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room air stratification in combined chilled ceiling and displacement ventilation systems.

by:Grade     2019-12-31
Displacement ventilation (DV)
Is an indoor air distribution method that can provide improved indoor air quality for pollutants discharged from heat sources (
Ventilation performance)
Compared to the diluted ventilation provided by the overhead hybrid system.
In the DV system, air is supplied at a very low speed through a supply device located near the floor level (
The most common is the low endwall diffusers)
Return near the ceiling.
Air supply temperature is slightly lower than room temperature.
DV can be used separately for cooling but should be coupled with another heating system. The ASHRAE (
Chen anglesman 2003)and REHVA (Skistad et al. , 2002)
The method is the most commonly used reference for DV system design and operation.
Supplying cold air on the floor in a layered environment may cause local thermal discomfort due to excessive ventilation and temperature stratification (
ASHRAE Standard 55, ASHRAE 2010). Hydronic-
Radio-based systems are related to energy conservation;
Therefore, it is a strong interest to combine the circulation system with DV\'s indoor air quality advantage.
The combination of cold hard floors and DV can increase the level and discomfort of riskat ankles, compared to \"DV-due to the temperature verticalstrategy-only\" system. Causone et al. (2010)
Laboratory experiments show that the combination of DV and floor cooling under a typical European office layout may result in an air temperature difference between the head and ankle exceeding the Comfort range specified in ASHRAE Standard 55 (ASHRAE 2010).
Temperature difference between 1. 1m and 0. 1 m (43 in. and 4 in. )
Change between 3. 2 K to 6. 6 K (5. 8 and 11. 9[degrees]F)
For the heat load between 31 and 76 w/[m. sup. 2](97. 8 and 239. 7 Btu/(h [ft. sup. 2]))
, The DVairflow rate varies between 35 and 80 L/s (74 and 170 cfm)
And the air temperature supplied varies between 16 [degrees]C and 22[degrees]C (60. 8 and71. 6[degrees]F).
They noticed that the vertical air temperature difference was reduced by increasing the airflow rate, thereby increasing the floor temperature.
They also showed that there was no significant increase in the risk of the draft. Mundt (1996)
It is found that the relevant mechanism affecting the temperature stratification level of DV system is the heat flow from the warm ceiling to the floor through radiation and the convection heat flow from the floor to the adjacent air layer. Causoneet al. (2010)
It is certain that in the case of cooling the floor, the large vertical gradient may be due to the ability of the cooling floor to remove the convection heat transfer on the floor surface and the reduction of air near the floor. The heat from the city. Causone et al. (2010)
It shows that the presence of radiation cooling floors does not affect the removal effect of pollutants (a. k. a.
Ventilation in Europe)
DV system.
In the above experiment, the average value of the measured pollutant removal effect was 1. 1 m (43 in. )of 2.
When the airflow rate is 50 L/s, 20 (106 cfm)
The average value is 5.
When the airflow rate is 80 L/s, 70 (170 cfm).
These values are similar to those obtained using DV alone.
Combination of refrigerated ceiling (CC)
DV is more attractive to the United States. S. markets.
There are two types of CC design :(1)
Radiation ceiling and (2)
Heat activated building system (TABS)
, Also known as the cycle plate.
Radiation ceiling has several advantages.
Their response time is fast, easy to control, and can adapt to fast-changing loads;
It is relatively easy to design;
Technology is well known.
They can also be used to transform applications and are compatible with traditional suspended ceiling systems.
The main disadvantages are related to the cost and the inability to store heat (peak-shave)
Their low operating average water temperature requires thoughtful spatial dew point control to avoid condensation.
Labels, usually built as circulating pipes in embedded plates, are cheaper than radiation plates and have the ability to cut peaks and shift, usually operating at higher cooling temperatures, reducing the risk of condensation.
The main disadvantages relate to the complexity of the design and control and the slow response of the hot block laboratory to the changing cooling load.
Alamadari, etc. (1998)
Finally, based on computational fluid dynamics (CFD)
Simulation shows that adding CCs to the DV system affects the air distribution properties of the dv cc board, thus changing the air temperature near the ceiling, resulting in downward convection, so that, increase the depth of the upper mixed warm and contaminated areas. In addition, the radiation heat transfer between the frozen panel and the wall reduces the surface temperature at room temperature below room temperature, causing downward convection near the wall, this may transport pollutants down from the top mixing area to the supply air and occupied area (Alamadari, etc. 1998).
They stressed that high thermal comfort was achieved in CFD simulations.
Reese and Jarvis (2001)
A node model was developed to represent indoor heat transfer in DV and CC systems.
According to the author, the model, by representing the air movement of the plume and other parts of the room, can correctly represent the relationship between the internal load and the air and surface temperature of the occupied part of the room.
The model is suitable for the annual energy simulation program (
Rees and Haves 2001)
But it cannot be applied as an independent design tool.
Novoselac and Srebic (2002)
The performance and design of the combined system of CC and DV were extensively criticized.
They said that the design of the combined system is more difficult than the design of the CC and DV systems that work independently, and because of the strong interaction of the two systems, use the design guidelines for CC and DV as a CC/DV system that is not suitable for a design combination for independent systems.
According to Novoselac andSrebic (2002)
One of the key parameters of the design is the cold load distribution between CC and DV systems. Tan et al. (1998)defined [eta]
Ratio of area cold load removed as CC to total room cold load. [eta]
May change between 0 and 1. If [eta]
Equal to 1, which means using a pure CC system.
On the other hand, if [eta]
Pure DV system is used.
Reduce the amount of cooling load removed by DV, I. e. , increasing [eta]
This means that the possibility of room layering is reduced, which in turn means a decrease in the ventilation efficiency of the system. Tan et al. (1998)
It is recommended to maintain a temperature gradient of at least 2 [degrees]C/m(1. 1[degrees]F/ft)
The DV system should remove at least 33% of the load (i. e. , [eta]= 0. 67). Behne (1999)
Indicates that good thermal comfort and air quality can be maintained when the DV system moves at least 20%-
25% of the total cold load. Ghaddar et al. (2008)
Developed a general design chart using a simplified plume multi-size CC/DV system
The thermal model of the conditional space developed by Ayoub et al. (2006).
Model developed by Ayoub et al. (2006)
Compared with CFD simulation.
The main limitation of the method is related to the fact that the design chart was developed for a ceiling cover factor of 100%.
The sensitivity analysis of the ceiling cover factor of 80% was carried out.
Factors with low ceiling coverage do not have data.
Understanding the minimum proportion of cold loads that dvo removes without breaking the layering is essential to maximize energy conservation and indoor air quality.
Ideally, only the minimum outdoor airflow rate required to maintain indoor air quality is provided, and potential loads are handled through a DV system, while all sensible internal heat loads are managed using a cooling ceiling.
In the tests reported in the literature, the active CC region maintained a constant ata high coverage rate.
The typical label is a large frozen surface area.
Research needs to be conducted to investigate what happens if the cooling area is reduced to a smaller percentage (typical radiant plate installation.
As described below, the study focuses on the layering of the internal office area.
The term \"internal area\" is used to refer to a non-thermal area (
Continuous space in a building controlled by a single thermostat)
Its thermal conditions are not affected by the presence of external windows.
The purpose of this study is to conduct laboratory experiments on typical AmericanS.
Internal area office configuration with CC/DV system1)
The ratio of cold load removed by CCs to total cold load and (2)
How the percentage of active ceiling area affects the air layering of the room.
Methods the experimental facilities and room description experiments were carried out in the climate Room (4. 27 m x4. 27 m x 3. 0 m [168 in. x 168 in. x 118 in. ])
Equipped with Radian panels located on the suspended ceiling with a height of 2. 5 m (98. 4in. )
Above the floor.
The climate Room is located in a huge test Hall.
The area of the climate chamber is 18. 2[m. sup. 2](196 [ft. sup. 2])
The volume is 54. 7 [m. sup. 3](1931[ft. sup. 3]).
There are no windows in the room.
Walls, ceilings and floors have similar structural and thermal properties.
From the outside, the chamber wall consists of 3 parts. 522 [m. sup. 2]K/W (20. 01([ft. sup. 2]h[degrees]F)/Btu)
Insulation, a stagnation of 0. 102-m (4-in. )airgap (0. 352 [m. sup. 2]K/W [2. 00 ([ft. sup. 2]h[degrees]F)/Btu])
Aluminum Extrusion wall, another layer of 0. 102 m (4-in. )
Polyurethane board (3. 522 [m. sup. 2]K/W [20. 01 ([ft. sup. 2]h[degrees]F)/Btu]).
By adding the component, the overall transmission ratio is 0. 135 W/[m. sup. 2]K(0. 0238 Btu/([ft. sup. 2]h[degrees]F)). [
Figure 1 slightly]
Aluminum radiation plate installed in the ceiling Area 1. 83 m (72 in. )long and 0. 61 m (24 in. )wide (Area is equal to 1. 11[m. sup. 2][12 [ft. sup. 2]]).
The copper tube is hot connected to the aluminum channel in the panel with a distance of 0. 15 m (6 in. ).
The ceiling is made of radiant ceiling panels in series.
Cotton fiber insulation on the panel (2. 288[m. sup. 2]K/W [13 ([ft. sup. 2]h[degrees]F)/Btu]).
Two radiation panels were used. In tests 12-1, 12-2, and 12-
3, 12 panels are used (13. 4 [m. sup. 2][144 [ft. sup. 2]]
The maximum limit is 73.
Area 5%); From tests 6-1 to 6-
5. 6 panels were used (6. 7[m. sup. 2][72 [ft. sup. 2]]
The maximum limit is equal to 36.
7%. ceilingarea)
, The second configuration is shown in Figure 1.
Figure 2 shows the location of two analog workstations, typical office thermal load, and instrument stations used to measure the vertical temperature distribution.
The entrance air is supplied to the room from 0. 6 m (24 in. )tall corner-
Installation of displacement diffuser (Figure 3).
In the course of the experiment, the radius of the booster is from 0. 31 m to 0. 46m (12 in. to 18 in. );
The different radial dimensions of the diffuser do not affect the thermal stratification in the chamber, but they affect the fluid dynamics field near the diffuser (adjacent zone).
In this study, the size and nature of adjacent areas are not important.
The heat source is summarized later in table 2.
Simulation of office heat source using floor-
Tower computers, flat screen and desk lamps are installed on the table and overhead lighting equipment.
According to EN 14240, the thermal dummy simulation of the occupant being heated (CEN 2004).
These simulators represent the load on the space by using a light ball enclosed in a metal sheet cylinder.
They try this radiantvective split with high quality
In order for the air to pass through, radiant paint and holes are used. [
Figure 2:[
Figure 3 slightly]
Measuring instruments and uncertainties for continuous monitoring of air temperature with resistance thermal device PT 100.
The sensor was calibrated before measurement.
The accuracy obtained is [+ or -]0. 15[degrees]C(32. 3[degrees]F)or better.
Temperature of water supply and return water ,[t. sub. w,s]and [t. sub. w,r]
Continuous monitoring with resistance thermal device pt100.
The sensor was calibrated before measurement.
The accuracy obtained is [+ or -](0. 03 + 0. 0005 x[t. sub. w]);
For the range of the measured value, the accuracy is [+ or-]0. 045[degrees]C (32. 1[degrees]F)or better.
Measure electricity with a power harmonic analyzer.
DV air supply temperature 【t. sub. air,s]
Measured inside the diffuser.
The exhaust leaves the room through a slot in the hanging ceiling, and finally leaves the return air chamber through the pipe to enter the surroundings.
[Exhaust]t. sub. air,r]
It was measured in that pipe.
The vertical tree is used to measure the air temperature at seven Heights (0. 1,0. 25,0. 6,1. 1,1. 7, 1. 9,2. 4 m [
4,10, 24,43, 67,75, 94 in. )]
Instrument station in the room (see Figure 2).
All air temperature sensors use a manufactured Meira cylinder to shield radiation from heat transfer.
The Earth temperature is measured at a height of 0. 6 m(24 in. )with a 0. 15-m [5. 6-in. ]black-
Thermometer (
Therefore, the thermal gain is equal to the cooling load.
Cooling load of radiation plate removal [CL. sub. CC]
Calculated with the following formula :[CL. sub. CC]= [m. sub. w], [c. sub. p,w]([t. sub. w,r]-[t. sub. w,s])(2)where [c. sub. p,w]
The specific heat capacity of water.
DV, load [removed]CL. sub. DV]
, Indirectly calculated as the difference between the total cold load and the cold load removed from the radiation ceiling.
The cold load removed by the Dvor can also be directly calculated by measuring the airflow rate and the temperature of the supply and return air.
Because the accuracy of the flow sensor is much higher than that of the flow rate sensor, this program is not used.
Table 1 summarizes the experimental situation.
The heat load in the room remains unchanged, equal to 631 W (34. 7 W/[m. sup. 2][109. 5Btu/(h [ft. sup. 2])]).
The thermal load is described in table 2.
Working temperature [t. sub. op]
Remain the same, almost equal to 24 [degrees]C (75. 2[degrees]F).
According to ISO 7726 Attachment G, the operating temperature is calculated as the average of the mean radiation temperature and the average seat temperature (ISO 1998).
The average temperature is the average value of the air temperature measured at 0. 1,0. 6, and 1. 1 m (
4, 24, 43. ).
In a layered environment, there is no height that can represent the \"perceived\" air temperature.
Therefore, the average air temperature measured at ASHRAEStandard 55 (ASHRAE 2010)
Height is used.
The average radiation temperature was measured at 0. 6 m (24 in. ).
DV air supply temperature 【t. sub. air,s]
Remain unchanged, equal to 18 [degrees]C(64. 4[degrees]F).
To keep the operating temperature set-
24: 00 [degrees]C (75. 2[degrees]F)
For all tests, the flow rate of water and cold water supply temperature are manually adjusted.
Air, water and mean radiation temperature;
Cooling water flow;
Record the airflow rate within at least 30 minutes after stabilization
The status condition was obtained.
Manually record the power consumption before starting the experiment.
The tests summarized in Table 1 are conducted in two phases.
The testing of all 12 panels was carried out in November 2009;
All the tests for 6 panels were conducted in August 2010.
For all tests, the position and size of the internal load are the same.
Super light configuration was replaced.
In the initial series of tests, in a ballast in the center of the room, the lamp above the head.
In the second series, the ceiling light is located in three positions parallel to the table on the ceiling.
All measuring instruments and data during the first visit
The recorder was installed indoors;
In the second visit, a datalogger (8 W)
Outside the room.
To compensate for the movement of data-
Recorder, 8w increase in internal load.
In order to verify that these minor changes did not affect the temperature stratification, there was no experiment with the radiation panel (only DV)
Both visits were repeated.
The temperature distribution obtained is very similar.
The average value of the difference in air temperature in the two cases calculated at each height is 0. 2[degrees]C (0. 36[degrees]F).
Results The main performance parameters of DV and CC systems obtained in the experiment are shown in Table 3.
The experiment is based on the calculation [eta]
The number of values and radians panels.
The operating temperature is controlled within the range of 23 °c. 7[degrees]C-24. 1[degrees]C (74. 7[degrees]F-75. 4[degrees]F);
Therefore, it can be concluded that the comparison is carried out under comfortable conditions of almost thermal equivalent (
The air speed and relative humidity are also constant).
The DV air supply temperature changes between 17. 9[degrees]C and 18. 1[degrees]C (64. 1[degrees]F and64. 6[degrees]F).
The airflow rate varies between 30 and 73. 5 L/s (64 to128 cfm)(1. 9-4.
8 air changes per hour).
The vertical air temperature distribution is shown in figure 4. [eta]
= 0 means a complete DV system is used. [
Figure 4 slightly][eta]= 0. 25, 0. 38, 0. 54, 0. 65, 0. 77, and 0.
81 said theCC removed 25%, 38%, 54%, 65%, 77% and 81% of the cold load, respectively.
It can be inferred from Figure 4 that the temperature stratification of the person sitting in the seat in the occupied area (up to 1. 1 m [43in. ]height)
For the condition range of the test, the change of cooling load between DV and CC will not be strongly affected.
At a higher level in the room, it can be seen that the temperature stratification decreases with the increase in the load amount removed by CC.
The suspended ceiling is located at 2. 5 m (98. 4 in. )from the floor.
Figure 4 reports the air temperature from the floor to the ceiling;
There is a gap between the ceiling and the exhaust.
I when the panel is activated. e.
, Cooling, exhaust [t. sub. air,r]
Lower than the temperature measured at 2. 4 m (94. 5 in. )by the panels.
Figure 4 shows that most of the temperature stratification occurs in the occupied area except for the pure DV test ([eta]= 0).
Relatively good
Mixed conditions (
Small temperature difference)
The higher height in the room is a good sign that these points fall above the layered height of the two characteristic lower and upper areas of the layered DV system. When [eta]
= 0, the temperature distribution indicates that the layered height is between 1. 1m and 1. 7 m (43 in. and 67 in. ).
When CC is opened ([eta]> 0)
, The layered height seems to be reduced to a height close to 0. 6 m (23 in. ).
Figure 5 shows the temperature distribution of four tests (6-2,12-1, 6-4, and 12-2)
, Where the cold load distribution is the same ([eta]= 0. 38 or [eta]= 0. 65)
But the number of panels is different (
To make up for this, R-
Square adjustment can be used. R-
The adjusted square is the value of r-
Adjust the square for more variables in the model.
Statistical analysis with R version 2. 10. 1.
The model containing all three variables has the highest R-
Square adjustment, but it presents a strong Many
Problem of collinearity;
So two.
The variable model is studied.
The best regression model in SI and IP is reported as follows: the radiation panel covers 70% of the ceiling.
The graph reproduced in Figure 9 shows the relationship between the vertical temperature gradient, ratio/Q and [eta].
R in the picture is called 【eta]. Tan et al. (1998)
The maximum ratio of suggested P/Q is equal to 18 KW /([m. sup. 3]/s)(28. 98Btu/(h cfm))
Because higher values can lead to a decrease in mixing and air quality in occupied areas.
Seven of the nine points obtained in these tests are drawn for comparison, as shown in Figure 9.
The temperature gradient has been calculated for the occupied area of the seated person.
The results show that the stratification is generally lower in the experiments in this study (roughly1[degrees]C[1. 8[degrees]F])
More than the method predicted.
Figure 9 shows the same [eta](
Figure for Rin)
The layers obtained are different for the same [eta]
, Lower layering when the number of panels is equal to six.
This approach tends to overestimate layering;
Therefore, it is possible to break the layering without being able to predict it.
In Figure 9, only 7 points are shown because [eta]= 0.
81, the p/Q ratio is 21, so it is beyond the scope of the figure;
The temperature gradient is 1 for this. 6[degrees]C (2. 9[degrees]F)fortest 12-3 and 1. 1[degrees]C (2[degrees]F)for test 6-5. [
Figure 9 omitted
The results given in this paper show that the CCcooling load ratio is not a sufficient predictor of stratification (see Figure 6).
Therefore, it is not enough to report the layering of a given [eta].
For example, [eta]= 0.
77, layered as 1. 1[degrees]C (2[degrees]F), and for [eta]= 0.
81, layered as 1. 6[degrees]C (2. 9[degrees]F).
The results of this study show that a better prediction of the stratified performance of the CC/DV system will be based on the ratio of the average radiation surface temperature and the total room cold load to the DV airflow rate.
In the experiment of 12 panels, 73.
5% of the ceiling area is covered with radiation panels.
The cold load of panel removal varies between 18 and 38 W/[m. sup. 2](
Panel area based on radiation)(56. 8 and 119. 9 Btu/(h [ft. sup. 2])]
Or between 13 and 28 W/[m. sup. 2](
Based on room area)[41. 0 and 88. 3Btu/(h [ft. sup. 2])].
The average surface temperature of the panel varies between 20. 1[degrees]C and23. 2[degrees]C (68. 2[degrees]F and 73. 7[degrees]F).
These values are representative of the hot-activated tablet app.
In the experiment of 6 panels, 36.
7% of the ceiling area is covered with radiation panels.
The cold load of panel removal varies between 24 and 73 W/[m. sup. 2](
Panel area based on radiation)(75. 7 and 230. 3 Btu/(h[ft. sup. 2]))
Or between 9 and 27 W/[m. sup. 2](
Based on room area)(28. 4and 85. 2Btu/(h [ft. sup. 2])).
The average surface temperature of the panel varies between 16. 5[degrees]C and 22. 6[degrees]C (61. 7[degrees]Fand 72. 8[degrees]F).
These values are typical values for metal radiation panel applications.
In the experiments reported in this paper, the limitation of this study was that the effect of the outer window on the distribution of airflow was not studied.
The experiment is carried out in the test room representing the internal area, where (almost)
Insulation wall.
Under cooling conditions, the rising hot plume may develop to a place close to the warm outside window.
There is no evidence of how this may affect temperature stratification and pollutant concentrations.
Therefore, the proposed model is valid only within the boundary conditions reported in this paper.
In this study, the effects of changes in cold load, Air Supply temperature, thermal comfort devices
The point, as well as the location and strength of the heat source, have not yet been investigated.
Conclusion a laboratory experiment was carried out to study the Air stratification of rooms with typical office space with radiation CC and DV.
* Average radiation ceiling surface temperature is a better predictor of head temperature difference (1. 1 m [43 in. ])andankle (0. 1 m [4 in. ])
Compared to other parameters related to the fraction of the total cooling load removed by the radiation CC, this result of the seated person in the occupied area explains the fact that when smaller activity is usede. g.
, For typical radiation ceiling layout)
, The cold radiation surface temperature required to remove the same amount of cold load (
Large area)
This caused more interference to the Air stratification of the room.
* For the range of test conditions covered, Room Air stratification in the occupied area :(1)
With CC removing a larger portion of the cooling load ,(2)
Increases as the surface temperature of the radiant ceiling rises, and (3)
Decreases as the displacement airflow rate increases.
* Although CC has an effect on layering, the results show that the minimum head-
Ankle temperature difference 1. 5[degrees]C (2. 7[degrees]F)
In occupied areas (
Sit or stand)
Will maintain the surface temperature of all radiant ceilings in 18 [degrees]C (64. 4[degrees]F)or higher.
* A model has been developed for DV and CC combinations to predict the temperature difference between the head and ankle as a function system of the average radiation surface temperature and the ratio between the cold load and the displacement airflow rate. DOI: 10. 1080/10789669. 2011.
592105 confirm that the current work is supported by the California Energy Commission (CEC)
Research on Public Interest Energy (PIER)
Construction projects and-
E. Physical donation of laboratory facilitiesH.
Prices, Winnipeg, Manitoba.
The author would like to thank Tom Epp for his help in setting out the work and preparing numbers.
Term CC = refrigerated ceiling [CL. sub. CC]
= Cold load for cold ceiling removal, W [CL. sup. DV]
= Cold load removed by DV system, W [c. sub. p,w]
= Specific heat capacity of water J /(KgK)
DV = displacement ventilation [m. sub. w]
= Water mass flow of [kg/hour meant. sub. w,s]and[t. sub. w,r], C[degrees]([degrees]F)[t. sub. w,r]
= Water temperature returned from the frozen ceiling ,[degrees]C([degrees]F)[t. sub. w,s]
= Water temperature to supply the cooling ceiling ,[degrees]C([degrees]F)[V. sub. air]
= Airflow rate of DV system, L/s (cfm)[eta]
= Cooling load ratio for cooling ceiling removal ,[CL. sub. CC]
Alamadari, F. Total Cold load, P. F. Butler, P. F. Grigg, and M. R. Shaw. 1998.
Cold ceiling and displacement ventilation.
Renewable energy source 15: 300-5. ASHRAE. 2010.
ANSI/ASHRAE Standard 55 2010, thermal environmental conditions for human habitation.
Atlanta, Georgia: American Society of Heating, Refrigeration and Air
Air conditioning engineer LimitedAyoub, M. , N. Ghaddar, and K. Ghali. 2006.
A simplified model of space cooling combined with a frozen ceiling and a displacement ventilation system.
HVAC Research and Development 12 (4): 1005-30. Behne, M. 1999.
Indoor air quality of the room with cooling ceiling.
Mixed ventilation or displacement ventilation.
Energy and Construction 30: 155-66. Causone, F. , F. Baldin, B. W. Olesen, and S. P. Corgnati. 2010.
Floor heating and cooling combined with displacement ventilation: Possibilities and Limitations.
Energy and architecture 42 (12):2338-52. CEN (
European Committee on Standards). 2004. EN 14240-
2004 ventilation of buildings--
Refrigerated ceiling-
Test and rating.
BRUSSELS, Belgium: CEN. Chen, Q. , and L. Glicksman. 2003.
System performance evaluation and design guidelines for displacement ventilation.
Atlanta, Georgia: Ashley. Ghaddar, N. , K. Ghali, and W. Chakroun. 2008.
The thermal model was simplified through experiments, and the optimized cold ceiling and displacement ventilation system were designed. ASHRAE RP-
1438 final report, Ashray, Atlanta, GA. ISO. 1993.
Guide for the representation of measurement uncertainty.
GENEVA, Switzerland: International Organization for standardizing. ISO. 1998.
ISO 7726 international standard: Ergonomics in thermal environment-
An instrument that measures physical quantities.
GENEVA, Switzerland: International Organization for standardizing. Mundt, E. 1996.
Performance of displacement ventilation system--
Experimental and theoretical studies.
Doctoral thesis, Bulletinno.
Building services engineering, KTH, Stockholm. Novoselac, A. , and J. Srebric. 2002.
A critical review of the performance and design of the combined cooling ceiling and displacement ventilation system was conducted.
Energy and architecture 34 (5):497-509. Rees, S. J. , and P. Haves. 2001.
A node model for office space displacement ventilation and cold ceiling systems.
36: 753-architecture and environment62. Skistad, H. , E. Mundt, P. V Nielsen, K.
Hargerstrom and J. Railo. 2002.
Non-displacement ventilation
Industrial premises. Guidebook no.
Reva, Brussels, Belgium. Tan, H. , T. Murata, K.
Blue Wood, Kurabuchi. 1998.
Cooledceiling/replacement ventilation composite air conditioning systemdesigncriteria.
Proceedings of Roomvent 98 1: 77-84. Zhang, H. , C. Huizenga, E. Arens, and T. Yu. 2005.
Modeling of thermal comfort in layered environments.
2005: Minutes of 10 international conference on indoor air quality and climate, Beijing, China.
Received on December 16, 2010;
In May 23, 2011, he received a doctorate in Stefano Schiavon, assistant professor ASHRAE.
Fred Bauman of PE member ASHRAE is a research expert.
Bendully, Peng, member ASHRAE, R & D manager.
Julian Rimer, Peng, ASHRAE, manager Hua Fanno Schiavon ,(1)
* Fred Bowman (1)Brad Tully, (2)
University of California, Berkeley (CBE)
390 Worcester Pavilion 1839, Berkeley, CA-947201839, USA (2)
MB ir2k3 Z9, Canada Company price Group, 638 Raleigh Street, Winnipeg * corresponding author e-
Email: stefanoschiavon @ gmail
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