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Analysis on Thermal Environment of Office Room Equipped with Radiant Cooling Workstation
Available online at www.sciencedirect.com
ScienceDirect
Available online at www.sciencedirect.com Procedia Engineering 00 (2017) 000–000
ScienceDirect
www.elsevier.com/locate/procedia
Procedia Engineering 205 (2017) 658–664
10th International Symposium on Heating, Ventilation and Air Conditioning, ISHVAC2017, 1922 October 2017, Jinan, China
Analysis on Thermal Environment of Office Room Equipped with Radiant Cooling Workstation Chunhui Songa, Nianping Lia,*, Yingdong Hea, De Hea, Meiling Hea and Haowen Chena a a
College of Civil Engineering, Hunan University, Changsha, Hunan, 410082,China
Abstract In this paper, an experiment was designed to compare the thermal environment between the fan coil unit system (Case 2) and the combination of radiant cooling workstation and fan coil unit system (Case 1). Through the numerical simulation software, namely Airpak, the indoor air temperature and velocity were compared in two cases. Compared to Case 2, there are 81.82% and 7.81% decrease in vertical and horizontal temperature difference, respectively. However, the vertical air temperature difference between head and feet is 0.97℃, which is higher than Case 2. Also, the mean air velocity in Case 1 is lower than Case 2, and the air velocity in occupied zone of Case 1 ranges from 0.065 m/s to 0.12 m/s. © 2017 The Authors. Published by Elsevier Ltd. © 2017 The Authors. Published by Ltd. committee of the 10th International Symposium on Heating, Ventilation and Air Peer-review under responsibility ofElsevier the scientific Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Conditioning. Air Conditioning. Keywords: Radiant cooling workstation (RCW); Thermal environment; Computational fluid dynamics (CFD)
1. Introduction Radiant cooling systems were introduced into China about 20 years ago [1], and there are quite a number of applications [2]. Comparing the conventional air conditioning system, much less energy is consumed to cool the conditioned room because relatively low pumping energy is required. The high operation temperature of radiant cooling systems enables a chiller to operate at high efficiency, which leads to the significant reduction in primary energy consumption [3]. D.Petras reported an investigation of energy performance and indoor environment in two modern office buildings equipped by a high temperature cooling system. The results showed that the high temperature cooling system has great potential to create a comfortable indoor environment at low energy consumption [4]. For the last two decades, there has been an increasing interest in task air condition (TAC), which is * Corresponding author. Tel.: +86-731-88822667; fax: +86-731-88822667. E-mail address: [email protected] 1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning.
1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning. 10.1016/j.proeng.2017.09.837
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defined as a space conditioning system that allows thermal conditions in small, localized zones to be individually controlled by occupants [5]. The individual control of task conditioning system contributed to create the preferred environment [6]. H.Zhang designed TAC system that heats only the feet and hands, and cools only the hands and face, which provided thermal comfort in a wide range of ambient environments. The results showed that TAC system used less than 41 W for cooling and 59 W for heating [7]. N.Mao demonstrated that using the ductless bedbased TAC could lead to a better ventilation performance and energy saving performance, but a poor thermal performance in terms of a higher draft risk than using the FAC system [8]. In this paper, a new type radiant cooling workstation (RCW) was proposed. An experiment was carried out in a test room in Changsha city. Through testing and comparing the indoor air temperature field and air velocity field of the combination of RCW and FCU system (Case 1) and FCU system (Case 2), the study results indicated the proposed RCW system outperform the traditional FCU system. 2. Experiments 2.1. Experimental facilities The room located in the Floor 1st of a laboratory building at Hunan University. The distribution of experimental room is shown in Fig.1. Except the north wall, the rest ones are interior walls. There is a door (1.0m×2.1m) located in the south wall, and a window (1.5m×1.8m) in the north wall. Above the work station there is a fluorescent lamp. A supply air outlet (0.8m ×0.25m) is placed at 2.3m above the floor level.
Fig. 1.The experiment room
RCW mainly consisted of five radiant panels (Fig. 2). The size of each radiant panel is 1200 mm×600 mm×20 mm. The surface is made of aluminum alloy while the back of panels was attached with rubber-plastic insulation lay. The diameter of plastic capillary tubes inside the panels was 3 mm. RCW is supplied with cool water which is recirculated by a cooling machine.
Fig. 2.Radiant Cooling Workstation
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2.2. Measurements The measured environmental parameters included indoor air temperature, supply air temperature and flow rate of FCU and radiant panel surface temperature. Indoor air temperature was measured using TR-72U thermometers at 0.1m, 0.6m, 1.1m, 1.7m, and placed in line-1(Fig.1). Radiant panel surface temperatures were measured using Pt100 thermometers at three different points. Supply air temperature of FCU was measured using TR-72U thermometers, and the supply air flow rate of FCU was measured using SwemaAir40 hot-wire anemometer. The measurement error of the TR-72U thermometers, Pt100 thermometers and SwemaAir40 hot-wire anemometer were ±0.3℃, ±0.3℃ and ±0.04 m/s, respectively. The indoor air temperature was kept at 26℃ during the whole experimental campaign. 3. CFD simulation CFD software Airpak was used in this study. The airflow pattern and temperature distribution in the facility were governed by the conservation laws of mass, momentum and energy. The flow was assumed to be three-dimensional, turbulent, steady state and incompressible. The Boussinesq approximation hypothesis was used for the buoyant force term. The Discrete Ordinates (DO) radiant model was used for radiation heat transfer. The turbulence was modelled with the indoor zero-equation model [9]. 3.1. Boundary conditions The CFD model boundary conditions were decided according to the results from experimental measurements. Table 1 presents the boundary conditions used for the CFD simulations. Table 1. Boundary conditions. Interior walls Human Supply air temperature Supply air flow rate Radiant panel surface temperature Fluorescent lamp Exterior wall Window
Case 1 Insulation Heat/area:1.2 met 24℃ 0.5 m/s 21.5℃ Total power:34 W Heat transfer coefficient: 1.0W/(K·m2) heat transfer coefficient: 2.8 W/(K·m2)
Case 2 Insulation Heat/area:1.2 met 21.5℃ 0.5 m/s Insulation Total power:34 W heat transfer coefficient: 1.0 W/(K·m2) heat transfer coefficient: 2.8 W/(K·m2)
3.2. Discretization The hexahedral unstructured mesh which is designed by Airpak discretized the computational domain. In the boundary layer next to non-slip walls, windows, thermal manikin, outlet, inlet, radiant panel surface and surface of fluorescent lamp, there were high gradients within a small region. To capture these gradients accurately, it was necessary to have fine mash spacing normal to these objects. The final mesh had 147,455 cells. The volume range was between 1.54313×10-7 m3 to 1.116256×10-3 m3, and all elements’ volume were less than 0.15 m3. 4. Results 4.1. Comparison between the results from experimental measurements and CFD simulations The results of grid independence test showed that the mesh density was sufficient to obtain precise results. To validate the model of CFD simulations, the results of air temperature obtained from experiments were used as comparison.
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Fig. 3 shows the comparisons between the experimental and the simulation results. The differences between the experimental and CFD results were quite small which could probably be explained by the measurement errors of thermometers. As in experimental study, the temperature of the floor is lower than that of the simulation, so the temperature measured at several test points which near the floor were lower than simulation results.
Fig. 3.Comparison between measured and simulated air temperature at line-1 of two cases (Case 1 in left, Case 2 in right).
4.2. Air temperature fields Fig.4 shows the temperature field at Z=0.93m section on the yx-axis direction in two cases. In front of the thermal manikin, the air temperature of the Case 1 was lower than that of Case 2 by 0.38℃ to 1.25℃. Especially in the location below the desk near the leg, the temperature difference reached the peak of 1.25℃. The vertical temperature difference between the head and feet were 0.97℃ in Case 1 and 0.05℃ in Case 2 respectively. There was a moderate difference, less than 3℃, which satisfies the relevant requirements and regulations [10]. From 0 m to 1.3 m in height, the temperature distribution of Case 1 was more uniform than Case 2. Fig.5 indicated the vertical temperature difference of Case 2 and Case 1 in line-1 were 0.82℃ and 0.89℃, respectively. Compared to Case 1 there was 7.87% decrease in vertical temperature difference of Case 2. Along the x-direction, the horizontal temperature difference of Case 2 and Case 1 were 0.17℃ and 1.12℃, respectively. There were an 81.82% decrease in horizontal temperature difference from Case 2 to Case 1.
Fig. 4. Air temperature fields at Z=0.93m section on the yx-axis direction for two cases (Case 1 in left, Case 2 in right).
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Fig. 5. Vertical (left) and horizontal (right) distribution of air temperature for two cases
Fig. 6.Air velocity fields at Y=1.1m(up) and Y=1.7m(down) section on the zx-axis direction for two cases (Case 1 in left, Case 2 in right)
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4.3. Air temperature fields Fig.6 shows the air velocity field at Y=1.1 m and Y=1.7 m section on the zx-axis direction at two cases. In both cases, a higher air velocity region in the unoccupied zone of the test room, caused by the cooled air directly delivered to this region. The distance between the center of this region and the FCU in Case 1 was longer than Case 2, and the maximum air velocity of the center region in Case 1 was lower than that of Case 2. Because the temperature of supply air in Case 2 was 2.4℃ lower than that in Case 1, which led to the higher speed of air subside. In both cases, the air velocity in most area of the two sections were lower than 0.15m/s which can prevent a feeling of discomfort for the most sensitive [11]. The mean air velocity of Case 1 lower than Case 2, because the cold source in Case 1 was relatively uniform and the air temperature distribution was more uniform which caused a weaker force of natural convection. At the section of Y=1.1 m in occupied zone, the air velocity of Case 1 and Case 2 were about 0.065 m/s and 0.12 m/s, respectively. However, the air velocity surrounding the thermal manikin was more uniform in Case 2, also the horizontal gradient of air velocity was lower. At the section of Y=1.7 m in occupied zone ,the air velocity of Case 1 and Case 2 both were about 0.12 m/s. Due to the thermal plume generated by the thermal manikin, in the region above the occupied zone, the air velocity reached up to 0.20 m/s. 5. Discussion This research indicates that the combination of RCW and FCU system can obtain a more uniform temperature field and lower air velocity field for higher supply air temperature and more discrete cold source etc. Since RCW immediately affects the air in occupied zone and keep more favorable radiation exchange between the occupants and the radiant surfaces, so that it’s more efficient in controlling the local thermal environment. In order to simplify the numerical model, the major source of errors come from the setting of the boundary condition, which might not match the experimental condition exactly. This system generates a high temperature difference between head and feet, which could lead to discomfort. To optimize the design of this system, further researches can incorporate different factors, such as temperature radiant panel surface temperature, air supply temperature, arrangement pattern of the radiant panel etc. Simulating the energy consumption and providing feasibility analysis of the application of this system to huger workspace are also encouraged. 6. Conclusions (1) A combination of RCW and FCU system was applied in an experimental room, and of the corresponding thermal environment were experimentally and numerically evaluated. The boundary conditions necessary for CFD simulation were obtained from the experimental data and the CFD model was further validated. (2) The air temperature fields obtained from the CFD study showed that Case 1 has 7.87% and 81.82% decrease in temperature difference of vertical and horizontal, respectively. The temperature in the occupied zone of Case 1 was lower than that of Case 2. The vertical temperature difference between head and feet in Case 1 was 0.97℃, which was great than that in Case 2. (3) The air velocity obtained from the CFD study was lower than 0.15 m/s at the height of 1.1 m and 1.7 m of both cases. The mean air velocity of Case 1 was lower than Case 2. Also, the air velocity of Case 1 was 0.065 m/s and 0.12 m/s at the height of 1.1m and 1.7m, respectively. Acknowledgements This study was financially supported by the National Natural Science Foundation of China (Project No. 51578220), and China National Key R&D program “Solutions to heating and cooling of buildings in the Yangtze River region” (Grant No. 2016YFC700303, 2016YFC00306). The authors would like to thank all subjects who participated in our study.
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References [1] J.L.Niu, 1998. Chilled-ceiling air-conditioning systems-application potentials and limitations, Mainland – Hong Kong HVAC Seminar’98, Beijing. [2] R.Hu, J.L.Niu. A review of the application of radiant cooling & heating system in Mainland China, Energy and Buildings. 52(2012)11-19. [3] K.N.Rhee, K.W.Kim. A 50 year review of basic and applied research in radiant heating and cooling systems for the built environment, Building and Environment. 91(2015)166-190. [4] D.Petras, M.Krajcik, J.Bugan, E.Durisova. Indoor Environment and Energy Performance of Office Buildings Equipped with a Low Temperature Heating/High Temperature Cooling System, Advanced Materials Research. 899(2014)36-41. [5] F.S.Bauman, E.A.Arens, 1996. Task/ambient conditioning system: engineering and application guidelines final report, Centre for Built Environment, University California at Berkeley, CA, USA. [6] H.Amai, S.Tanabe, T.Akimoto, T.Genma. Thermal sensation and comfort with different task conditioning systems, Build Environ. 42(2007)3955-64. [7] H.Zhang, E.Arens, D.Kiw, E.Buchberger, F.Bauman, C.Huizenga. Comfort, perceived air quality, and work performance in a low-power taskambient conditioning system, Build Environ. 45(2010)29-39. [8] N.Mao, D.M.Pan, S.M.Deng, M.Y.Chan. Thermal, ventilation and energy saving performance evaluations of a ductless bed-based task/ambient air conditioning (TAC) system. Energy Build. 66(2013)297-305. [9] B.Zhao, X.T.Li, Q.S.Yan. Simulation of indoor air flow in ventilated room by zero-equation turbulence model, Tsinghua Sci.Technol.41 (10) (2001)109-113. [10] N.Mao, D.M.Pan, M.Y.Chan, S.M.Deng. Experimental and numerical studies on the performance evaluation of a bed-based task/ambient air conditioning system, Applied Energy. 136 (2014)956-967. [11] T.Catalina, J.Virgone, F Kuznik. Evaluation of thermal comfort using combined CFD and experimentation study in a test room equipped with a cooling ceiling, Building and Environment. 44(2009)1740-1750.
ScienceDirect
Available online at www.sciencedirect.com Procedia Engineering 00 (2017) 000–000
ScienceDirect
www.elsevier.com/locate/procedia
Procedia Engineering 205 (2017) 658–664
10th International Symposium on Heating, Ventilation and Air Conditioning, ISHVAC2017, 1922 October 2017, Jinan, China
Analysis on Thermal Environment of Office Room Equipped with Radiant Cooling Workstation Chunhui Songa, Nianping Lia,*, Yingdong Hea, De Hea, Meiling Hea and Haowen Chena a a
College of Civil Engineering, Hunan University, Changsha, Hunan, 410082,China
Abstract In this paper, an experiment was designed to compare the thermal environment between the fan coil unit system (Case 2) and the combination of radiant cooling workstation and fan coil unit system (Case 1). Through the numerical simulation software, namely Airpak, the indoor air temperature and velocity were compared in two cases. Compared to Case 2, there are 81.82% and 7.81% decrease in vertical and horizontal temperature difference, respectively. However, the vertical air temperature difference between head and feet is 0.97℃, which is higher than Case 2. Also, the mean air velocity in Case 1 is lower than Case 2, and the air velocity in occupied zone of Case 1 ranges from 0.065 m/s to 0.12 m/s. © 2017 The Authors. Published by Elsevier Ltd. © 2017 The Authors. Published by Ltd. committee of the 10th International Symposium on Heating, Ventilation and Air Peer-review under responsibility ofElsevier the scientific Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Conditioning. Air Conditioning. Keywords: Radiant cooling workstation (RCW); Thermal environment; Computational fluid dynamics (CFD)
1. Introduction Radiant cooling systems were introduced into China about 20 years ago [1], and there are quite a number of applications [2]. Comparing the conventional air conditioning system, much less energy is consumed to cool the conditioned room because relatively low pumping energy is required. The high operation temperature of radiant cooling systems enables a chiller to operate at high efficiency, which leads to the significant reduction in primary energy consumption [3]. D.Petras reported an investigation of energy performance and indoor environment in two modern office buildings equipped by a high temperature cooling system. The results showed that the high temperature cooling system has great potential to create a comfortable indoor environment at low energy consumption [4]. For the last two decades, there has been an increasing interest in task air condition (TAC), which is * Corresponding author. Tel.: +86-731-88822667; fax: +86-731-88822667. E-mail address: [email protected] 1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning.
1877-7058 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 10th International Symposium on Heating, Ventilation and Air Conditioning. 10.1016/j.proeng.2017.09.837
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Chunhui Song et al. / Procedia Engineering 205 (2017) 658–664 Chunhui Song et al./ Procedia Engineering 00 (2017) 000–000
659
defined as a space conditioning system that allows thermal conditions in small, localized zones to be individually controlled by occupants [5]. The individual control of task conditioning system contributed to create the preferred environment [6]. H.Zhang designed TAC system that heats only the feet and hands, and cools only the hands and face, which provided thermal comfort in a wide range of ambient environments. The results showed that TAC system used less than 41 W for cooling and 59 W for heating [7]. N.Mao demonstrated that using the ductless bedbased TAC could lead to a better ventilation performance and energy saving performance, but a poor thermal performance in terms of a higher draft risk than using the FAC system [8]. In this paper, a new type radiant cooling workstation (RCW) was proposed. An experiment was carried out in a test room in Changsha city. Through testing and comparing the indoor air temperature field and air velocity field of the combination of RCW and FCU system (Case 1) and FCU system (Case 2), the study results indicated the proposed RCW system outperform the traditional FCU system. 2. Experiments 2.1. Experimental facilities The room located in the Floor 1st of a laboratory building at Hunan University. The distribution of experimental room is shown in Fig.1. Except the north wall, the rest ones are interior walls. There is a door (1.0m×2.1m) located in the south wall, and a window (1.5m×1.8m) in the north wall. Above the work station there is a fluorescent lamp. A supply air outlet (0.8m ×0.25m) is placed at 2.3m above the floor level.
Fig. 1.The experiment room
RCW mainly consisted of five radiant panels (Fig. 2). The size of each radiant panel is 1200 mm×600 mm×20 mm. The surface is made of aluminum alloy while the back of panels was attached with rubber-plastic insulation lay. The diameter of plastic capillary tubes inside the panels was 3 mm. RCW is supplied with cool water which is recirculated by a cooling machine.
Fig. 2.Radiant Cooling Workstation
Chunhui Song et al. / Procedia Engineering 205 (2017) 658–664 Chunhui Song et al / Procedia Engineering 00 (2017) 000–000
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2.2. Measurements The measured environmental parameters included indoor air temperature, supply air temperature and flow rate of FCU and radiant panel surface temperature. Indoor air temperature was measured using TR-72U thermometers at 0.1m, 0.6m, 1.1m, 1.7m, and placed in line-1(Fig.1). Radiant panel surface temperatures were measured using Pt100 thermometers at three different points. Supply air temperature of FCU was measured using TR-72U thermometers, and the supply air flow rate of FCU was measured using SwemaAir40 hot-wire anemometer. The measurement error of the TR-72U thermometers, Pt100 thermometers and SwemaAir40 hot-wire anemometer were ±0.3℃, ±0.3℃ and ±0.04 m/s, respectively. The indoor air temperature was kept at 26℃ during the whole experimental campaign. 3. CFD simulation CFD software Airpak was used in this study. The airflow pattern and temperature distribution in the facility were governed by the conservation laws of mass, momentum and energy. The flow was assumed to be three-dimensional, turbulent, steady state and incompressible. The Boussinesq approximation hypothesis was used for the buoyant force term. The Discrete Ordinates (DO) radiant model was used for radiation heat transfer. The turbulence was modelled with the indoor zero-equation model [9]. 3.1. Boundary conditions The CFD model boundary conditions were decided according to the results from experimental measurements. Table 1 presents the boundary conditions used for the CFD simulations. Table 1. Boundary conditions. Interior walls Human Supply air temperature Supply air flow rate Radiant panel surface temperature Fluorescent lamp Exterior wall Window
Case 1 Insulation Heat/area:1.2 met 24℃ 0.5 m/s 21.5℃ Total power:34 W Heat transfer coefficient: 1.0W/(K·m2) heat transfer coefficient: 2.8 W/(K·m2)
Case 2 Insulation Heat/area:1.2 met 21.5℃ 0.5 m/s Insulation Total power:34 W heat transfer coefficient: 1.0 W/(K·m2) heat transfer coefficient: 2.8 W/(K·m2)
3.2. Discretization The hexahedral unstructured mesh which is designed by Airpak discretized the computational domain. In the boundary layer next to non-slip walls, windows, thermal manikin, outlet, inlet, radiant panel surface and surface of fluorescent lamp, there were high gradients within a small region. To capture these gradients accurately, it was necessary to have fine mash spacing normal to these objects. The final mesh had 147,455 cells. The volume range was between 1.54313×10-7 m3 to 1.116256×10-3 m3, and all elements’ volume were less than 0.15 m3. 4. Results 4.1. Comparison between the results from experimental measurements and CFD simulations The results of grid independence test showed that the mesh density was sufficient to obtain precise results. To validate the model of CFD simulations, the results of air temperature obtained from experiments were used as comparison.
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Fig. 3 shows the comparisons between the experimental and the simulation results. The differences between the experimental and CFD results were quite small which could probably be explained by the measurement errors of thermometers. As in experimental study, the temperature of the floor is lower than that of the simulation, so the temperature measured at several test points which near the floor were lower than simulation results.
Fig. 3.Comparison between measured and simulated air temperature at line-1 of two cases (Case 1 in left, Case 2 in right).
4.2. Air temperature fields Fig.4 shows the temperature field at Z=0.93m section on the yx-axis direction in two cases. In front of the thermal manikin, the air temperature of the Case 1 was lower than that of Case 2 by 0.38℃ to 1.25℃. Especially in the location below the desk near the leg, the temperature difference reached the peak of 1.25℃. The vertical temperature difference between the head and feet were 0.97℃ in Case 1 and 0.05℃ in Case 2 respectively. There was a moderate difference, less than 3℃, which satisfies the relevant requirements and regulations [10]. From 0 m to 1.3 m in height, the temperature distribution of Case 1 was more uniform than Case 2. Fig.5 indicated the vertical temperature difference of Case 2 and Case 1 in line-1 were 0.82℃ and 0.89℃, respectively. Compared to Case 1 there was 7.87% decrease in vertical temperature difference of Case 2. Along the x-direction, the horizontal temperature difference of Case 2 and Case 1 were 0.17℃ and 1.12℃, respectively. There were an 81.82% decrease in horizontal temperature difference from Case 2 to Case 1.
Fig. 4. Air temperature fields at Z=0.93m section on the yx-axis direction for two cases (Case 1 in left, Case 2 in right).
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Fig. 5. Vertical (left) and horizontal (right) distribution of air temperature for two cases
Fig. 6.Air velocity fields at Y=1.1m(up) and Y=1.7m(down) section on the zx-axis direction for two cases (Case 1 in left, Case 2 in right)
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4.3. Air temperature fields Fig.6 shows the air velocity field at Y=1.1 m and Y=1.7 m section on the zx-axis direction at two cases. In both cases, a higher air velocity region in the unoccupied zone of the test room, caused by the cooled air directly delivered to this region. The distance between the center of this region and the FCU in Case 1 was longer than Case 2, and the maximum air velocity of the center region in Case 1 was lower than that of Case 2. Because the temperature of supply air in Case 2 was 2.4℃ lower than that in Case 1, which led to the higher speed of air subside. In both cases, the air velocity in most area of the two sections were lower than 0.15m/s which can prevent a feeling of discomfort for the most sensitive [11]. The mean air velocity of Case 1 lower than Case 2, because the cold source in Case 1 was relatively uniform and the air temperature distribution was more uniform which caused a weaker force of natural convection. At the section of Y=1.1 m in occupied zone, the air velocity of Case 1 and Case 2 were about 0.065 m/s and 0.12 m/s, respectively. However, the air velocity surrounding the thermal manikin was more uniform in Case 2, also the horizontal gradient of air velocity was lower. At the section of Y=1.7 m in occupied zone ,the air velocity of Case 1 and Case 2 both were about 0.12 m/s. Due to the thermal plume generated by the thermal manikin, in the region above the occupied zone, the air velocity reached up to 0.20 m/s. 5. Discussion This research indicates that the combination of RCW and FCU system can obtain a more uniform temperature field and lower air velocity field for higher supply air temperature and more discrete cold source etc. Since RCW immediately affects the air in occupied zone and keep more favorable radiation exchange between the occupants and the radiant surfaces, so that it’s more efficient in controlling the local thermal environment. In order to simplify the numerical model, the major source of errors come from the setting of the boundary condition, which might not match the experimental condition exactly. This system generates a high temperature difference between head and feet, which could lead to discomfort. To optimize the design of this system, further researches can incorporate different factors, such as temperature radiant panel surface temperature, air supply temperature, arrangement pattern of the radiant panel etc. Simulating the energy consumption and providing feasibility analysis of the application of this system to huger workspace are also encouraged. 6. Conclusions (1) A combination of RCW and FCU system was applied in an experimental room, and of the corresponding thermal environment were experimentally and numerically evaluated. The boundary conditions necessary for CFD simulation were obtained from the experimental data and the CFD model was further validated. (2) The air temperature fields obtained from the CFD study showed that Case 1 has 7.87% and 81.82% decrease in temperature difference of vertical and horizontal, respectively. The temperature in the occupied zone of Case 1 was lower than that of Case 2. The vertical temperature difference between head and feet in Case 1 was 0.97℃, which was great than that in Case 2. (3) The air velocity obtained from the CFD study was lower than 0.15 m/s at the height of 1.1 m and 1.7 m of both cases. The mean air velocity of Case 1 was lower than Case 2. Also, the air velocity of Case 1 was 0.065 m/s and 0.12 m/s at the height of 1.1m and 1.7m, respectively. Acknowledgements This study was financially supported by the National Natural Science Foundation of China (Project No. 51578220), and China National Key R&D program “Solutions to heating and cooling of buildings in the Yangtze River region” (Grant No. 2016YFC700303, 2016YFC00306). The authors would like to thank all subjects who participated in our study.
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