1、Kishor Khankari is the President at AnSight LLC in Ann Arbor, Michigan Analysis of Supply Airflow Rate on the Performance of Radiant Cooling Systems Kishor Khankari, PhD Fellow ASHRAE ABSTRACT Radiant cooling systems can potentially be more energy efficient than the traditional all air HVAC systems
2、partly due to their low supply air requirements. This paper with the help of Computational Fluid Dynamics (CFD) analysis compares the cooling performance of a traditional all air system with the radiant cooling systems for a typical perimeter office. The effect of supply airflow rate on the cooling
3、performance of both radiant floor and radiant ceiling systems is evaluated. The cooling performance is evaluated in terms of three-dimensional airflow patterns, temperature distribution, thermal comfort of occupants, and resulting cooling capacity of the radiant surfaces. These analyses indicate rad
4、iant cooling systems can provide comparable thermal comfort for occupants at much reduced airflow rates than the traditional all air system. However, the location of the radiant surface and associated supply airflow rates can significantly affect the thermal comfort and cooling capacity of the radia
5、nt systems. The radiant floor system may require higher airflow rate than the radiant ceiling system to provide comparable thermal comfort. However, increasing the airflow rate can significantly reduce the cooling capacity of the radiant floors than the radiant ceilings. Low airflow rates with radia
6、nt floor can create hot zones whereas high airflow rates with radiant ceiling can create cold zones for the occupants. Further the supply airflow rate can significantly affect the relative distribution between the convective and radiative heat transfer components of the radiant floors than the radia
7、nt ceilings. This study concludes that the radiant ceiling with low supply airflow rate can provide better thermal comfort and higher cooling capacity, and therefore, a better option for cooling perimeter offices than radiant floors. It is noted that CFD can be an effective tool in minimizing HVAC d
8、esign risks and gaining valuable insights which can help in optimizing the design of radiant cooling systems. INTRODUCTION Radiant systems primarily employ hydronic cooling and heating with reduced fan power due to reduced requirement of the supply air. Therefore such systems can be potentially more
9、 energy efficient than the traditional all air HVAC systems (Kwang and Olesen 2015, Sastri and Rumsey 2014, Thornton et al. 2009). Air is supplied to the occupied spaces with the radiant heating and cooling systems to maintain certain level of indoor air quality and humidity. Radiant heat transfer o
10、ccurs primarily between the surfaces such as walls, glazing, and radiant (hydronic) surfaces whereas the convective heat transfer occurs between the adjacent air and the radiant surfaces. Although in principle radiant heat transfer is supposed to provide more than 50 percent of the sensible heating
11、or cooling, adequate air movement adjacent to the radiant surfaces is essential for transferring the heat to and from the surfaces to the occupied spaces. In addition the temperature difference between the adjacent air and the radiant surfaces can also affect the cooling capacity of the radiant surf
12、aces. The distribution between the radiative and convective heat transfer to and from the radiant surfaces can depend on several factors including the location and type of sensible heat loads in the space, type and location of radiant system, type and location of supply air diffusers, supply air flo
13、w rate, and the temperature of the supply air. According to previous studies the radiant heat transfer in most cases can be limited to 0.97 Btu/hftF (5.5 W/mK) whereas the convective heat transfer can vary between 0.09 and 0.97 Btu/hftF (0.5 and 5.5 W/mK) which can depend on the surface type, surfac
14、e orientation, and on heating or cooling mode (Kwang and Olesen 2015). The main objective of this study is to compare the performance of a radiant cooling system with the all air overhead supply HVAC system for a small perimeter office using Computational Fluid Dynamics (CFD) simulations. Radiant fl
15、oor and radiant ceiling configurations are analyzed with two different supply airflow rates. The effect of supply airflow rates on the cooling capacity of radiant surfaces are analyzed along with the relative distribution of radiative and convective heat transfer components. The performance of these
16、 systems are compared by analyzing airflow patterns, temperature distribution, and the resulting thermal comfort of the occupants in a three dimensional office space. These analyses were performed systematically by varying a single parameter at a time. VIRTUAL OFFICE SET-UP A three-dimensional, stea
17、dy state, non-isothermal Computational Fluid Dynamics (CFD) model of a perimeter office is developed for this study. As shown in Figure 1 the virtual office has three occupants seated around a table with three computers, a window on one side of the room, a ceiling light, 4-way air supply diffuser, a
18、nd a return grill. The room has about 170 sq. ft. (15.8 m2) floor with 9 feet (2.74 m) ceiling height. The total sensible heat load in the room was assumed to be 19.7 Btu/h per sq. ft. (62 W/m2). The sensible heat load due to three occupants was assumed 4.5 Btu/h per sq. ft. (14 W/m2) whereas sensib
19、le heat load due to the computers and lighting was assumed to be 3.6 Btu/h per sq. ft. (11 W/m2) and 3.1 Btu/h per sq. ft. (10 W/m2), respectively. The room has South facing window with the solar heat gain of 8.6 Btu/h per sq. ft. (27 W/m2). These sensible cooling loads are part load design conditio
20、ns which are more prevalent than the peak design load conditions. The air is supplied through a 4-way diffuser at an angle of 15 degrees to the ceiling. This diffuser is placed close to the window such that one of the four supply air streams is directed towards the window for cooling the window glaz
21、ing. In the case of all air overhead system the total supply airflow rate and the supply air temperature were specified at 150 CFM (71 l/s, 6 ACH) and 55 F (13 C), respectively. In the case of floor and ceiling radiant systems the supply air flow rates were reduced to 25 percent and 50 percent of th
22、e supply airflow rate for the all air system and the supply air temperature was maintained at 55 F (13 C). It was ensured that as per ASHRAE Standard 62.1 the supply airflow rate is always maintained above the typical minimum ventilation airflow requirement of 0.15 cfm/ft2 (1.3 L/s.m2) for the offic
23、e spaces. According to ASHRAE Standard-55 occupants may feel uncomfortable due to contact with the cold floor, and therefore, for rooms with sedentary occupants the recommended minimum floor temperature for cooling is 66 F (19 C) (ASHRAE, 2013b). Accordingly in the case of radiant floor system the f
24、loor temperature was specified at 70 F (21 C) to ensure the surface temperature is not too close to the recommended minimum temperature. Whereas in the case of radiant ceiling system the ceiling temperature was specified at 65 F (18 C). This study assumes the room design set point temperature is 75
25、F (24 C) and the relative humidity is 55 percent which corresponds to the dew point temperatures of 58 F (14 C). The supply air conditions were assumed to be 55 F (13 C at 60% relative humidity (41.4 F (5 C) dew point temperature). It is further assumed that the radiant systems can control and maint
26、ain the constant temperature of the radiant surfaces and the supply air conditions can maintain the humidity levels in the space. Thermal comfort of occupants was analyzed by employing Predicted Mean Vote (PMV) index as described in ASHRAE Fundamentals Handbook (ASHRAE 2013). This index was computed
27、 assuming CLO values of 0.5 and the metabolic heat production rate (MET) of 1.2., and relative humidity of 55 percent. The walls of the room other than the glazing and the radiant surface were assumed to be adiabatic. The standard k-e turbulence model was employed to compute the turbulent viscosity
28、of the air. Discrete-ordinates radiation model was employed for radiation calculations with angular theta (N) and phi (N) divisions specified to one which controls the angles used to discretize each octant of the angular space. A total of 5 cases analyzed in this study are described in the Table 1.
29、Table 1: Description of the cases analyzed Case HVAC System Supply airflow rate Percent of the all air system Supply air temperature 1 All air 150 CFM (71 l/s) 100 55 F (13 C) 2 Radiant floor 75 CFM (35.4 l/s) 50 55 F (13 C) 3 Radiant floor 37.5 CFM (56.5 l/s) 25 55 F (13 C) 4 Radiant ceiling 75 CFM
30、 (35.4 l/s) 50 55 F (13 C) 5 Radiant ceiling 37.5 CFM (56.5 l/s) 25 55 F (13 C) Figure 1 Schematic diagram of CFD models of a perimeter office: a) all air overhead system, b) radiant floor, and c) radiant ceiling system. RESULTS AND DISCUSSION Figure 2 shows computational results for the all air ove
31、rhead system. Such systems are also known as mixed air systems which potentially promote mixing of the room air to maintain uniform air temperature in the space. In this case the supply air from the 4-way diffuser is split into four directions three streams flow into the room over the occupants and
32、computers whereas the fourth stream sweeps a part of the hot window glazing. As shown in Figure 2a, this air stream after sweeping the window flows over the room floor and then moves upward towards the room return. The other air streams fall into the room, flow over the occupants and computers while
33、 gaining the heat from these sources and then rise upward due to buoyancy forces. It also shows a small air recirculation pattern as the air moves around the ceiling light. The actual three-dimensional airflow patterns (not shown here) are more complex than shown in Figure 2a which are drawn on a si
34、ngle vertical plane. The resulting air temperature distribution as shown in Figure 2b indicates the room is divided into two temperature zones a lower temperature zone behind the occupant on the side of the window and a higher temperature zone in the rest of the room. As mentioned before the cold su
35、pply air after gaining the heat from all heat sources moves upward towards the return which creates such a zone of high temperature in this section of the room. Figure 2c shows resulting distribution of PMV, the thermal comfort index, at the 4 feet (1.2 m) height from the floor within a range of -0.
36、5 (cold) to 0.5 (hot) in which according to ASHRAE 55 standard about 90 percent of the occupants would experience acceptable thermal environment (ASHRAE 2013). Although this analysis indicates that all air HVAC system can provide mostly thermal neutral (PMV=0) comfortable environment around the occu
37、pants, it could also form a few isolated cold regions near the side walls of the room. It appears that the cold air exiting from the sides of the 4-way diffuser Lap t opsRetu rnLight Supply dif f us e r 4 -w ayW indo w(a)Supply dif f us e r 4 -w aySupply airf low rat e 50%R adiant f loor (70F , 21C
38、)(b)Supply dif f us e r 4 -w aySupply airf low rat e 50%R adiant c eiling (65F , 18C )(c)pass straight through this region without gaining heat from any heat sources. A combination of high air velocities and low air temperature in this region could create a cold sensation for the occupants. Also not
39、e in the region where cold air sweeps at the center of the window the PMV values are in a comfortable range than at either end of the window. Figure 2 Results for the all air overhead supply analysis showing a) airflow patterns, b) temperature distribution, c) PMV distribution. Figure 3 shows airflo
40、w patterns for the floor and ceiling radiant systems. It should be noted the supply air flow rates are significantly reduced in these analyses. In contrast to the all air system due to reduction in the cooling capacity of the supply air the air temperatures in the space increase and the buoyancy eff
41、ect becomes prominent. As shown in these figures the hot air in the vicinity of the occupants and computers rises upward and mixes with the incoming cold supply air stream near the ceiling. As a result the temperature of this mixed air stream decreases which causes it to slide downward along the wal
42、l and then moves along the floor back towards the heat sources. Thus, it creates a large recirculation airflow patterns between the wall and the two occupants. Figure 3 Results for the radiant floor and radiant ceiling analysis showing airflow patterns. (a) (b) (c)a) f loor : airf lo w rate 25%b) an
43、t f loor : airf lo w rate 50%c ) eiling: air f lo w rate 25%d) eiling: airf lo w rate 50%In the case of radiant floor systems a smaller recirculation zone is also formed under the light (Figure 3a and 3b). In the case of ceiling radiant systems the cold supply air moves along the cold ceiling withou
44、t creating such a circulation zone under the light (Figure 3c and 3d). In these cases another air recirculation pattern is created near the window. The hot air rising near the window surface and the falling cold air stream form this recirculation pattern. These analyses show a little difference in t
45、he overall airflow patterns between the 25 percent and 50 percent reduced airflow rates. Figure 4 shows resulting air temperature distribution in the room with radiant cooling systems. In general air temperatures in the room are higher than in the case of all air systems. The low temperature zone as
46、 mentioned above in the case of all air system significantly diminishes in both cases of radiant systems and it remains mostly in the vicinity of the supply diffuser. In the case of radiant floor system with 25 percent reduced airflow rate (37 CFM (56.5 l/s), Figure 4a), a zone of high air temperatu
47、re is formed in the vicinity of the computer and occupants which extends all the way up to the ceiling. Increasing the airflow rate (50 percent reduced airflow rate) the height of this high temperature zone is reduced which remains closer to the table top. Increasing the supply airflow rate from 25
48、percent to 50 percent reduces the overall air temperature in the room (Figure 4a and 4b). In the case of radiant ceiling with low supply airflow rate the zone of higher air temperature extends up to 5 feet (1.5 m) from the floor whereas a zone of low temperature is formed near the ceiling. In the ca
49、se of radiant ceiling systems increasing the supply airflow rate from 25 to 50 percent significantly reduces the overall air temperature in the room (Figure 4c and 4d). Figure 4 Results for the radiant floor and radiant ceiling analysis showing temperature distribution. Figure 5 shows resulting distribution of thermal comfort index (PMV) at the 4 feet (1.2 m) height from the floor. The analysis in the case of radiant floor system with low airflow rate indicates the occupants may feel hot in the vicinity of the window and near the other sources of heat the in room (Figure 5a). T
