Índice absoluto de marginación

This chapter studies the heat transfer principles of window mullion heating and cooling. Two heat transfer models have been set up through differential analysis. The simulation results have been compared with 10 day’s measured data. The comparison shows that the heat transfer models predict the measured temperatures with a root mean square error (RMSE) of the hot water return temperature, the mullion surface temperature, and the window surface temperature of 0.90°F, 0.98°F and 1.15°F, respectively. The simulation study leads to the following conclusions:

Hot water supply temperature and chilled water supply temperature are the primary factors that affect the heating or cooling capacity of window mullions and the mullion surface

temperature. Hot/chilled water return temperature, mullion surface temperature and window frame temperature are all quasi-linear functions of the hot water supply temperature.

Window surface temperature distribution is affected by the mullion surface temperature and the outside air temperature. The temperature gradient on the glazing surface within one foot from mullions is much higher than in the central part of the window. The temperatures in the central 2 feet of a 4-foot window show almost no influence by the mullion surface temperature.

The sensitivity study in the discussion section showed that the conductive thermal resistance of the mullion double tubes plays a decisive role in controlling the mullion and window frame temperatures. The double tubes and gap fillings increase the thermal resistance of the mullion tubes, which results in a lower surface temperature for heating and a higher surface temperature for cooling. The higher surface temperature for cooling may be intended to lower the risk of moisture condensation on the surface of the mullion in the cooling condition. However, the enhanced thermal resistance decreases the heating and cooling capacity of the mullion. If the mullions are only used for heating, a single tube structure is recommended.

Window frame thermal resistance affects the frame surface temperature and heat loss from the mullion to the outside air. The change of frame surface temperature and the heat loss to the outside air is a non-linear function of the change of framed thermal resistance, as expected. At a hot water temperature of 125oF, the frame temperature drops about 2.6°F, when frame resistance decreases from 2.0( ft2-hr-°F )/Btu to 1.0( ft2-hr-°F )/Btu. The frame temperature increases about 0.9°F when frame resistance increases from 2.0( ft2

-hr-°F )/Btu to 3.0( ft2-hr-°F )/Btu. The smaller the window frame thermal resistance, the larger the temperature drop from the mullion tube to the frame and the higher the heat loss from the window frame to the outside.

From a design perspective, the window width or spacing between the mullions has little impact on the heating capacity or mullion surface temperature. However, the space between the mullions will somewhat affect the window’s inner surface temperature distribution and the window’s average temperature. When the window width decreases from 5 feet to 3.5 feet, the average window surface temperature increases from 62.32°F to 62.73°F at an ambient temperature of 38°F. Increasing the mullion fin length will increase the heating or cooling capacity of the mullion because of the heat transfer area increase. If the fin length increases from 45mm (1.83 inchs) to 60mm (2.36 inches), the heating capacity of one mullion will increase about 3.1%.

The effect of solar radiation on the temperature distribution of the window panes depends upon window orientation, building location and season of the year. For a south-facing window under the sun on a typical winter day, the inside glass temperature will increase about 1°F when solar radiation is considered.

CHAPTER III

THE PERFORMANCE STUDY OF OVERHEAD RADIANT PANELS

3.1 Introduction

The overhead radiant panels are the other type of radiant device in the IW, as shown in Figure 1.2. This type of radiant panel is also called a free-hanging ceiling radiant panel. The overhead radiant panel can be used for both heating and cooling in the IW, and there is no topside insulation on these panels. The overhead radiant panels are supposed to meet a part of the sensible load of the IW, with panel output varied by controlling the supply water temperature. Significant research has been done regarding the heat transfer models, and the thermal comfort and efficiency of ceiling radiant panels. Chen and Kooi (1988) developed a radiant panel simulation model which considered the radiant ceiling panel as an indoor surface exchanging heat with room air by convection and other room surfaces through radiation. Kilkis et al. (1994) proposed an in-slab type panel model. They pointed out that the heat transfer in a panel-cooled room and the cooling panel itself might be represented by a quasi-steady state natural convection model by assuming uniform panel surface temperatures. Stetius and Feustel (1995) developed a 2-D radiant panel model by simplifying heat diffusion equations for an in-slab type panel. Conroy and Mumma (2000) derived an analytical model for a top insulated metal ceiling radiant panel. This model was based on the study of solar collectors conducted by Duffie and Beckman (1991). The basic methodology in this model was to determine the panel cooling capacity by finding the unknown mean panel surface temperature (Tpm) in an iterative process. However, the detailed structure of radiant panels varies greatly; it is hard to use one general model to estimate the ceiling panel capacity in the IW. The objective of this chapter is to develop a specific model to estimate the heating and cooling capacity of the radiant panels used in the IW (no topside insulation) with a focus on the impact of thermal contact resistance between the tubes and aluminum panels.

Some researchers (Awbi and Hatton 2000, Jeong and Mumma 2003b), in recent years, have proposed a mixed convection heat transfer coefficient to calculate the radiant panel capacity, because the ventilation diffusers near the ceiling panels create a forced air flow across the ceiling panels. However, the air velocity near the panel surfaces is related to the diffuser locations. In the IW, the ventilation diffusers are either on the occupants’ desks or on the floor, and thus have

little impact on the convection heat transfer across the radiant panels. This chapter will use the natural convection heat transfer coefficient recommended by ASHRAE 2004.