.7 • C).The phenomenon with the partial occupancy cooling cases was similar to the full occupancy.
In the heating cases, the radiator was installed under the window.The air temperature was over 21 • C in most of the area with the wallgrille diffuser.The highest temperature at 0.1 m height was reached in the area with the radiator.However, this condition did not occur with the other three types of diffusers.The temperature with the displacement unit was lowest compared to the other diffusers.With the ceiling diffuser and perforated duct, the mean air temperature was nearly 21 • C.Moreover, the horizontal air difference was lower with the ceiling diffuser and perforated duct (0.5 • C) and higher with the displacement unit (1.7 • C).In the case of partial occupancy, the heated window (cooling season) and the cold window (heating season) had minor effect on the temperature distribution.Fig. 7 shows the horizontal air temperature distribution at the height of 0.9 m was similar to 0.1 m.However, the air temperature was more uniform at 0.9 m than 0.1 m height.In particular, the horizontal temperature difference was smallest with the displacement unit at 0.9 m height being only 0.2 • C with the cooling cases and 0.5 • C with the heating case.Therefore, the airflow was nearly fully mixed at the height of 0.9 m.The difference between the measured air temperature and the designed temperature was negligible with the ceiling diffuser and perforated duct.

Heat removal efficiency
Table 4 shows the heat removal efficiency in the classroom with different air distribution methods.With mixing ventilation concepts (a wall-grille, a ceiling diffuser, and a perforated duct diffuser), the heat removal efficiency was quite similar, which all were close to 1.This indicates the airflow in the studied condition was near fully mixed.However, as it could be expected, the displacement unit made a much higher heat removal efficiency in the classroom.The reason is that the cold air supplied at the floor level is displaced and mixed with room air then extracted from the exhaust.Therefore, a larger vertical temperature difference existed with the displacement unit.The difference of heat removal efficiency between different heat gain levels was insignificant.

Air velocity profile
Figs. 8 and 9 show the measured air velocity at all locations and heights with the four types of diffusers under the cooling and heating cases.For the thermal environment, the design criteria for cooling season were 0.12 m/s, 0.19 m/s, and 0.24 m/s for Category A, B, and C, respectively [27].The corresponding values were 0.10 m/s, 0.16 m/s, and 0.21 m/s for heating season.
In the case of the cooling season, there existed high velocities (over 0.25 m/s) over the occupied zone with the wall-grille.Particularly, the air velocity was much higher in the area near the heated window side.The reason was that the throw pattern from the wall-grille reached the window horizontally then went down attached the surface of the window.At the height of 0.5 m, the flow pattern returned to the opposite side.This indicates that there was a large-scale circulating airflow created by the wall-grille in the space.Between the height of 0.1 m and 0.5 m, the air velocity was over 0.20 m/s in most space.Above 0.9 m height, the air velocity was below 0.20 m/s in the occupied zone.
However, the air velocity was very low (<0.12 m/s) with the displacement unit except for measurement points close to the cornerinstalled supply units.Furthermore, maximum air velocity just occurred at the height of 0.1 m but within the acceptable level.This is because there were two displacement units at the corners, which made the supplied air velocity much lower.Furthermore, the displacement ventilation concept was not sensitive to heat gain variation.With the ceiling diffuser, air velocities were reasonably low in the whole space (below 0.23 m/s).The air velocity at the heated window side was slightly higher than the opposite side.With the perforated duct diffuser, there existed locally high velocities (over 0.25 m/s) at the height of 0.1 m.With the partial occupancy, the air velocities decreased to 0.23 m/s.This indicates that the performance of the perforated duct diffuser was sensitive to higher heat gains in the classroom.
For the thermal environment, the design criteria with the heating season were 0.10 m/s, 0.16 m/s, and 0.21 m/s for Category A, B, and C, respectively [27].The air velocity in the heating season was quite similar to the cooling season (Fig. 10).The air velocity was reasonably low (below 0.14 m/s) in the whole space except at the ankle level with the displacement unit, ceiling diffuser, and perforated duct.With the wall-grille, the highest air velocity happened at the height of 0.1 m, where also the highest temperature occurred.Still, the higher velocity existed at the cold window side with the wall-grill.Based on the results in the cooling and heating season, the heating window and cold window had very limited influence on the air velocity distribution.Fig. 11 shows the average standard deviation (SD) of air velocity in all the cases.As expected, the standard deviation with the displacement  unit was the lowest in different conditions.This means the airflow pattern was quite stable even at the ankle level.The performance of the ceiling diffuser and perforated duct was similar.The most turbulent airflow occurred with the wall-grille.With the partial occupancy, the standard deviation was slightly lower than the full occupancy.Therefore, the heat gain had some effect on the standard deviation.

Draft rate
Fig. 12 shows the average draft rate at the height of 0.1 m-1.3 m with the cooling and heating seasons.The distribution of the draft rate followed the trend of the air velocity.In the full occupancy cooling case (Fig. 12a), the highest draft rate occurred at the height of 0.5 m (over 20%) with the wall-grille, which fulfills Category C of the thermal environment.In the partial occupancy heating case (Fig. 12b), the higher draft rate happened at the height of 0.1 m, which may cause thermal discomfort at the ankle.With the displacement unit, the draft rate was quite low except at the height of 0.1 m where the displacement unit was installed.However, the draft rate was kept at a reasonable value.With the ceiling diffuser and perforated duct, the higher draft rate also happened at the height of 0.1 m but within the acceptable level.The draft rate of the ceiling diffuser was lowest in both cooling and heating seasons and meet Category B (less than 20%).As a result, the performance of the ceiling diffuser was more stable with different thermal  environments.

Air diffusion performance index
Fig. 13 shows the effective draft temperature in all the cases according to Eq. ( 3).It should be noted that the 25th and 75th percentiles of the database fall within −1.7 to 1.1 • C in all the cases.The effective draft temperature was more concentrated around the 0 • C in the heating case and varied much in the full occupancy cooling case.This means the performance regarding thermal comfort was slightly better in the heating case than the cooling case with all the types of diffusers.The possible reason was that the difference between room temperature and As a result, the heat gain level was a key factor that would influence the effective draft temperature with the same volume of supply air.With the four types of diffusers, the interquartile range (75th percentiles-25th percentiles) was the smallest with the ceiling diffuser.Additionally, the median value was nearer to zero degrees with the ceiling diffuser, which represents a neutral thermal sensation.Therefore, the thermal comfort was better and more stable with the ceiling diffuser compared to the other diffusers.However, the performance of the displacement unit was the worst.The lower effective draft temperatures were found at the height of 0.1 m with the displacement unit.This is because of the lower air temperature and higher air velocity at 0.1 m height.
Fig. 14 shows that the air diffusion performance index (ADPI) was mostly more than 80% except for the displacement unit in the cooling mode.This is because that the higher vertical temperature difference exists.It should be noted that the ADPI was only slightly less than 80% with the displacement concept.This indicates that the 80% comfort acceptance was achieved with the wall-grille, ceiling diffuser, and perforated duct in both cooling and heating cases.The thermal conditions in the heating season was better than the cooling season with all air diffusers.The difference of ADPI was quite small between the full occupancy and partial occupancy in the cooling case.Therefore, the ADPI was not much sensitive to the heat gain level.

Discussion
Proper air distribution plays an important role in both the comfort of the occupants and the air quality of ventilated spaces.The air distribution should be introduced with the method that is able to meet various thermal requirements of occupied spaces with varied heat gain conditions.The quality of the indoor climate in the classrooms have been found to be poor in several surveys.To analyze thermal comfort conditions in a classroom, the physical measurements were conducted in the laboratory conditions in this study.The performance of four typical air distribution methods was studied in a mock-up classroom during the winter and summer seasons.This is essential information for the design of high performing classroom ventilation.
The supply air jet from the ceiling diffuser tended to be conveyed by thermal plumes from the heat loads in the cooling mode.However in the heating mode, air distribution was quite uniform, when there is window plume.Displacement ventilation was also the least sensitive of all studied supply concepts for different load conditions.However with displacement ventilation, there was high velocities in the vicinity to the supply units.Still, the average conditions of thermal comfort in the occupied zone were the best.During design process, the location of the supply units should be carefully analyzed to prevent near-zone draft risk.The space constrain should be also analyzed when the location for the displacement units are selected.
It should be noted that the performance of the wall-grille could be optimized only for one selected load condition.In principle, the throw length of wall-grille could be optimized for winter conditions leading to lower velocities close to the window workplaces, e.g., by selecting a larger supply area of the wall-grille.However, this increases draft risk in summer conditions when thermal plumes would affect the performance causing early jet detachment from the ceiling.Further, the detachment can lead to local thermal comfort problems.The performance of perforated duct diffuser was also quite unstable and sensitive when higher heat gains exist.In those conditions, supply air could unexpectedly drop down causing increased draft risk in certain locations.
Based on the results, air distribution method has significant effect on the local air temperature, velocity and draft risk.Some air distribution methods are sensitive to the convective flow and the whole performance can be changed when the internal heat gains varied.Thus, the air distribution methods selected for the classroom should be robust to internal conditions and seasons.
In mixing ventilation concepts, load conditions have a significant effect on the air distribution.In further study, the system performance should be analyzed in different conditions when the novel air distribution strategies are designed.In practice, it is not possible to analyze the interaction of convection flows and jets in the design phase without using CFD-simulation or laboratory mock-ups.It is recommended to use CFD or mock-up analysis to guarantee the performance of air distribution in varied load conditions.
The heat gain level and its distribution are the key factors that affect the air distribution and thermal environment in the classroom.This means that the results are valid only with the given experimental set-up.Therefore, a better understanding of heat gain distribution on air distribution is important with the practical application of different air distribution methods.

Conclusion
The measured and compared air distribution methods were a corridor-wall grille, a displacement ventilation unit, a ceiling diffuser, and a perforated duct diffuser in a simulated classroom.
• With the studied mixing ventilation concepts (a wall-grille, a ceiling diffuser, and a perforated duct diffuser), the heat removal efficiency was quite similar and airflow pattern was near fully mixed.However, the heat removal efficiency was highest with the displacement ventilation.• The corridor-wall grille created high velocities in both summer and winter conditions close to the window due to the high momentum flux of the supply jet.The performance of a wall-grille is not very sensitive to the strength and location of heat gains.• With the perforated duct diffuser in the classroom, the air distribution is quite unstable and sensitive to higher heat gains.This led to the thermal discomfort in the occupied zone.• In the studied classroom with the displacement ventilation concept, the draft rate (10%) was quite low except in the vicinity of the displacement unit.The air distribution pattern was not changed much in different load conditions.However, the air diffusion performance index was lower than other air diffusers.• With the ceiling diffuser, air velocities were reasonably low (below 0.23 m/s) within the occupied zone in all the test cases.The performance of the ceiling diffuser concept was quite appropriate in the varied load conditions.
Based on the results, the displacement ventilation and ceiling diffuser are recommended solutions for classrooms in the studied conditions.

Fig. 1 .
Fig. 1. a) Locations of the measured section in whole classroom and b) The height of measured sensor.

Fig. 2 .
Fig. 2. Air distribution in different cases: a) mixing ventilation with corridor-wall grille with adjustable vanes evenly spreading supply air jet onto ceiling area, b) displacement ventilation with low velocity unit, c) mixing ventilation with radial ceiling diffuser, and d) mixing ventilation with perforated duct diffuser.

Fig. 3 .
Fig. 3. Smoke visualization of four air distribution units in cooling case with full occupancy (54 W/m 2 ).

Fig. 4 .
Fig. 4. Smoke visualization of ceiling diffuser air distribution in the partial occupancy winter case.

Fig. 5 .
Fig. 5. Vertical temperature distribution in the middle of the room (at point 12) with a) full occupancy cooling case, b) partial occupancy cooling case, and c) partial occupancy heating case.

Fig. 6 .Fig. 7 .
Fig. 6.Horizontal air temperature distribution at the height of 0.1 m with all tested cases.Fig. 7. Air temperature distribution at the height of 0.9 m with all tested cases.

Fig. 8 .Fig. 9 .
Fig. 8. Air velocity distribution with four types of air supply units in the case of cooling with full occupancy.Color codes: level 0-0.12 m/s green, 0.13-0.20 m/s yellow, 0.20-0.25 m/s light blue and over 0.25 m/s red.(For interpretation of the references to colour in this figure legend, the reader is referred to the Web version of this article.)

Fig. 10 .Fig. 11 .
Fig. 10.Air velocity distribution with four types of air supply units in the case of heating with partial occupancy.Color codes: level 0-0.10 m/s green, 0.11-0.17m/s yellow, 0.18-0.22m/s light blue and over 0.22 m/s red.(For interpretation of the references to colour in this figure legend, the reader is referred to the Web version of this article.)

Fig. 12 .
Fig. 12.Average draft rate with four types of air supply units in the cases of a) full occupancy cooling and b) partial occupancy heating.Color codes: level 0-10% green, 10-20% yellow, 20-30% red and over 30% purple.(For interpretation of the references to colour in this figure legend, the reader is referred to the Web version of this article.)

Table 2
Measuring instruments in the tests.

Table 3
Room, black ball and exhaust temperature in different measurement cases.

Table 4
Heat removal efficiency with the four types of air diffusers in the cooling case.