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Performance comparison of conventional and chilled ceiling/displacement ventilation systems in Kuwait.


Buildings in Kuwait account for more than 50% of all energy use in the country [Mesheshwari et al. (2005)]. In the summer, the HVAC systems represent 70% of peak load. The combined chilled ceiling and displacement ventilation (CC/DV) system is a system that has potential for energy savings with improved air quality in spaces where the cooling load does not exceed 100 W/[m.sup.2] [Novoselac and Srebric (2002)]. The CC/DV system has not been introduced in Kuwait HVAC market. The current paper is concerned with assessing the perspective of this system performance, applicability, and energy saving potential in Kuwait climate and building code.

In displacement ventilation, the cooler air entering the room at the floor level displaces the warmer room air that rises due to its natural buoyancy effect. Consequently, the bottom occupied zone contains the fresh cool air with no recirculation flow while the heat and contaminants produced by the room activities rise to the ceiling level where they are exhausted [Jiang et al. (1995) and Yuan et al. (2001)]. The chilled ceiling carries a portion of the sensible cooling load and the DV system carries the rest of the sensible load in addition to the latent cooling loads. Design and operation of integrated chilled ceiling (CC) and displacement ventilation (DV) system is not as straight forward as the case when one system is used in meeting the cooling load. The CC/DV modeling is more involved due to nonlinearity in heat transport processes and stratification present in the room. The system operation required setting of several parameters that need to be simultaneously specified to insure meeting thermal comfort, IAQ inside the space while condensation on the ceiling is prevented. The operational variables include: the supply air conditions (flow rate, temperature and humidity) of the DV system and the chilled ceiling temperature. Ghaddar et al. (2008) proposed design charts that are based on the ratio R of the CC cooling load to the total load, the thermal comfort represented by the temperature gradient (dT/dZ), and the amount of displaced air parameter (P = (Q/m)). Their design chart parameters included the temperature ranges of the supply air temperature and the chilled ceiling for any R in the feasible design regions where dT/dZ is less than 2.5 K/m and with the facility to read off the stratification height H and insure that it is above 1.2 m. Stratification height is the elevation at which the density gradients disappear in the rising air and the plume spreads horizontally at a minimum level for acceptable indoor air quality in the occupied zone. The design chart of Ghaddar et al. (2008) provided information to select an accepted design, but it does not give information about expected performance under transient conditions initiated by external solar and environmental load and internal load schedule. The charts are based on a wall-plume-multi-layer model of Ayoub et al. (2006) that is insensitive to position of the heat sources in the room. Ghali et al. (2007) extended the plume-multilayer to transient response to take into consideration variable load and heat storage in walls and predict system energy consumption. Their model is well-suited for the current application because it accounts for plumes associated with non-uniformly heated walls; and it accurately predicts the stratification height and the vertical air and wall temperature gradients as a function of supply conditions, and the chilled ceiling temperature. This model will be instrumental in the assessment of the CC/DV system performance in Kuwait, but needs to be validated with experiments for cases when transient external load is present which could be a substantial load for Kuwait weather. The climate conditions of Kuwait and building construction practices necessitate careful design of the combined system to achieve energy savings. The climate is characterized by a high solar load and well-insulated building envelops with conductance U less than 0.72 W/[m.sup.2]K.

The first objective of the work is to experimentally test the transient wall-plume-multilayer model when transient external load is present in the conditions of Kuwait to check the attainment of thermal comfort and indoor air quality with an accept able stratification height. The second objective is to investigate the potential for energy savings and economic feasibility of the use of the CC/DV system compared to current air conditioning practice of using conventional mixed convection systems in Kuwait for a case study of an office space. A design procedure based on Ghaddar et al. (2008) design charts will be followed to size each of the two subsystems (chilled ceiling and displacement ventilation) and determine the chilled ceiling load removal ratio R to the total room load such that thermal comfort and stratification height are attained at the peak load [Ghaddar et al. (2008)]. A comparison will be made of the annual energy consumption and cost of the CC/DV system compared to that of the conventional system at similar comfort and indoor air quality conditions.


We need to assess accuracy of predicted comfort and IAQ parameters (dT/dZ, and H) of the transient 1-D model of Ghali et al. (2007) in presence of external load in Kuwait weather. The objective of the experiment is to compare the experimentally measured and predicted values of the conditioned room stratification height and vertical temperature gradient at the given design parameters of chilled ceiling temperature and supply air temperature and flow rate during the system operation under transient space load.

Experimental Setup and Measurements

The test facility conditioned space by a CC/DV system has a length of 4.59 m, a width of 4.96 m and height of 3.80 m. The room has one aluminum door (2.5 x 1.03 [ms.sup.2]) hollow core without thermal breaks and two glass windows (1.1 x 1.4 [m.sup.2]). The North and West walls are internal partitions with U = 1.704 W/[m.sup.2]*K, while the East and South have U = 0.568 W/[m.sup.2]*K. The floor conductance U is 0.398W/[m.sup.2]*K and the roof conductance is U = 1.931 W/[m.sup.2]*K. The U-values for the walls, partitions, roof and floor are based on the Thermal Standard of the State of Kuwait, (1994). The schematic diagram of CC/DV room is shown in Figure 1. On the North wall, the air supply diffuser is located 10 cm above the floor. It has a length of 2.5 m and a height of 0.42 m. The air exhaust slot (1.5 x 0.42 [m.sup.2]) is positioned above the diffuser by 1.75 m and 0.55 m from the ceiling. The chilled ceiling panel that covers 80% of the ceiling is suspended from the ceiling and is well insulated from the back side to prevent conduction through the roof. The chilled water pipes are 15 mm diameter copper tubes. The pipes are pressed in 85% aluminum sheet of 1 mm thickness. The chilled water is supplied from a main chiller in parallel to the headers of the panel. Three heat sources with 150 W each is placed in 3 metallic cylinders at the center of the room as shown in Figure 1 to represent the human with surface temperature not exceeding 35[degrees]C. The cylinder is 1.2 m in height and 0.4 m in diameter.


We are interested in measuring the velocity of air inside the environmental chamber, the temperature of air, the stratification height inside the chamber, and the temperature gradient with respect to the height of the chamber (dT/dZ). The temperature distribution inside the room is measured by mounting T-type shielded thermocouples (accuracy of [+ or -]0.5[degrees]C) on four rods inside the room at different locations (see Figure 1). Two small rods are positioned in front of the diffuser to measure the supply air temperature. The thermocouples are placed 10 cm from each other on the rods. Two thermocouples are mounted on the North wall, one on the East wall, two on the South wall and one on the West wall. Thermocouples are also mounted on the chilled ceiling and two external thermocouples are mounted outside the window to measure the ambient temperature. The data acquisition system of type OMB-Multiscan-1200 is capable of recording data every second. The measured temperature distribution is used to determine the vertical temperature gradient. Local velocities in the room are measured in an air column using a 1-D anemometer Extech 407412 type. Velocity measurements spanned several positions in the area away from the heat sources at incremental heights from a height Z = 0.55 m from the floor to a height Z = 2.0 m (2cm apart). The air velocity is recorded from the equally spaced mounted sensors on the 6 column/rod positions every minute. We are interested in the vertical component of the velocity since this is the component that will allow determining the stratification height based on the level where the supply flow rate is equal to the upward flow at that level. When the vertical component first dies out (v = 0) moving upward from the floor towards the ceiling, then this will be the stratification height with accuracy of [+ or -]0.02m.

Two experiments were conducted within one week during August 2007. The first experiment monitored the room temperature distribution and velocity from 12 hr to 16 hr with transient dominant external load at mean value of 47.3 W/[m.sup.2]. There was no direct radiation on the external wall during the experiment, but the outdoor temperature increased steadily from 40.8[degrees]C to 44.5[degrees]C at 4 pm. The second experiment was done when the three heat sources were turned on during the operation of the system from 12-16 hr of the same day with a mean load of 73.85 W/[m.sup.2]. The internal heat sources were steady and did not vary in time as they were turned on at noon. In both experiments, the system was run at 8 am, but the measurements of temperature and velocity were recorded starting at noon. The heat sources started at noon as well. For the experiments of CC/DV room with and without heat sources, the ceiling temperature was set 18[degrees]C and the supply air temperature was set at 22[degrees]C. The air supply flow rate entering the room was 0.5 kg/s.

Model Validation Results

The transient plume-multi-layer model of Ghali et al. (2007) is used to predict the room vertical temperature distribution and the stratification height. Accurate initial conditions of the room temperature and walls are needed to predict the transient behavior. Fine tuning of initial conditions was done using the measured wall temperature and external air temperature. The weather data of Kuwait for the few days preceding the experiments dates were obtained and simulations were done to make sure that initial boundary conditions used at the date of the experiment meet the measured values of surface temperature of the walls at the start time of the experiment.

The room air is divided into four vertical zones or layers. The first zone is the floor zone (zone 1, height of 0.75 m), then the middle occupied zone (zone 2, height of 0.75 m), followed by an upper zone (zone 3, height of 1.55 m), then by the ceiling adjacent zone (zone 4, height of 0.75 m). The measured air temperature of each zone is averaged and is plotted as a function of time for the two experiments in Figure 2a for the case when the internal heat sources were not turned on and in Figure 2b when the heat sources were on. The room mean load was 47.3 W/[m.sup.2] in the first case and 73.8 W/[m.sup.2] in the second case. The predicted values of temperature are plotted on the same figures and show good agreement with the experimental value. The vertical temperature gradient remained below 2.5 K/m at all times. Since the ceiling temperature was not changed in both cases, the internal load caused an increase in the occupied zone temperature, but it was still within comfort values. The vertical gradient in presence of internal heat sources was less than the case with lower internal load. The air in the floor zone will have elevated temperatures due to the presence of the hot surfaces which also increases the adjacent upper zones temperature.


The stratification height was estimated form the velocity measurements and is predicted using the plume-multi-layer model of the test room. Figure 3 shows the measured and predicted stratification height in the room for (a) room with external heat source only, and (b) room with both internal and external heat sources. The agreement between the experimental and model values is good with error less than 4 cm in both cases. The presence of internal heat sources has increased upward flow and also increased the stratification height. Note that the stratification height when measured very close to the sources, it will give higher velocities. The velocity measurements at different locations and heights provide fairly accurate estimate of the velocity profile from which the stratification height is determined. The values of the stratification height were measured away from the heat sources.



The test case considered for the study is a 5 x 10 x 3 m space. The north wall has a single glazed glass window with an area equal to 7.5 [m.sup.2]. The ceiling and floor are considered internal partitions. The walls are external walls. It is assumed that all the walls have the same construction with an overall heat transfer coefficient U of 0.57 W/[m.sup.2]*K. The chilled ceiling covers 80% of the area. Twelve occupants are present in the office according to the schedule shown in Figure 4. The heat load is 250 W/person, including lighting and equipment loads, and the moisture generation is 2x[10.sup.-5] kg/s/person. The desired indoor design conditions of the occupied space are taken as 23 [degrees]C and 55% relative humidity. The space peak load was 78 W/[m.sup.2] of which 60 W/[m.sup.2] is internal load by people and equipment and 18 W/[m.sup.2] external load due to transmitted solar load through walls. The space latent peak load portion is 22 W/[m.sup.2] due to people and ventilation requirements of 7 l/s/person (ASHRAE, 2007).


A design procedure based on Ghaddar et al. (2008) design charts was followed to size each of the two subsystems (chilled ceiling and displacement ventilation) to determine the chilled ceiling load removal ratio R to the total room load at peak outdoor and internal load conditions such that thermal comfort and stratification height are attained at this peak load. The design chart of Ghaddar et al. (2008) gives a relationship between the ratio of total space cooling load to ventilation airflow mass flow rate, the ratio of cooling output of the CC to total space load, and the proposed values of supply air temperature and the chilled ceiling temperature. At a chilled ceiling temperature [T.sub.c] of 16 [degrees]C and supply air temperature [T.sub.s] of 23.5[degrees]C, the design parameters obtained are presented in Table 1. The design parameters insure that high air quality associated with a 100% fresh air zone of stratification height larger than 1.1 m and thermal comfort where vertical temperature gradient is below 2.5 K/m are both attained at peak load. In displacement ventilation with the case of warm concentrated contamination sources of [CO.sub.2] and others generated by people, all contaminants are transported directly into the upper zone by the convection flows (Yuan et al., 2001). The [CO.sub.2] transport is assumed here to follow thermal transport at the same velocity of the air (Yuan et al., 2001). The total space load is 3.9 kW, of which 1.8 kW is removed by the chilled ceiling, 1.0 kW sensible load and 1.1 kW of latent load removed by the DV system.
Table 1. The Combined CC/DV System Design Parameters for the Test Cast
at Outdoor Peak Temperature of 48 [degrees] C and Humidity Ratio of
0.0157 kg x [H.sub.2] O/kg of Air During July

Design Parameter                                                 Value

Chilled ceiling temperature in [degrees] C                          16
Supply air temperature in [degrees] C                             23.5
Supply air flow rate in kg/s                                      0.25
Floor temperature in [degrees] C                                    24
Supply air humidity ratio in kg [H.sub.2]O/kg of dry air          0.01
Return air humidity ratio in kg [H.sub.2]O/kg of dry air         0.011
Ceiling adjacent air layer Temperature in [degrees] C            26.76
Ceiling dew point temperature in [degrees] C                      15.5
Average walls temperature in [degrees] C                          21.7
Predicted Mean Vote PMV                                           0.15
Stratification height in m                                         1.1
Vertical temperature gradient in the occupied zone dT/dZ in K/m   1.57

To size the system chiller, the dehumidification load should be added because the outdoor air is to be cooled and dehumidified to the specified supply conditions. Figure 5 shows a schematic of the system suggested to meet the cooling load. It includes the use of a cooling coil to cool and dehumidify outdoor air supplied to the space. The use of a sensible wheel at effectiveness of 0.8 permits the pre-cooling of outside air which is at least 20[degrees]C higher in Kuwait weather than desired supply temperature to the room. The system peak cooling load is 9.69 kW of which 2.67kW is latent load on the cooling coil, and 6.97 kW is sensible load on the supply air cooling coil after recovery and the chilled ceiling.


Conventional System Sizing and Design at Peak Load

The office space test case described in the previous section with same load schedule was simulated using TRNSYS (2004) software to determine the required size of a conventional mixed ventilation system for equal PMV comfort values obtained in the CC/DV system design. The ventilation requirement of the same test case space according to ASHRAE (2005) is 0.12 [m.sup.3]/s for 12 persons in the space. The supply air temperature at peak load was taken at 14 [degrees]C result ing in supply flow rate of 0.4 [m.sup.3]/s. Two simulation cases are considered. The first case uses 100% fresh air supply (0.4 [m.sup.3]/ s) while the second case uses 0.12 [m.sup.3]/s fresh air (30%) and 0.3 [m.sup.3]/sec from the return air. The supply air humidity ratio is kept at less than 10 g/kg to maintain an indoor relative humidity of less than 55%. For both simulation cases, the peak load occurs in July. For the 100% fresh air system, the peak load is 19.8 kW of which 12.5 kW is sensible load and 7.38 kW is latent load related to occupants and outdoor air dehumidification. For the partial fresh air system, the peak load is 6.9 kW of which 4.18 kW is sensible and 2.72 kW is latent. The thermal comfort predicted values of PMV were comparable to the values obtained for the CC/DV system.

Based on peak load, the partial fresh air conventional system size is considerably smaller than the cooling unit size of the CC/DV system. However, a considerable reduction in the cooling unit size from 19.8 kW in the conventional 100% fresh air system to 9.69 kW in the CC/DV system is realized. The reduction in chiller size originates from the reduction in the ventilation load (fresh air supply) due to the high outdoor air temperature. The 100% fresh air conventional system provides high indoor air quality equivalent to the CC/DV system. The CC/DV system has a higher initial system cost than the conventional system. The cost effectiveness cannot be assessed on the basis of the system size, but on the energy consumption of the system during the cooling season.

The simulation based transient wall-plume-multi-layer model of Ghali et al. (2007) gives an estimation of the energy consumption of the designed system over a period of time to compare performance with a conventional system performance. The conventional system annual energy consumption is also calculated during the cooling season for Kuwait weather while maintaining the same level of comfort. In the transient analysis of the CC/DV system, the ambient conditions are not constant. The hourly direct and diffuse solar radiation incident on the walls and the values of ambient temperature are derived directly from hourly weather data files of Kuwait for a typical day of each month of the season when cooling is needed from April to October [Carrier, Version 4.20a, 2006]. The internal and external wall convection coefficients are assumed 3.05 W/[m.sup.2]*[degrees]C and 16.66 W/[m.sup.2]*[degrees]C. The sensitivity of the results to the time step and the grid size in the wall was examined for convergence. It was found that using a time step less than 100 s and a spatial grid size less than 1.2 mm resulted in a change less than 0.1[degrees]C in the internal wall temper atures. For each month, the simulations were done over a period of 10 days to reach a convergent steady periodic solution for a 24 hour period representative of the month.


The transient simulation of the CC/DV system is performed using the plume-multi-layer model for the space to calculate the system cooling energy demand from April to October. The supply mass flow rate and temperature are kept constant at 0.25 kg and 23.5[degrees]C, respectively. In order to have a level of thermal comfort comparable to that of the conventional system, the chilled ceiling temperature was varied to meet the thermal comfort constraint during transient load. The ceiling temperature was set at 16[degrees]C at peak load. The stratification height is found to vary between 1.1 m and 1.25 m, which satisfies the indoor air quality constraint.

Figure 6 shows the hourly variation of the system load removed by the 100% fresh air conventional system as compared to the load removed by the displacement ventilation system and the chilled ceiling system. Note that for the given space load, the DV system load shown in Figure 6 is much higher than the CC system since it also includes the latent load removal and the cooling of supply air from ambient conditions to supply conditions. The fraction of the chilled ceiling load ratio to total load. Over a period of 24 hours during July, the fraction of the load removed by the chilled ceiling varied from 49% to 55% of the space sensible load.


Figure 7 shows the hourly variation of the average wall surface temperature in the conventional system, the room average air temperature, the occupied zone temperature and the average walls surface temperature in the CC/DV system for the month of July. For the conventional mixed ventilation system, the space air temperature is constant at 23[degrees]C while the average wall surface temperature varied between 24.3[degrees]C and 25.5[degrees]C. The average wall surface temperature in the case of the CC/DV system is about 21.7[degrees]C, which is significantly lower than the conventional system due to the presence of the chilled ceiling set at 16[degrees]C. The occupied zone temperature in the CC/DV system ranges from 23.4[degrees]C to 24.1[degrees]C and it is higher than that of the conventional system (23[degrees]C) although the two systems have comparable levels of thermal comfort. This difference in the air temperature is offset by the lower average surface temperature of the CC/DV system. Therefore the radiant temperature which is a factor in the thermal comfort will be lower for the CC/DV than that for the conventional system; hence improve the thermal comfort.


The monthly cooling energy loads per unit floor area per unit time of the CC/DV system and the mixed ventilation systems at 100% fresh air supply and at 30% fresh air supply are shown in Figure 8 for the test case. The maximum cooling energy demand occurs in the month of July. The 100% fresh air conventional system provides similar indoor air quality in the occupied zone to the CC/DV system. The monthly mean cooling power consumption is averaged over the 24 hours per unit floor area in W/[m.sup.2]. The electrical power consumption can be obtained by dividing the mean thermal power consumption by the coefficient of performance of the chiller system. The mean cooling energy required of the CC/DV varied from 119 W/[m.sup.2] in October to 162 W/[m.sup.2] in July. The mean cooling energy required will lead to mean electrical power consumption that is within the standard code of practice in Kuwait requires that the maximum electrical energy consumption for air condition space should not exceed 70 W/[m.sup.2]. This is easily satisfied for typical coefficient of performance of chillers exceeding 2.5 and higher.


The total cooling energy of the CC/DV system is 21.8 GJ compared to 44.3 GJ for the conventional system during the cooling season. A reduction in total energy demand by 53% is attained when CC/DV system replaces the mixed ventilation system while satisfying the same comfort and air quality conditions. At 30% fresh air of supply flow, the conventional mixed system uses 14.8 GJ. This is less than the energy use of the CC/DV system. However, we have to take into account that the CC/DV system supplies 100% fresh air while we used only about 30% fresh air in the mixed system. This means that the internal air quality of the CC/DV system would be better than that of the partly fresh air system.

It is important to perform an economic feasibility analysis on the

test case because the CC/DV system has a higher initial cost. This is due to cost of the heat wheel and the cost for installation of the chilled panels that ranges from 10 to 36 $/[m.sup.2]. On the other hand, the cooling system size is reduced from 19.8 kW in the 100% fresh air conventional system to approximately 9.69 kW in the CC/DV system. The cost of other equipment is assumed to be the same for both systems even though the maintenance cost for the CC/DV is a lot less. The incremental initial cost, excluding chilled ceiling installation of the CC/DV system, is considered for three values of $700, $1,000, and $1,300 over the initial cost of the conventional system. The coefficient of performance of the vapor compression system is assumed to be equal to 2.5 and the electric energy cost is assumed to be 0.15 $/kWh based on unsubsidized electrical energy production cost. The efficiency of the heat wheel is assumed to be 0.8. The saving in the yearly energy bill is about 3,125 $/year assuming a discount rate of 0.08 and an inflation rate of 0.05. Figure 9 shows the payback period of the CC/DV system versus the cost of the chilled ceiling per unit area of the space for the different incremental initial costs of the system over the conventional system. A payback period of less than 3 years can be achieved at all considered incremental costs. The cost effectiveness analysis is only indicative since it is limited to the test case considered in this study.



The performance of chilled ceiling and displacement ventilation (CC/DV) system has been experimentally and theoretically studied using the plume-multilayer model of Ghali et al. (2007) for a test case in Kuwait to assess the added value in satisfying comfort at high indoor air quality as compared with conventional system performance. The transient plume-multi-layer model predicted well in comparison with experimental values the vertical temperature distribution and stratification height in a test room conditioned by a CC/DV system in presence of external load in Kuwait weather.

The performance of the CC/DV system has been applied to a test office space in Kuwait and the energy consumption of the system is evaluated over the whole cooling season from April till October and results are compared with the 100% fresh air conventional system energy consumption at the same thermal comfort level and IAQ. The energy consumption of 100% fresh air systems was less than 50% of the conventional system energy consumption at 100% fresh air. The use of CC/DV system was shown to be feasible and cost effective for use in Kuwait climate with a payback period less than 3 years depending on installation cost of the ceiling panel. However, the CC/DV is not more efficient than a 30% fresh air conventional HVAC, but it provides a better indoor air quality.


The financial support of the ASHRAE (Project RP-1483), the Lebanese National Council for Scientific Research, Kuwait University Research Board and Easa Hussain Al-Yousifi & Sons Co. are highly acknowledged.


ASHRAE Standard 62.2, 2007. Ventilation and Acceptable Indoor Air Quality in Low-Rise Residential Buildings, American Society of Heating Refrigeration and Air conditioning Engineers, (ANSI approved).

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Carrier Hourly Analysis Program, Carrier, Version 4.20a, 2006.

Code of Practice No. MEW/R-6, 1994. Ministry of Electricity & Water, State of Kuwait.

Ghali K., N. Ghaddar, and M. Ayoub. 2007. Chilled ceiling and displacement ventilation system: an opportunity for energy saving in Beirut. International Journal of Energy Research, 31: 743-759.

Ghaddar N., K. Ghali, R. Saadeh, and A. Keblawi. 2008. Design charts for combined chilled ceiling displacement ventilation system (1438-RP). ASHRAE Transactions, 143(2):574-87.

Jiang W. and T. A. Reddy. 2007. General methodology combining engineering optimization of primary HVAC&R plants with decision analysis methods-part 1: deterministic analysis. HVAC&R Research, 13(1): 93-118.

Mesheshwari, G.P., F. Al-Ragom, H. Al-Taki, A. Mirza, and R. Alasseri. 2005. Optimization of lean season strategies in air-conditioned office buildings. Proceedings of the Third International Conference on Energy Research and Development, Kuwait, Nov. 20-23, 2005, pp. 115-129.

Novoselac A. and J. Serbric. 2002. A critical review on the performance and design of combined cooled ceiling and displacement ventilation systems. Energy and Build ings, 34: 497-509.

TRNSYS, A Transient Simulation Program, Version 15. 2004, Solar Energy Laboratory, University of Wisconsin-Madison, Madison, USA.

Yuan, X., Q. Chen, L. Glicksman. 2001. A critical review of displacement ventilation, ASHRAE Transactions. 4101 (RP-949): 78-90.

Ammar Bahman

Associate Member ASHRAE

Walid Chakroun, PhD


Nesreen Ghaddar, PhD


Ralph Saade

Kamel Ghali, PhD

This paper is based on findings resulting from ASHRAE Research Project RP-1438.

Ammar Bahman is a graduate student and Walid Chakrou is a professor at Kuwait University, Kuwait City, Kuwait, Nesreen Ghaddar is a professor and Ralph Saadesh is a graduate student at the American University of Beirut, Beirut, Lebanon, Kamel Ghali is a professor at Beirut Arab University, Beirut.
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Author:Bahman, Ammar; Chakroun, Walid; Ghaddar, Nesreen; Saade, Ralph; Ghali, Kamel
Publication:ASHRAE Transactions
Article Type:Report
Geographic Code:7KUWA
Date:Jan 1, 2009
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