Analysis on combinations of indoor thermal microclimate parameters in radiant cooled residential buildings and drawing of new thermal comfort charts

Article Analysis on combinations of indoor thermal microclimate parameters in radiant cooled residential buildings and drawing of new thermal comfort charts Building Serv. Eng. Res. Technol. 206, Vol. 37() 66 84! The Chartered Institution of Building Services Engineers 205 DOI: 0.77/04362445596924 bse.sagepub.com Xuemin Sui and Xu Zhang 2 Abstract This paper proposed to discuss some issues about indoor microclimate parameters design in radiant cooled residential buildings. Based on the thermal comfort model, the relationships between low mean radiant temperature and other thermal microclimate parameters such as indoor air temperature, air velocity, and relative humidity were analyzed. The improved and reasonable combinations of four thermal microclimate parameters including radiant cooling surface temperature, air temperature, air velocity, and relative humidity were further derived. Furthermore, a number of new thermal comfort charts using radiant cooling surface temperature and air temperature as dominant parameters were drawn. The results show that the air temperature can be increased 3 C within the recommended thermal comfort range when the mean radiant temperature is reduced 3 C. Under the design goal of thermal neutral, when the chilled ceiling surface temperature is in the range of 6 26 C, the air temperature should be set at 25.3 29.3 C; when the chilled floor surface temperature is in the range of 9 26 C, the air temperature should be set at 25.8 29.2 C. The design values of air velocity and relative humidity are not recommended to be changed, and still can follow the design standards for conventional all-air systems. Practical application: In this paper, some issues about indoor microclimate parameters design in radiant cooled buildings were discussed. A number of new thermal comfort charts using radiant cooling surface temperature and air temperature as dominant parameters were provided, and the design values of indoor air temperature, mean radiant temperature, air velocity, and relative humidity for radiant cooling system in residential buildings were also suggested. The research results can provide a more intuitive reference for the design of radiant cooling systems in residential buildings. Keywords Residential building, radiant cooling, thermal comfort, thermal microclimate parameter, thermal comfort chart School of Environmental Science and Engineering, Chang an University, Xi an, PR China 2 HVAC&Gas Institute, School of Mechanical Engineering, Tongji University, Shanghai, PR China Corresponding author: Xuemin Sui, School of Environmental Science and Engineering, Chang an University, Xi an 70054, China. Email: suixuemin@63.com

Sui and Zhang 67 Introduction Radiant cooling system was first seriously investigated in European countries in the late 980s, and has become one of the popular design alternatives for space cooling in north-west Europe. In the early 2000s, radiant cooling systems began to be used in China. Since then, this technology has been increasingly applied in public buildings, 2 4 and has received increasing attention. Compared with conventional all-air systems, radiant cooling systems exhibit the superiority of thermal comfort and energy saving potential. The advantages of this system are as follows: 5 0 () reducing indoor vertical temperature gradient, almost no air flow, avoiding local discomfort; (2) decrease in air supply volume reduces cooling feeling due to cold air, and thus reduces the discomfort caused by draft; (3) reduction in air supply volume minimizes air transport energy consumption; (4) employing water to eliminate heat load instead of air greatly reduces energy consumption of cooling transport; (5) radiant cooling reduces the sensing temperature of human body and accordingly reduces system energy consumption; (6) high temperature of chilled water allows the use of natural resources and natural cooling in some seasons; (7) reducing indoor noise. In recent years, the indoor thermal environment and energy consumption problems of China s urban residential buildings have become severe. Summer air-conditioning devices have become increasingly prevalent, mainly dominated by convective air-conditioners. Of course, split air-conditioning unit will give better control of indoor air temperature and then improve human comfort and productivity. But there also exist many examples of discomfort due to draught, uniform temperature, noise, and sick building syndrome. On the other hand, the rapid increase of China s urban residential area objectively causes the rapid growth of residential building energy consumption. 2 Due to the shortcomings of current residential air-conditioning system, there is an urgent need for a new air-conditioning system to replace the traditional system. Many heating, ventilation, and air conditioning (HVAC) engineers and occupants turned their eyes to radiant cooling system, and it began to be gradually used in residential buildings of China. Currently, the residential applications of radiant cooling cases are mainly the technical systems of radiant cooling combined with ground-source heat pump, and the district buildings are centralized cooled. 3 5 However, centralized cooling cannot achieve occupants flexible adjustments and cannot meet occupants individual needs. In view of the above problems, the authors and their research team conducted a lot of studies on residential temperature humidity separate control air-conditioners. 6 2 These airconditioners used air source heat pump as the hot and cold sources, used radiant floor for heating in winter, used radiant ceiling for cooling in summer, and also employed outdoor air supply system to meet both indoor health requirements and dehumidification needs. Air-conditioning demand and air-conditioning usage patterns of residential buildings are different from that of public buildings. For the applications and promotions of radiant cooling systems in residential buildings to achieve good thermal comfort and energy saving, there are many design questions to be solved, including indoor thermal environment design, flexible system control, anti-condensation control, etc. In this study, we investigated the indoor thermal microclimate parameters design of residential buildings with radiant cooling systems. The determination of indoor design parameters is the first step of load calculation process in designing any HVAC system. There are four basic thermal microclimate parameters affecting human thermal sensations including indoor air temperature, relative humidity of indoor air, air velocity near human body, and mean radiant temperature. 22 The thermal comfort zone of the ASHRAE standard 55-200 23 does not directly give the design ranges of indoor air temperature and mean radiant temperature, but it recommends the design range of operative temperature, which combines the effects of air temperature and mean radiant temperature.

68 Journal of Building Services Engineering Research & Technology 37() Relevant specifications in early Chinese design codes and manuals commonly take the first three parameters as the indoor design parameters, which based on the assumption that indoor environment is uniform, and indoor air temperature is equal to mean radiant temperature. 24 26 In the conventional air-conditioning design, there is no distinction between indoor air temperature and operative temperature or between indoor air temperature and indoor surface temperature in wall conduction load calculation. But this idea can not be used for radiant cooling systems, because indoor air temperature is not equal to mean radiant temperature in radiant cooled environment. Xuemin Sui et al. 9,27 took a large number of field measurements on the practical application of radiant cooling systems in residential buildings and evaluated the indoor thermal environment. The field measurement data showed that the mean radiant temperature was about 3 C lower than the air temperature. It is concluded that the mean radiant temperature is reduced due to the presence of radiant terminal device, and its value is lower than indoor air temperature. If the design values and ranges of the microclimate parameters combinations for traditional all-air systems were used for radiant cooling systems, it is bound to cause a certain bias between the design value and the actual value of human thermal comfort index. However, so far, the current radiant standards do not clearly give the indoor design thermal environment parameters for radiant cooling system. The international radiant cooling/heating standard ISO 855 28 pointed out that compared with a convective heating and cooling system, a radiant heating system could achieve the same level of operative temperature at a lower air temperature and a radiant cooling system at a higher air temperature, but it did not give quantitative analysis and quantitative values. The European radiant cooling/heating standard EN 5377 29 pointed out that the operative temperature should be considered as relevant for thermal comfort assessment and heat loss calculations. But for current cooling load calculation method, indoor air temperature is needed for wall conduction load calculation. The new Chinese design code GB 50736-202 30 supplemented the design air temperature range for radiant cooling system based on the design temperature of 24 26 C for civil buildings with convective air-conditioning system, and proposed the indoor design air temperature should increase 0.5.5 C. However, this specification is not comprehensive enough, because low mean radiant temperature can also be compensated by some other measures to achieve the same comfort, such as changing the air velocity or the relative humidity. Therefore, for radiant cooling system, it is necessary to further combine and design the four basic microclimate parameters and to determine the appropriate combinations of them within the thermal comfort range. In addition, this provision can not provide a precise reference for designers because the increase of indoor design air temperature depends on the value of mean radiant temperature. Under what conditions should the indoor design temperature be increased 0.5 C, and under what conditions should be increased.5 C? This question does not be clearly solved. However, even if the difference is only C, it still has a non-negligible effect on building energy consumption. Previous research show the indoor design air temperature is increased by C, energy consumption can be reduced by 8 0% in the cooling conditions. 3 For different application cases, the mean radiant temperature values are different and difficult to be evaluated. The reduction of mean radiant temperature is mainly caused by the presence of radiant cooling panel surface, if some new thermal comfort charts taking radiant cooling surface temperature as dominant parameter could be drawn, it would provide a more intuitive reference for the design and application of radiant cooling system. The purpose of this study is to address these issues. In the present article, based on the thermal comfort model, the thermal microclimate parameters combinations in residential buildings were derived within the recommended thermal comfort range. The relationships between low

Sui and Zhang 69 mean radiant temperature and other thermal microclimate parameters including indoor air temperature, air velocity, and relative humidity were analyzed. Then the reasonable combinations of four microclimate parameters including radiant cooling surface temperature, indoor air temperature, air velocity, and relative humidity were further derived. In addition, a number of thermal comfort charts using radiant cooling surface temperature and air temperature as dominant parameters were drawn, and the reasonable range of the design value of each parameter was determined. The research results can provide a more intuitive reference for the design of radiant cooling system in residential buildings. Thermal comfort model Thermal comfort model is the base of thermal comfort evaluation methods and thermal environment standards. The purpose of this section is to describe the selection of thermal comfort model used in the study. According to the simplified method of human thermal system, there are mainly three types of thermal comfort models including single-node model, two-node model, and multi-node model. 32 In addition, there is another model called adaptive model mainly for non-airconditioned buildings. 33 The Predicted Mean Vote/Predicted Percentage Dissatisfied (PMV/ PPD) model is the most widely used singlenode model at present. Based on human thermal equilibrium theory, it is mainly used to predict human thermal responses and evaluate the overall thermal sensation in steady-state thermal environment controlled by air-conditioning systems. 34 The two-node and multi-node models all consider the physiological parameters and thermoregulatory mechanisms of human body, and they can be used to predict human thermal comfort in dynamic and non-uniform environments. 32 The so-called adaptive model is mainly used for naturally ventilated buildings. 35 37 It is not actually used to predict people s comfort responses, but is used to predict thermal environment parameters which can make people feel comfortable. It reflects the indoor comfort temperature variation trends along with the mean monthly outdoor air temperature. As for the determination of indoor design thermal parameters in air-conditioning system design, it is based on such a steady-state thermal environment well controlled by air-conditioning system. A lot of standards such as ISO 7730, 38 ASHRAE Standard 55-200, 23 Chinese Standard GB 5009-2003, 24 JGJ34-200, 25 and GB 50736-202, 30 all use the PMV/PPD indices to describe and evaluate thermal environment, and also give the suggested design parameters of indoor thermal environment based on these indices. However, those provisions are aimed at conventional convectional air conditioning systems. The purpose of this study is to determine the design thermal environment parameters of residential buildings which employ radiant cooling systems. The thermal environment is well controlled and belongs to steadystate thermal environment, so the PMV/PPD model was also be used to derive the reasonable indoor thermal microclimate parameters combinations in the paper. The PMV/PPD model has been validated by many laboratory and field studies around the world since many years ago. 39 43 Besides, early researchers conducted a large number of experimental studies on thermal comfort in radiant cooled environment and proposed that the PMV/PPD model was also applicable to radiant cooled environment. 44 Factors that affect the PMV are metabolic rate, clothing insulation, air temperature, mean radiant temperature, air speed, and relative humidity. PPD is an index expressing the thermal comfort level as a percentage of thermally dissatisfied people, and is directly determined from PMV. Much more details including calculation methods of PMV and PPD are described in ISO 7730. 38 It recommends the values of PMV/ PPD indices to be adopted: 0.5 PMV 0.5; PPD 0%. The main methodology of the paper is using MATLAB programming for determining combinations of microclimate parameters, based on the PMV/PPD model.

70 Journal of Building Services Engineering Research & Technology 37() When considering the dynamic or uneven effects of thermal environment, the two-node model or multi-node model should be selected. As for the issue on human thermal comfort in non-uniform radiant cooled thermal environment, it will be addressed in the follow-up studies. Effect of low mean radiant temperature on thermal comfort In this section, based on the PMV/PPD model, 38 programming with MATLAB, the effect of low mean radiant temperature on human thermal comfort was analyzed. The following calculating background conditions were considered: () indoor human activities commonly are sitting and resting, and the body metabolism rates are met; (2) residents generally are used to wearing shorts, blouses, pajamas, slippers, and other household clothing in summer, which have low insulation values in range of 0.25 0.65 clo. In order to correspond to the summer condition of ASHRAE 55-200, 23 the I cl ¼ 0.5 clo condition was selected for analysis; (3) the indoor relative humidity is 50%, and the air velocity is 0. m/s. In practical applications, the difference between mean radiant temperature and indoor air temperature is often used to reflect the relative relationship between these two parameters. Figure shows the variation trends of the PMV index along with the temperature difference (4t, 4t ¼ t r t a ). It can be seen that, under different constant air temperature conditions, the variation of temperature difference causes a significant change of the PMV index. For example, under the condition with the indoor air temperature of 26 C, when the temperature difference is 0 C, the PMV index is 0.03, which means human thermal sensation is neutral. But when the temperature difference is 3 C, the PMV index is 0.55, which means human thermal sensation is slightly cool, and this value has exceeded the thermal comfort standard limit. 38 It is concluded that low mean radiant temperature makes the PMV index lower, and causes a cold sensation of human body. So the comfort environment design of radiant cooled room should consider not only the indoor air temperature but also the effect of mean radiant temperature. The design of indoor thermal microclimate parameters should attach importance to mean radiant temperature, and re-analyze the 3 ta=24 C ta=26 C ta=28 C ta=30 C 2 PMV 0 - -2-3 -8-7 -6-5 -4-3 -2-0 Δt ( C) Figure. PMV index variation trends along with the difference between mean radiant temperature and air temperature.

Sui and Zhang 7 (a) 40 PMV=-0.5 PMV=-0.2 PMV=0 PMV=0.2 PMV=0.5 (b) 40 PMV=-0.5 PMV=-0.2 PMV=0 PMV=0.2 PMV=0.5 Air temperature ( C) 35 30 25 Air temperature ( C) 35 30 25 20 6 7 8 9 20 2 22 23 24 25 26 27 Mean radiant temperature ( C) 20 6 7 8 9 20 2 22 23 24 25 26 27 Mean radiant temperature ( C) Figure 2. Air temperature variation trends along with low mean radiant temperature within thermal comfort range. (a) I cl ¼ 0.3 clo; (b) I cl ¼ 0.5 clo. appropriate combinations of indoor microclimate parameters. Relationships between low mean radiant temperature and other thermal microclimate parameters The analysis in Effect of low mean radiant temperature on thermal comfort section shows that the mean radiant temperature has a significant impact on the PMV index, and the decrease of mean radiant temperature caused a decrease of it. Whether the effects caused by low mean radiant temperature can be compensated by other thermal microclimate parameters such as air temperature, air velocity, and relative humidity? It was discussed based on MATLAB program for the PMV-PPD model. The relationships between low mean radiant temperature and these other thermal microclimate parameters were analyzed by the corresponding derivation procedure, respectively. Relationship between low mean radiant temperature and indoor air temperature Figure 2 shows the air temperature variation trends along with low mean radiant temperature within the thermal comfort range under the condition of v ¼ 0. m/s. It can be seen from Figure 2(a) and (b) that air temperature should be increased to reach the same thermal comfort when mean radiant temperature decreases. That means low mean radiant temperature can be compensated by higher air temperature. Neutral temperature is the temperature at which human body feels neither hot nor cold. 23 At this temperature, human body does not need to take additional physiological regulation to maintain its thermal equilibrium. In a thermal neutral condition, the value of PMV is 0. Whether in climate chamber experiments or in on-site investigations, neutral temperature is most often studied as an indicator. 45,46 So in this article, the relationship between low mean radiant temperature and air temperature was discussed when thermal sensation of human body was in thermal neutral or thermal comfort. Figure 2(a) shows the condition of I cl ¼ 0.3 clo. Under the thermal neutral conditions, when the mean radiant temperature is 26 C, the corresponding air temperature is 28.3 C; when the mean radiant temperature is 23 C with a reduced value of 3 C, the corresponding air temperature is 3 C with an increasing value of 2.7 C; when the mean radiant temperature is 20 C with a reduced value of 6 C, the corresponding air temperature is 33.6 C with an

72 Journal of Building Services Engineering Research & Technology 37() increasing value of 5.3 C. Figure 2(b) shows the condition of I cl ¼ 0.5 clo. Under the thermal neutral conditions, when the mean radiant temperature is 26 C, the corresponding air temperature is 26.2 C; when the mean radiant temperature is 23 C with a reduced value of 3 C, the corresponding air temperature is 28.9 C with an increasing value of 2.7 C; when the mean radiant temperature is 20 C with a reduced value of 6 C, the corresponding air temperature is 3.5 C with an increasing value of 5.3 C. Thus, it can be concluded that the design value of indoor air temperature for radiant cooling systems increases compared to all-air systems, and the increase level depends on the decrement of the mean radiant temperature. The increment of the design air temperature recommended by the current Chinese design code GB 50736-202 30 is low when used for residential buildings. The recommended increment should be increased by.2 2.2 C when the mean radiant temperature is about 3 C lower than the air temperature. Relationship between low mean radiant temperature and air velocity Figure 3 shows the air velocity variation trends along with low mean radiant temperature within the thermal comfort range at constant air temperature of 26 C and constant relative humidity of 50%. Figure 3(a) and (b) shows the conditions of I cl ¼ 0.3 clo and I cl ¼ 0.5 clo, respectively. As can be seen from the two figures, each curve has an inflection point. These inflection points have zero air velocities. The position of inflection point depends on the PMV value, and it decreases with decreasing PMV value. On the right of inflection point, in order to reach the same thermal comfort, the lower the mean radiant temperature is, the lower the air velocity is. While on the left of inflection point, it shows an opposite variation trends that the lower the mean radiant temperature is, the higher the air velocity is. For example, in Figure 3(a), under the thermal neutral conditions, the mean radiant temperature of 20 C results in an air velocity of 0.02 m/s to reach the same thermal comfort, while the mean radiant temperature of 6 C results in an air velocity of 0.3 m/s. However, in actual situations, lower mean radiant temperature combining with higher air velocity will increase the cold sensation of human body. In this case, the solution derived from the thermal comfort model apparently departs from practical common sense. From the above discussion, it can be concluded that the effect of the reduction of mean (a) 0.4 PMV=-0.5 PMV=0 PMV=0.5 (b) 0.4 PMV=-0.5 PMV=0 PMV=0.5 Air velocity (m/s) 0.3 0.2 0. Air velocity (m/s) 0.3 0.2 0. 0 6 7 8 9 20 2 22 23 24 25 26 27 Mean radiant temperature ( C) 0 6 7 8 9 20 2 22 23 24 25 26 27 Mean radiant temperature ( C) Figure 3. Air velocity variation trends along with low mean radiant temperature within thermal comfort range. (a) I cl ¼ 0.3 clo; (b) I cl ¼ 0.5 clo.

Sui and Zhang 73 (a) Relative humidity (%) 450 400 350 300 250 200 50 00 50 PMV=-0.5 PMV=0 PMV=0.5 (b) PMV=-0.5 PMV=0 PMV=0.5 350 Relative humidity (%) 0-50 6 7 8 9 20 2 22 23 24 25 26 27 6 7 8 9 20 2 22 23 24 25 26 27 Mean radiant temperature ( C) Mean radiant temperature ( C) 300 250 200 50 00 50 0 Figure 4. Relative humidity variation trends along with low mean radiant temperature under the thermal neutral condition. (a) I cl ¼ 0.3 clo; (b) I cl ¼ 0.5 clo. radiant temperature on thermal sensation can be compensated by lower air velocity. But the air velocity can only be reduced within a limited range. When air velocity drops to 0, if the mean radiant temperature continues to drop, it needs to be compensated by higher air velocity, which is contrary to the traditional common sense of thermal sensation. The reason for this result is that Fanger s thermal comfort model can only predict thermal sensation in a limited range. Outside this range, it may lead to a disagreement between the equation solution and common sense. Accordingly, the conclusion can be drawn that in order to reach the same comfort, low mean radiant temperature can be compensated by lower air velocity within a limited range. Relationship between low mean radiant temperature and relative humidity Figure 4 shows the relative humidity variation trends along with low mean radiant temperature within thermal comfort range at constant air temperature of 26 C and air velocity of 0. m/s. It can be seen from the figure, at different values of PMV, the relative humidity increases with the decrease of mean radiant temperature. However, the capability of using higher relative humidity to compensate for the effect of low mean radiant temperature is limited, because the value of relative humidity cannot be unlimitedly increased. Taking the thermal neutral condition of I cl ¼ 0.5 clo for example, a mean radiant temperature of 27 C corresponds to a relative humidity of 34.7%, while a mean radiant temperature of 24 C corresponds to a relative humidity of 94.6%. When the mean radiant temperature continues to drop below 24 C, the relative humidity value solved by the model equation has exceeded the maximum admissible value, and therefore the relative humidity can not be increased to compensate for the effect of low mean radiant temperature. It also can be seen from the figure that the compensating capability of relative humidity improves with the decrease of PMV. Using high relative humidity to compensate for low mean temperature can play a positive role in reducing energy consumption of air conditioning system. Especially in some high humid regions, increasing indoor design relative humidity can reduce dehumidification load, and thus reduce energy consumption of air conditioning system. But for indoor relative humidity design, in addition to considering the human thermal comfort, another important issue is the condensation of radiant cooling panel. Anti-condensation of radiant panel restricts the design range of indoor relative humidity. When the indoor control parameters such as air temperature and relative humidity are constant, the common

74 Journal of Building Services Engineering Research & Technology 37() Table. Dew point temperatures under different relative humidities at a constant air temperature of 27 C. Dry bulb temperature ( C) Relative humidity (%) 27 50 5.70 27 60 8.58 27 70 2.06 27 80 23.25 27 90 25.22 27 00 27.00 anti-condensation measure is to control the water supply temperature of radiant panel to keep the panel surface temperature higher than the dew point temperature of indoor air. Table shows the values of dew point temperature under different relative humidities at a constant air temperature of 27 C. High relative humidity results in a high dew point temperature, which limits the surface temperature of radiant cooling panel. It can be seen from the table that the control surface temperature has been close to indoor air temperature when relative humidity is very high. For these conditions, the use of radiant cooling has lost its practical significance. Therefore, despite the conclusion that low mean temperature can be compensated by high relative humidity is significant to be recommended in view of climate characteristics of high humidity regions, the design range of relative humidity should not be increased due to the limited range of relative humidity for compensating, coupled with the constraints to relative humidity due to anti-condensation of radiant cooling panel. Improved combinations of parameters and new thermal comfort charts Dew point temperature ( C) According to the analysis in Relationships between low mean radiant temperature and other thermal microclimate parameters section, to reach the same thermal comfort, low mean radiant temperature can be compensated by higher air temperature, but should not be compensated by changing air velocity and relative humidity. The reduction of mean radiant temperature is caused by the radiant cooling panel surface, so the thermal microclimate parameters that affect human thermal sensation in radiant cooled environment can be further summarized as: indoor air temperature, indoor air relative humidity, air velocity near human body, and radiant cooling surface temperature. If the combination relationship of these four parameters in thermal comfort could be found, it would provide a more intuitive reference for the design and application of radiant cooling system. The following section aims to address this issue. In order to quantify the relationship between radiant cooling surface temperature and mean radiant temperature, the composition of mean radiant temperature should be first analyzed. The mean radiant temperature can be calculated by equation (): 22 T 4 mrt,p ¼ XN i¼ F p i T 4 i ðþ where, F p i is the view factor between body (denoted as p) and plane (denoted as i); T i is the surface temperature (K). Equation () shows that the value of mean radiant temperature depends on the temperature of each wall as well as the view factor between human body and each wall surface. Therefore, in order to get the values of mean radiant temperature of different radiant cooling surface temperatures, the temperature of each interior wall surface of different radiant cooling surface temperatures should be firstly found. The wall surface temperatures and their distribution characteristics can be determined by experiments. To study the thermal environment characteristics of radiant cooled residential building, a lot of field measurements have been conducted. 9,27 The residential building tested is a three-layer linking row villa located in Shanghai City. The testing period was 20 August 2008 to September 205, during which the outdoor temperature was 7 34 C.

Sui and Zhang 75 36 Outdoor air Chilled ceiling surface Interior surface of south window Interior surface of west wall Interior surface of north wall Interior surface of east wall Floor surface Indoor air Mean radiant temperature 34 32 Temperature ( C) 30 28 26 24 22 :00 :09 :8 :27 :39 :48 :57 2:06 2:5 2:24 2:33 2:42 2:5 3:00 Time Figure 5. Indoor and outdoor air temperatures and interior wall surface temperatures. The layout of the measured building, the thermal properties of building envelopes, the tested radiant cooling system, the measurement equipment, and methodology have been elaborated in Sui et al. 9 Figure 5 shows the test data on indoor thermal environment at :00 3:00 on September 2008, which shows the relative relationship of the various parameters such as outdoor air temperature, indoor air temperature, interior wall surface temperatures, and mean radiant temperature. For the tested radiant chilled ceiling system, the cooling radiant terminal device was designed with a concrete core structure with pipes embedded in concrete slabs between each story, similar to radiant floor heating systems. The cooling surface covered the whole surface of the ceiling. In the testing phase, the chilled ceiling surface temperature is about 23 C, and the indoor air temperature is about 28 C. The south exterior wall is a large floor-toceiling window with a good exterior shading device. Its temperature is about 30.3 C due to the effect of outdoor air temperature. The surface temperatures of interior walls are lower than indoor air temperature due to the presence of chilled ceiling, and the mean radiant temperature is about 26.5 C. Due to the limitations of experimental conditions, the thermal storage of chilled ceiling, as well as the effect of outdoor meteorological parameters, the interior wall surface temperatures of different chilled ceiling surface temperatures under the same design condition cannot be obtained by experiments. This problem can be solved by thermal radiation heat transfer network. The geometry of the tested room is 4.6 m 3.4 m 2.8 m. Because the interior surface temperature of exterior wall mainly depends on the outdoor meteorological parameters, its value of different chilled ceiling surface temperatures under the same outdoor meteorological parameters condition can be considered as a

76 Journal of Building Services Engineering Research & Technology 37() E b3 ε3 ε A 3 3 X A, 3 J 3 X A 2,3 2 J J 2 E b ε ε A X A,2 ε 2 ε A 2 2 E b2 Figure 6. Radiation heat transfer network of a closed cavity with three gray surfaces. constant value. According to the experimental results, it is 30.3 C. The adjacent rooms are all air-conditioned rooms, so the temperature difference between the interior wall surface and the adjacent rooms can be negligible. As can be seen from the measured data, the interior surface temperatures of four interior walls have little difference. And further, because each interior surface has the same emitting rate, for simplicity, they may be regarded as equated with a wall with uniform surface temperature. Thus the radiant heat transfer between the wall surfaces of radiant cooled room can be simplified as a radiant heat transfer between three gray surfaces of a closed cavity including the exterior wall, the chilled ceiling, and the simplified interior wall. The radiation heat transfer network is shown in Figure 6. According to Kirchhoff s Current Law, the node equations of the radiation heat transfer network are listed as following: 47 Node : Node 2: Node 3: E b2 J 2 " 2 þ J 2 J þ J 3 J 2 ¼ 0 ð3þ " 2 A 2 X,2 A X 2,3 A 2 E b3 J 3 " 3 þ J J 3 þ J 2 J 3 ¼ 0 ð4þ " 3 A 2 X,3 A X 2,3 A 2 The net rate of radiation heat transfer from the interior wall surface can be calculated by equation (5): 3 ¼ E b3 J 3 " 3 ¼ 0 ð5þ " 3 A 3 Derived from the above equations, the calculation formula of the interior surface temperature of interior wall can be expressed by: E b J " þ J 2 J þ J 3 J ¼ 0 ð2þ " A X,2 A X,3 A T 2 3 ¼ X 3T 2 þ X 32T 2 2 T 2 3 ¼ 0:45T2 þ 0:55T2 2 ð6þ ð7þ

Sui and Zhang 77 Table 2. Comparison between experimental and calculated values. Outdoor air temperature ( C) Chilled ceiling surface temperature ( C) Interior wall surface temperature ( C) Experimental values 33.5 23.3 26.8 26.5 32.2 23.2 26.5 26.4 33. 23. 26.7 26.3 3.3 23.0 26.6 26.3 32. 23.0 26.6 26.3 33.6 23.4 26.6 26.5 Calculated values where A, A 2, and A 3 are the surface areas of the exterior wall, chilled ceiling, and interior wall, respectively (m 2 ); ", " 2, and " 3 are the surface emissivities of the exterior wall, chilled ceiling, and interior wall, respectively; E b, E b2, and E b3 are the emissive powers of the exterior wall, chilled ceiling, and interior wall, respectively (W/m 2 ); T, T 2, and T 3 are the surface temperatures of the exterior wall, chilled ceiling, and interior wall, respectively (K); J, J 2, and J 3 are the effective radiation rates of the exterior wall, chilled ceiling, and interior wall, respectively (W/m 2 ); X 2 is the view factor from the exterior wall to the chilled ceiling; X 3 is the view factor from the exterior wall to the interior wall; X 23 is the view factor from the chilled ceiling to the interior wall; X 3 is the view factor from the interior wall to the exterior wall; X 32 is the view factor from the interior wall to the chilled ceiling. Because some simplifications about the room and the wall conditions are made in the radiation network method described above, the applicability of equation (7) needs to be validated. This article validated it with experimental data. Table 2 gives the comparison between the experimental data and the calculated values at the experimental conditions that the chilled Table 3. Interior wall surface temperature of different chilled ceiling surface temperatures. Chilled ceiling surface temperature ( C) Interior surface temperature of exterior wall ( C) 6 30.3 22.5 7 30.3 23. 8 30.3 23.6 9 30.3 24. 20 30.3 24.7 2 30.3 25.2 22 30.3 25.8 23 30.3 26.3 24 30.3 26.9 25 30.3 27.4 26 30.3 27.9 Interior surface temperature of interior wall ( C) ceiling surface temperature is about 23.3 C. It can be seen that the calculated and experimental values are basically consistent. Therefore, the interior surface temperature of the interior wall under different radiant cooling surface temperatures can be calculated by equation (7). The calculation results are shown in Table 3. It shows that the interior surface temperature of the interior wall increases with the increase of the chilled ceiling surface temperature. Because the internal heat gain is relatively small during the experimental period, only from the point of view on radiant heat transfer between wall surfaces, the chilled floor condition has the same interior wall surface temperature with the chilled ceiling condition. Therefore, the interior surface temperature of interior wall of different chilled floor surface temperatures can also use the data in Table 3. Another key parameter for the calculation of mean radiant temperature is the view factor between human body and each wall surface. The view factor between body (denoted as p)

78 Journal of Building Services Engineering Research & Technology 37() c b a b a c a b c a Figure 7. Geometric meanings of a, b, c. Table 4. View factors between human body and surrounding surfaces of the room. Surrounding surfaces of the room Surface Ceiling Floor East wall West wall South exterior window North wall View factor 0.284 0.4088 0.0864 0.086 0.452 0.452 and plane (denoted as i) can be calculated by the following formula: 48 F p i ¼ X pffiffiffiffiffiffiffiffiffiffiffiffiffiffi tan Y p ffiffiffiffiffiffiffiffiffiffiffiffiffiffi 4 þ X 2 þ X 2 ð8þ Y þ pffiffiffiffiffiffiffiffiffiffiffiffiffiffi tan X p ffiffiffiffiffiffiffiffiffiffiffiffiffiffi þ Y 2 þ Y 2 where X ¼ a/.8c and Y ¼ b/.8c; meanings of a, b, and c are defined in ASHARE Handbook, 22 just as shown in Figure 7. Based on the geometrical dimensions of the room, the view factors between the human body and the surrounding surfaces of the room were calculated, as shown in Table 4. When the radiant terminal device employs chilled ceiling, its surface temperature can be set at 6 26 C within the range allowed by actual projects. But when chilled floor is employed, the floor surface temperature limit should be considered due to thermal comfort requirements. According to ASHRAE 55-200 standard, 23 the acceptable floor temperature range for people wearing lightweight indoor shoes is 9 29 C within the comfortable requirements. So the acceptable chilled floor surface temperature range is 9 29 C. Based on the data in Table 3, using equation (), the mean radiant temperatures of different radiant cooling surface temperatures under chilled ceiling and chilled floor conditions were calculated, and the results are listed in Table 5. Because the view factor between human body and chilled floor is larger than that between human body and chilled ceiling, the radiation heat transfer between the human body and the chilled floor is larger than that between the human body and the chilled ceiling under the same cooling surface temperature conditions; therefore, the mean radiant temperature of human body in chilled floor condition is lower than that in chilled ceiling condition even though these two types of cooling surfaces have the same surface temperature. Based on the calculated results in Table 5 as well as Fanger s thermal comfort model, by Programming with MATLAB, the relationship between radiant cooling surface temperature and indoor air temperature under thermal neutral conditions was derived. In the derivation analysis, both the relative humidity and the air

Sui and Zhang 79 speed were constant with a value of 50% and 0.5 m/s, respectively. The thermal resistances of clothing were taken three sets of values of 0.25 clo, 0.5 clo and 0.65 clo, which respectively represented three types of representative summer dresses such as lightweight pajamas, Table 5. Mean radiant temperature of human body at different cooling surface temperatures. Surface temperature of radiant cooling ( C) Mean radiant temperature ( C) Radiant ceiling cooling ( C) 6 22.8 7 23.3 8 23.9 9 24.4 22.9 20 24.9 23.6 2 25.4 24.2 22 25.9 24.9 23 26.5 25.5 24 27.0 26.2 25 27.5 26.8 26 28.0 27.5 Radiant floor cooling ( C) short-sleeved T-shirt + suit pants, long-sleeved shirt + dress pants. The derivation results are shown in Figure 8. In order to correspond to the summer condition of ASHRAE 55-200, the I cl ¼ 0.5 clo condition was selected for analysis. As can be seen from Figure 8(a), an indoor air temperature of 25.3 29.3 C should be taken when the chilled ceiling surface temperature is in the range of 6 26 C. Similarly, as can be seen from Figure 8(b), when the chilled floor surface temperature is in the range of 9 26 C, an indoor air temperature of 25.8 29.2 C should be taken. Figure 9 shows a comparison between the chilled ceiling condition and the chilled floor condition. As can be seen from the figure, at the same radiant cooling surface temperature, under thermal neutral conditions, the indoor air temperature required in the chilled floor condition is higher than that in the chilled ceiling condition; the temperature difference is greater with the decrease of the cooling surface temperature, and the difference is up to C when the cooling surface temperature is 9 C. This is because the lower the cooling surface temperature is, the greater the radiant heat exchange between the human body and the cooling surface in chilled floor condition is and the lower the mean radiant temperature of human body is, compared with the chilled ceiling condition. (a) (b) Icl=0.25clo Icl=0.5clo Icl=0.65clo 32 32 Icl=0.25clo Icl=0.5clo Icl=0.65clo 3 3 30 30 29 29 28 28 27 27 26 26 25 25 24 24 23 23 6 7 8 9 20 2 22 23 24 25 26 9 20 2 22 23 24 25 26 Chilled ceiling surface temperature ( C) Chilled floor surface temperature ( C) Air temperature ( C) Air temperature ( C) Figure 8. Thermal comfort charts of different clothes resistances (v ¼ 0.5 m/s, ¼ 50%). (a) Chilled ceiling; (b) chilled floor.

80 Journal of Building Services Engineering Research & Technology 37() 30 Air temperature ( C) 29 28 27 26 I cl =0.5clo Chilled ceiling Chilled floor 25 6 7 8 9 20 2 22 23 24 25 26 Cooling surface temperature ( C) Figure 9. Comparison between chilled ceiling condition and chilled floor condition. (a) 30 (b) 30 v=0. v=0.2 v=0.3 v=0. v=0.2 v=0.3 29 29 28 28 27 27 26 26 25 25 24 24 23 23 6 7 8 9 20 2 22 23 24 25 26 9 20 2 22 23 24 25 26 Chilled ceiling surface temperature ( C) Chilled floor surface temperature ( C) Air temperature ( C) Air temperature ( C) Figure 0. Thermal comfort charts of different air velocities (I cl ¼ 0.5 clo, ¼ 50%). (a) Chilled ceiling; (b) chilled floor. It leads to higher indoor air temperature required in chilled floor condition. Figure 0(a) and (b) respectively shows thermal comfort charts of chilled ceiling condition and chilled floor condition under different air velocities, which respectively presents the relationship between air temperature and chilled ceiling surface temperature or chilled floor surface temperature of different air velocities, under thermal neutral conditions. Taking into account that high air velocity will cause local thermal discomfort, its range in this paper is taken in accordance with the recommended range for traditional all air conditioning system in summer, and the air velocities of 0., 0.3, and 0.5 m/s were taken. Different combinations of parameters can achieve the same thermal comfort. Taking the data in Figure 0(a) for example, human body in the thermal environment with air temperature of 29. C, cooling surface temperature of 6 C, and air velocity of 0. m/s, has the same thermal comfort with that in the thermal environment with air temperature of 27.7 C, cooling surface temperature of 23 C, and air velocity of 0.3 m/s. In addition, it can be seen from Figure 0 that the lower the cooling surface temperature is, the smaller the effect of air velocity is.

Sui and Zhang 8 (a) 3 Air temperature ( C) 30 29 28 27 26 25 24 30% 40% 50% 60% 70% 6 7 8 9 20 2 22 23 24 25 26 Chilled ceiling surface temperaure ( C) (b) 3 Air temperature ( C) 30 29 28 27 26 25 24 30% 40% 50% 60% 70% 9 20 2 22 23 24 25 26 Chilled floor surface temperature ( C) Figure. Thermal comfort charts of different relative humidities (I cl ¼ 0.5 clo, v ¼ 0.5 m/s). (a) Chilled ceiling; (b) chilled floor. Figure (a) and (b) respectively shows thermal comfort charts of chilled ceiling condition and chilled floor condition at different relative humidities, which respectively presents the relationship between air temperature and chilled ceiling surface temperature or chilled floor surface temperature of different relative humidities, under thermal neutral conditions. The relative humidity range is not particularly limited in ASHRAE s thermal comfort standard. 23 But from the point of view of thermal comfort, extreme values of relative humidity should be avoided. The recommended level for relative humidity in ISO7730-2005 standard is between 30% and 70%, and this range is appropriate to avoid some uncomfortable problems including skin dryness or wetness, eye irritation, and respiratory inflammation. Therefore, a 30 70% relative humidity range was used in this section. From Figure (a) and (b) it can be seen that under the conditions of the same cooling surface temperature, when the relative humidity increases, the indoor air temperature should be reduced to achieve thermal neutral condition. Conclusions Based on the thermal comfort model, the relationships between low mean radiant temperature and other thermal microclimate parameters such as indoor air temperature, air velocity, and relative humidity, were analyzed, and the improved and reasonable combinations of four thermal microclimate parameters including radiant cooling surface temperature, indoor air temperature, air velocity, and relative humidity were further derived. Furthermore, a number of new thermal comfort charts taking radiant cooling surface temperature and air temperature as dominant parameters were drawn, which can provide a more intuitive reference for the design of radiant cooling systems. The main conclusions were drawn as follows:. Under the constant indoor air temperature conditions, low mean radiant temperature makes the PMV index drop. A 3 C drop of mean radiant temperature can cause a PMV index drop of 0.52. 2. In order to reach the same thermal comfort, low mean radiant temperature can be compensated by increasing air temperature. Within the recommended thermal comfort range, when the mean radiant temperature is reduced by 3 C, the indoor air temperature should be increased by 3 C; when the mean radiant temperature is reduced by 6 C, the indoor air temperature should be increased by 2.5 5 C. Under thermal neutral conditions, when the mean radiant temperature is reduced by 3 C, the indoor air temperature should be increased by 2.7 C; when the

82 Journal of Building Services Engineering Research & Technology 37() mean radiant temperature is reduced by 6 C, the indoor air temperature should be increased by 5.3 C. The increase of design air temperature recommended by the current Chinese design code GB 50736-202 is low when used for residential buildings. 3. In order to reach the same thermal comfort, low mean radiant temperature can be compensated by lower air velocity. But because the air velocity can only be reduced within a limited range, the compensating capability is limited. Low mean radiant temperature can also be compensated by higher relative humidity. This method is very useful from the energysaving view point, but it will increase the possibility of condensation of radiant devices and is not recommended to be used. 4. Radiant cooling surface temperature and indoor air temperature are the two most important parameters in radiant cooling system design. The new thermal comfort charts taking these two parameters as dominant parameters can provide a more intuitive reference for the design of radiant cooling systems. Under the design goal of thermal neutral, when the chilled ceiling surface temperature is in the range of 6 26 C, the indoor air temperature should be taken as 25.3 29.3 C; when the chilled floor surface temperature is in the range of 9 26 C, the indoor air temperature should be taken as 25.8 29.2 C. The design values of air velocity and relative humidity are not recommended to be changed, and still can follow the design standards for conventional all-air systems. 5. The results of this study apply only to residential buildings. The research methods used in the paper applies equally to public buildings, but the results may show different patterns. Due to the lack of test data about the indoor thermal environment in public buildings cooled by radiant systems at present, the relations between the radiant cooling surface temperature and the mean radiant temperature cannot be established. So this study does not extend to public buildings. It will be addressed in the follow-up studies. Declaration of conflicting interests The author(s) declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article. Funding The author(s) disclosed receipt of the following financial support for the research, authorship, and/or publication of this article: the Key Project of the National Eleventh Five Year Research Program of China (Grant No. 2006BAJ0A05), the National Natural Science Foundation of China (Grant No. 5308049), the Postdoctoral Science Foundation of China (Grant No. 203M532002), the Construction Scientific Project of Xi an City (Grant No. SJW20323), and the Natural Science Foundation of Shanxi Province (Grant No. 205JQ553). References. Wilkins CK and Kosonen R. Cool ceiling system: a European air-conditioning alternative. ASHRAE J 992; 4 45. 2. Chao C and Zhang Y. Feasibility analysis of energy efficiency reformation of an existing office build in glob by with floor panel cooling and heating system. Heat Ventilating Air Cond 200; 40: 56 60. 3. Zou ZS and Li GF. Design of air conditioning system for Tianjin Railway Station. Heat Ventilating Air Cond 200; 40: 32 35. 4. Fan L and He KJ. Study and application of floor panel heating/cooling system in lobbies of Wukesong Gymnasium. Heat Ventilating Air Cond 2005; 35: 87 90. 5. Memon RA, Chirarattananon S and Vangtook P. Thermal comfort assessment and application of radiant cooling: a case study. Build Environ 2008; 43: 85 96. 6. Imanari T, Omori T and Bogaki K. Thermal comfort and energy consumption of the radiant ceiling panel system: comparison with the conventional all-air system. Energ Build 999; 30: 67 75. 7. Tian Z and Love JA. Energy performance optimization of radiant slab cooling using building simulation and field measurements. Energ Build 2009; 4: 320 330. 8. Niu JL and Zhang LZ. Energy savings potential of chilled-ceiling combined with desiccant cooling in hot and humid climates. Energ Build 2002; 34: 487 495. 9. Wang S and Morimoto M. Evaluating the low exergy of chilled water in a radiant cooling system. Energ Build 2008; 40: 856 865. 0. Oxizidis S and Papadopoulos AM. Performance of radiant cooling surfaces with respect to energy consumption and thermal comfort. Energ Build 203; 57: 99 209.. Yang W. The investigation of thermal comfort and thermal adaptation in residential buildings during the summer