Thermal environment in simulated offices with convective and radiant cooling systems under cooling (summer) mode of operation

Accepted Manuscript Thermal environment in simulated offices with convective and radiant cooling systems under cooling (summer) mode of operation Panu Mustakallio, Zhecho Bolashikov, Kalin Kostov, Arsen Melikov, Risto Kosonen PII:

S0360-1323(16)30038-5

DOI:

10.1016/j.buildenv.2016.02.001

Reference:

BAE 4387

To appear in:

Building and Environment

Received Date: 22 November 2015 Revised Date:

31 January 2016

Accepted Date: 1 February 2016

Please cite this article as: Mustakallio P, Bolashikov Z, Kostov K, Melikov A, Kosonen R, Thermal environment in simulated offices with convective and radiant cooling systems under cooling (summer) mode of operation, Building and Environment (2016), doi: 10.1016/j.buildenv.2016.02.001. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

ACCEPTED MANUSCRIPT Thermal Environment in Simulated Offices with Convective and Radiant Cooling Systems under Cooling (Summer) Mode of Operation Panu Mustakallio(1*, Zhecho Bolashikov(2, Kalin Kostov(2, Arsen Melikov(2, Risto Kosonen(1,3 1)

Halton Oy, Haltonintie 1-3, 47400 Kausala, Finland International Centre for Indoor Environment and Energy, DTU Civil Engineering, Technical University of Denmark, Nils Koppels Alle, Building 402, 2800 Lyngby, Denmark 3) Department of Energy Technology, Aalto University, Sähkömiehentie 4, 02150 Espoo, Finland 2)

Corresponding author, Tel: +358 40 849 7575 – Fax: +358 20 792 2088, Email: [email protected]

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*

ABSTRACT

The thermal environment in a double office room and in a six-person meeting room obtained with chilled beam (CB), chilled beam with radiant panel (CBR), chilled ceiling with ceiling installed mixing ventilation (CCMV) and four desk

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partition-mounted local radiant cooling panels with mixing ventilation (MVRC) under summer (cooling) condition was compared. MVRC system was measured only for the office room case. CB provided convective cooling while the

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remaining three systems (CBR, CCMV and MVRC) provided combined radiant and convective cooling. Solar radiation, office equipment, lighting and occupants were simulated to obtain two different heat load conditions: 38 W/m2 and 64 W/m2 in the case of office room, and 71 W/m2 and 86 W/m2 in the case of meeting room. Air temperature, globe (operative) temperature, radiant asymmetry, air velocity and turbulent intensity were measured and draught rate calculated. Manikin-based equivalent temperature (MBET) was determined by using two thermal manikins to identify

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the impact of the local thermal conditions generated by the studied systems on occupants’ thermal perception. The results revealed that the differences in the thermal conditions achieved with the four systems were not significant. CB and CBR provided slightly higher velocity level in the occupied zone. The operative temperature in the studied cases

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with chilled ceiling in operation with mixing ventilation was almost the same as the operative temperature obtained with the active chilled beam (i.e. only convective cooling). The heat load distribution played major role for the airflow

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pattern in all studied systems.

Keywords: physical measurements, chilled beam, chilled ceiling, radiant cooling, convective cooling, mixing ventilation, thermal comfort, office workplace

Abbreviation CB, Chilled beam CBR, Chilled beam with radiant panel CCMV, Chilled ceiling with ceiling installed mixing ventilation MVRC, Four desk partition-mounted local radiant cooling panels with mixing ventilation 1

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MBET, Manikin based equivalent temperature

1. Introduction Present standards ISO 7730 [1] and EN 15251 [2] recommend maximum values for the indoor thermal climate parameters (for both whole body thermal sensation and local thermal comfort, air and operative/mean radiant

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temperature, air velocity, etc.) for winter and summer conditions in order to achieve thermally comfortable indoor environment for occupants. However, in rooms with high heat loads (high cooling demand) it becomes challenging to achieve the targeted thermal environment (especially local thermal comfort) due to the increased convective flows (high volumes of air supplied in order to remove the generated heat). Therefore cooling systems based on advanced

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convective, radiant or combined heat exchange are used.

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In earlier research, indoor climate conditions in full-scale measurements of single office room, generated with radiant ceiling panels and mixing ventilation based on air supply from radial ceiling diffuser were compared to the purely convective cooling system with active chilled beam installed flush within the suspended ceiling [3, 4]. Brief comparison of thermal comfort conditions and energy consumption obtained by radiant ceiling panel system and conventional (convective) all-air system has been reported [5]. The differences between radiant and convective systems have been

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discussed. In the literature distributed by system manufacturers, radiant systems have been recommended for being able to provide favorable difference in the operative temperature leading to energy efficiency, better thermal comfort and less draught discomfort compared to convective systems.

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The main research data and design guidelines of radiant systems, chilled beam systems and mixing ventilation have been summarized in the guidebooks by Federation of European Heating, Ventilation and Air Conditioning Associations

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(REHVA) [6, 7, 8]. The research on radiant cooling and heating systems during last 50 years has been reviewed [9]. Chilled ceiling in cooling mode has been extensively studied, but not many studies on the thermal environment achieved with chilled ceiling combined with mixing ventilation and comparison of its performance with the performance of purely convective system have been reported [10, 11]. Most often the research of chilled ceiling combined with total volume ventilation has focused on the effect of supply air distribution on cooling capacity of the chilled ceiling. Recently research on room thermal environment generated by personal ventilation combined with chilled ceiling was reported [12]. The journal articles reporting research on thermal environment generated with active chilled beams is limited to only two articles [13, 14] while the performance of passive chilled beams (not in the focus of the present study) is reported in numerous publication.

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This study was performed to compare the performance of four systems based on radiant and convective cooling, namely chilled beam (CB), chilled beam with radiant panel (CBR), chilled ceiling with ceiling installed mixing ventilation (CCMV) and four desk partition-mounted local radiant cooling panels with ceiling installed mixing ventilation (MVRC) with regard to the generated thermal environment and human response. The focus of this research was on the performance of cooling mode of operation. Operating conditions in heating mode differ from these when the effect of

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buoyancy for the convection flow from both radiant and convective heating systems is reversed increasing the portion of radiation from ceiling integrated radiant systems [3, 4]. The CB and CCMV systems are often used in modern office buildings. One of the hypotheses of this research was that the performance of different combinations of radiant and convective cooling systems will have substantial effect on the flow distribution and thermal environment in rooms.

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Second hypotheses was that the use of radiant ceiling panels combined with mixing ventilation will reduce substantially the operative temperature and thus cooler environment will be achieved compare to convective cooling provided by

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active chilled beam. This paper compares the thermal environment in cooling mode of operation obtained with the systems in a double office room and in a six-person meeting room. MVRC system was measured only for the office room cases. The results of the human subject experiments performed in the case of office room will be presented in a separate paper.

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2. Methods

Measurements were performed in a climate chamber (4.12 x 4.20 x 2.89 m3, L x W x H) under steady state conditions at 26 °C design room air temperature. Two cases were simulated: office room with two occupants and meeting room with six occupants. Two cooling conditions were simulated: design (maximum) heat load - 64 W/m2 (office room) and 85

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W/m2 (meeting room), and usual heat load - 37 W/m2 (office room) and 71 W/m2 (meeting room). Design heat load was defined by including portion of direct solar load onto floor and warmer window surface temperature in office room case

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in addition to the usual heat load conditions, i.e. occupants, PC and lighting. In the meeting room case design heat load was defined only by adding direct solar load, due to limitation of maximum cooling power from chilled ceiling system. The cases were designed according to the recommendations of EN 15251 standard for category II indoor climate conditions. Heat load from occupants, computers, four ceiling lighting units and solar radiation was simulated. The heat load for the studied conditions is specified in Table 1. Surface temperature of simulated windows (water panels) and floor surface temperature (electrically heated foils below part of floor covering right below the simulated windows) was controlled to mimic the solar heat gains. Two full-sized thermal manikins with complex body shape of an average Scandinavian woman 1.7 m in height were used to resemble occupants. One of the manikins consisted of 17 body parts, the other of 23. In the office room case the places of thermal manikins were swapped for obtaining results with the 23body-parts’ thermal manikin when located at both workstations. In the meeting room case four additional heated 3

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dummies of simplified body geometry were used to simulate the remaining four occupants [15]. The positioning of the heat load in the room is shown in Figure 1. Table 1. Heat balance for studied cases Office room (floor area 17.3 m ) Design heat load Usual heat load 2 persons 2 persons 156 W 156 W 9 W/m² 9 W/m² 2 computers 2 computers 130 W 130 W 8 W/m² 8 W/m² 160 W 160 W 9 W/m² 9 W/m² 34 30 °C °C 404 W 202 W 250 W 0 W 38 W/m² 12 W/m² 1100 W 648 W 37 W/m² 64 W/m² 26 l/s 26 l/s 16 16 °C °C 312 W 312 W 18 W/m² 18 W/m² 788 W 336 W W/m² W/m² 46 19

Computers (about 65 W/computer)

Lighting Solar load - window surface temperature 2

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with 6.3 m window and 26 °C room ~ Solar load - direct solar load on the floor Total solar load Total heat loads

Meeting room (floor Design heat load 6 persons 468 W 27 W/m² 6 computers 390 W 23 W/m² 160 W 9 W/m² 30 °C 202 W 250 W 26 W/m² 1470 W 85 W/m² 54 l/s 16 °C 648 W 37 W/m² 822 W W/m² 48

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Supply air flow rate Supply air temperature Supply air cooling power in 26 °C room Cooling power demand from water

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area 17.3 m ) Usual heat load 6 persons 468 W 27 W/m² 6 computers 390 W 23 W/m² 160 W 9 W/m² 30 °C 202 W 0 W 12 W/m² 1220 W 71 W/m² 54 l/s 16 °C 648 W 37 W/m² 572 W W/m² 33

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Heat balance of test for In cooling conditions with Occupants (about 78 W/occupant)

The operating principle of the four cooling systems is schematically shown in Figure 2. In the case of CCMV cooling panels were integrated into the suspended ceiling tiles. The chilled ceiling covered 77% of the total ceiling surface. The top surface of the tiles was not insulated, thus chilled ceiling transferred heat also above suspended ceiling. Supplied air

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was distributed by two linear diffusers. Each diffuser supplied ventilation air in two directions from two slots sized 0.472 m x 0.020 m each (Figure 2). Supply air temperature in all cases was kept at 16 ± 0.5 °C and water inlet temperature at 15 ± 0.5 °C with return water 2-3 °C warmer. The same exposed chilled beam (coil length 2.10 m placed 2.5 m above the floor) was used in the cases CB and CBR (Figure 2). In the case CBR the surface area of the radiant

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panels was 3.6 m² (3.0 m (L) x 1.2 m (W)). The chilled beam was always removed from the ceiling when chilled ceiling or partition incorporated radiant panels cases were measured. Personal radiant panels were installed at the desks as

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shown in Figure 3. In this case with design heat loads, the air volume flow supplied from the linear diffusers was increased to 44 l/s to compensate for the decreased cooling power of the panel radiators.

The operating conditions of the four systems are listed in Table 2. The ventilation flow rate and the air temperature maintained in the room (measured continuously at a reference point described in the following) corresponded to Category II defined in EN15251. The water flow rate and temperature in the cooling panels and the ventilation airflow rate were determined and maintained to achieve Category II.

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Air temperature, globe (operative) temperature, mean velocity and turbulent intensity were measured at 8 heights (0.05 m, 0.1 m, 0.3 m, 0.6 m, 1.1 m, 1.7 m, 2.0 m, 2.4 m above floor), at 25 grid-locations in the room (Figure 1). At the locations of the desks (office room) and the table (meeting room) the measurements were performed at heights 1.1, 1.7, 2.0 and 2.4 m. Draught rate index [1] was calculated based on the measured parameters. Supply air flow patterns from diffusers and chilled beam were visualized with smoke for observing the overall flow field and flow directions. Surface

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temperature (walls, floor, ceiling, windows) and radiant temperature asymmetry were also measured. Radiant asymmetry was measured at 1.1 m height at location 13 (Fig. 1) in CCMV, CB and CBR cases and in the vicinity of the thermal manikins (locations 7 and 17 for office room and 9 and 17 for meeting room). In MVRC cases, radiant asymmetry was measured at the workstation at 1.1 m height between both occupants and the radiant panels, and at 0.6

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m and 1.1 m heights between both occupants and side walls. Heat loss and surface temperature of manikins’ body segments was measured and used to calculate the manikin-based equivalent temperatures, MBET [16]. Air temperature

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and globe (operative) temperature sensors were of a thermistor type with accuracy of ± 0.2 °C [17]. Air temperature sensors were shielded to compensate for radiation interference. Sensors for measuring operative temperature were placed in gray globes (spheres) and fulfilled the requirements for this type of measuring instruments recommended in the ISO Standard 7726 [17, 18]. Multi-channel wireless low-velocity thermal anemometer with eight velocity sensors was used. Velocity sensors were of an omnidirectional hot-sphere type with accuracy of ± 0.02 m/s with extra ± 1% of

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the reading for the range 0.05-0.5 m/s. The omnidirectional hot-sphere type sensors measure air speed which is referred as velocity in this paper. All measurement sensors were calibrated prior to the measurements. All measurement results were 5 minutes averaged readings.

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All measurements were initiated when steady state condition was achieved. This was documented by two loggers placed between locations 3 and 4 (location Tref in Figure 1) which measured air temperature and operative temperature at 0.6m

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and 1.1m. It was considered that the thermal environment is under steady state when the air temperature was 26 °C ± 0.2 °C. The velocity, air and globe (operative) temperature sensors were placed on single measurement pole in order to measure in all 8 heights at the same location. After recording for 5 min the air velocity, air temperature and globe (operative) temperature the stand was moved to the next location from the grid. The new measurement started after 10 min allowing the flow conditions in the test room to come to steady state after the disturbance from the person replacing the stand. When all 25 locations were measured the 2 manikins were logged for 5 min. At the end the radiant asymmetry was measured and air flow visualization with smoke was performed and recorded on video.

The thermal manikins were operated to keep the surface temperature of the manikin to be the same as the skin temperature of an average person in state of thermal comfort at light activity level (1-1.2 Met) [16]. The power supplied 5

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to each body segment to keep the skin temperature constant was measured. The power and the surface temperature were then used to calculate the manikin based equivalent temperature (MBET). MBET was used as an approximation of the equivalent temperature defined as: “the temperature of a homogenous "room", with mean radiant temperature equal to air temperature and zero air velocity, in which a person exchanges the same heat loss by convection and radiation as in

MBET = tbp - ( Pbp / hܾ‫) ݌‬

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the actual conditions”, ISO 14505 [19]. The MBET is calculated with equation 1.

(1)

where, MBET Manikin Based Equivalent Temperature (°C) surface temperature of the body segment (°C)

Pbp

power provided to each body segment (W/m2)

hbp

heat transfer coefficient for the body segment (W/m2°C)

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tbp

The heat transfer coefficient is obtained by measurements with the thermal manikin in a climate chamber under steadystate environment at different temperature (air temperature equal to the mean radiant temperature and absence of radiant temperature asymmetry) and without air movement (velocity less than 0.05 m/s). Under these conditions heat transfer

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coefficient for the whole body and the separate body segments of the manikin is obtained and used to determine the MBET for the whole body and for the body segments.

The 23 body-fragment manikin was measured at both locations 7 and 17 for office cases by swapping locations with the

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17 body-fragment manikin. In the meeting room cases, the 23 body-fragment manikin was placed at location 17 and the

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17 body-fragment manikin was at location 9 (Figure 1).

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ACCEPTED MANUSCRIPT Tref

b)

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a)

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Tref

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Fig. 1 Top view of the test room with measurement pole locations and thermal manikins: a) in office room case, and b) in meeting room) case

b)

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a)

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Height 2.89 m

c)

d)

Fig. 2 Operating principle of the four cooling systems that were studied: a) CCMV, b) CB, c) CBR and d) MVRC.

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a)

Fig. 3 Photos of the experimental setup a) in office room case with CB, CBR and CCMV, and b) in meeting room case

3. Results

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Summary of measurements results, including average, minimum, maximum and standard deviation of the measured (calculated) values at heights 0.1, 0.6, 1.1 and 1.7 meters, are presented in Table 2 and 3 for general view of the achieved

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thermal environment under each system. Manikin-based equivalent temperatures (MBET) of selected body segments of the 23-body segment thermal manikin determined in both workstations are shown in Table 4 for office room, and of both the 23-body segment and the 17-body segment thermal manikins in Table 5 for meeting room. Thermal conditions with all studied systems were very similar. Similar behavior of the air distribution pattern could be seen in smoke visualizations for all cases with supply air jets being pushed and dominated by the strong upward buoyancy flow [8]

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(generated by the warm window surface and the occupants) towards the wall opposite to the simulated windows.

Average room air temperature and globe (operative) temperature were rather similar for all studied systems. The average

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globe (operative) temperature was slightly higher than room air temperature in all cases. Only small difference, 0.2°C on average, between CCMV mostly based on radiant cooling and CB based entirely on convective cooling were measured in

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the design heat load case. There was significant horizontal globe (operative) temperature difference between window side and door side of the room (in design heat load, 1.5 °C on average and in usual heat load around 1.0 °C). The maximum horizontal temperature difference in design heat load cases was about 2.0 °C. It was similar with all cooling systems and was caused by the one-sided location of the heat loads, i.e. close to the simulated windows. Only in MVRC cases the difference was a bit lower compared to the other three systems. Due to the horizontal temperature difference, the globe (operative) temperature near the window was about 0.4-1.8 °C higher than room design temperature measured at the reference point (Figure 1) under all cases. Vertical temperature difference in the room under all studied cases was small (0.1 - 0.7 °C). This difference was a bit smaller with the radiant systems, CCMV and CBR, due to the bigger view factor towards the floor when compared to the MVRC case (the radiant panels were perpendicular to the floor surface). In the design heat load cooling case the difference can be seen most clearly. 8

ACCEPTED MANUSCRIPT The radiant asymmetry was in all cases between the windows and the opposite wall a few degrees Celsius (1.9 - 5 °C). This was due to the simulated solar heat load gains and the fact that the supply air jet cooled the wall when being deflected towards it by the strong buoyancy flow close to the windows. In CCMV case, radiant asymmetry between floor and ceiling was higher than with other systems (4.1 °C) due to the chilled ceiling surface. As specified in the method

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section the radiant asymmetry in MVRC case was measured at different locations than in the other cases. The radiant asymmetry in MVRC case differed a few degrees (1.7 - 6.5 °C) between all three measured directions, and it was highest between the upper radiant panel and the upper body of the thermal manikins at 1.1 m height (6.5 °C).

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The measured averaged room air velocities were quite similar for all four systems. With design heat loads under CCMV, CB and CBR cases relatively high velocities were measured (0.24-0.34 m/s). Highest velocities were found near the floor

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and at the wall opposite to the simulated windows. The strong buoyancy flow generated at the window pushed the supply air jets toward the opposite wall. When usual heat loads were used, the velocity levels were mostly lower in all cases except in the meeting room case where the highest velocities were measured under the usual heat load cases with CB and CBR systems. This difference with CB and CBR under meeting room case however was smoothed away when the average of the three highest velocities was compared. The highest room air velocity measured with the purely convective

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CB system was 0.05 m/s higher than the combined radiant-convective (CCMV) system. The integration of radiant panels to the purely convective system decreased the measured highest velocity in the case of office room but not in the meeting room case. The highest velocities of MVRC in the occupied zone were lower when compared with other systems. The average turbulence intensity was rather similar with all systems, between 40% and 50%. Also the average of the three

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highest turbulence intensity levels was roughly the same. Very high turbulence intensity measurements in MVRC case under the usual heat loads were caused by the very low air velocity level (average 0.06 m/s) [8]. The average draught rate

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was lower than 11% in all cases. It was 2-3% higher in purely convective CB cases. The highest draught rate in CB and CBR cases were about 5% higher than for the CCMV cases. Under the usual heat loads, draught rates got lower for all systems. This was most pronounced in the CCMV and MVRC cases. Draught rate levels in MVRC cases were lower than with the other three systems.

Manikin based equivalent temperatures were very similar for the two heat load levels. In office room cases, thermal manikins were located in similar (symmetrical) locations with regard to the window, but in meeting room cases one of the manikins was located near the window (23-body-segment) and the other near opposite wall (17-body segment). Main difference in MBET in office room cases with different cooling systems was that under design heat load, the lower MBET level was logical with slightly lower temperatures in the CCMV case. Under the usual heat load, the manikin 9

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based equivalent temperature was highest for the CCMV case. This result is not expected and needs to be studied further. The MVRC system gave a bit lower MBET for most of the body segments than other systems, except for the top of head and the back due to small view factor to the radiant panels. The MBET range of all body parts was from 23.8 – 27.3 °C in MVRC cases, a bit higher than with the other three systems, which can be also seen in the standard deviation of temperatures. Especially the MBETs of legs were low (23.8 - 25.1 °C) in MVRC case. Also in the CCMV cases, the

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standard deviation of the MBET in individual body parts was slightly bigger than in the CB and CBR cases. Main difference in MBETs in the meeting room cases was that with chilled ceiling the top of the head was nearly in design temperature (26 oC) for the person sitting near the window, while with the other three systems was 1.5-2.0 °C warmer. In the other end of the room the MBET with chilled ceiling was 3.0-3.5 °C cooler while with other three systems it was

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closer to the room design temperature of 26 oC, but still about 1-1.5 °C cooler. In most of the cases the back of the

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thermal manikins was a bit warmer than the other parts of the body due to the effect of the chair.

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Table 2. Summary of measurement results in the office room cases IN DESIGN HEAT LOAD (64 W/m2) CASE CB

CBR

MVRC

CCMV

CB

CBR

MVRC

26.1

25.8

26.1

25.9

26.0

25.8

25.9

25.8

25.3 28.0 0.6

24.9 26.7 0.5

25.3 27.6 0.6

25.0 27.2 0.4

25.5 26.8 0.3

25.0 26.4 0.3

25.2 26.7 0.3

24.2 26.7 0.4

26.3 25.3 28.2 0.7 0.1 -0.2 0.6 0.2

26.1 25.1 27.2 0.6 0.3 -0.1 0.7 0.2

26.3 25.4 27.8 0.6 0.2 0.0 0.5 0.1

26.1 25.4 27.5 0.5 0.1 0.0 0.6 0.1

26.1 25.5 27.1 0.3 0.1 -0.1 0.5 0.1

25.9 25.1 26.6 0.4 0.1 -0.1 0.4 0.1

26.1 25.2 26.9 0.4 0.0 -0.3 0.3 0.1

25.9 24.5 26.8 0.4 0.1 -0.1 0.3 0.1

26.4

26.9

26.8

25.4 1.0 27.1 25.7 1.4 0.7 0.2

25.7 1.2 27.4 25.9 1.5 0.5 0.2

25.9 1.0 27.3 25.9 1.3 0.4 0.1

26.4

26.2

26.4

26.5

25.7 0.7 26.8 25.9 0.8 0.4 0.2

25.6 0.7 26.6 25.7 0.9 0.3 0.1

25.7 0.7 26.7 25.8 0.9 0.3 0.0

25.9 0.6 26.7 26.0 0.8 0.2 0.0

0.3 0.5 4.0

0.2 0.2 4.2

0.6 0.6 3.3

0.3 0.3 2.3

0.4 0.5 3.2

0.2 0.5 2.5

0.8 0.7 2.3

0.6

1.5

-3.4

0.7

0.7

0.8

-1.7

0.8 0.13 0.27 0.29 0.05

1.7 0.12 0.24 0.25 0.05

-6.5 0.10 0.20 0.21 0.04

3.0 0.11 0.21 0.23 0.04

-0.8 0.12 0.26 0.27 0.05

0.3 0.11 0.25 0.26 0.05

-4.3 0.06 0.13 0.14 0.03

39 76 14

45 74 13

45 77 12

47 78 15

40 71 12

42 72 15

48 84 16

46 100 58

7.9 14.7 16.3 3.3

9.5 19.8 20.8 4.6

8.1 18.0 19.5 4.2

6.1 14.1 14.9 3.7

5.7 12.3 13.0 2.8

7.8 18.2 18.4 4.6

6.9 16.6 17.4 4.2

2.0 7.2 8.3 2.1

0.0 -0.1 5.0 0.3 4.1 0.13 0.23 0.24 0.04

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26.8

25.7 1.1 27.4 25.8 1.6 0.6 0.1

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CCMV

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Measurement results in occupied zone at heights 0.1, 0.6, 1.1 and 1.7 m Average air temperature [°C] Min. air temperature [°C] Max. air temperature [°C] Std. dev. of air temperature [°C] Average operative temperature [°C] Min. operative temperature [°C] Max. operative temperature [°C] Std. dev. of operative temperature [°C] Average operative - air temperature [°C] Min. operative - air temperature [°C] Max. operative - air temperature [°C] Std. dev. of operative-air temperature [°C] At height 1.1 m: Avrg. air temperature of window side [°C] Avrg. air temperature of door side [°C] Avrg. horizontal air temp. diff. [°C] Avrg. oper. temp. of window side [°C] Avrg. oper. temperature of door side [°C] Avrg. horizontal oper. temp. diff. [°C] Avrg. oper. - air temp. of window side [°C] Avrg. oper. - air temp. of door side [°C] At heights 0.1 m - 1.7 m: Avrg. vertical air temperature diff. [°C] Avrg. vertical oper. temperature diff. [°C] Max. radiant asymmetry (window-door) [°C] Max. radiant asymmetry (side-side wall, except in MVRC radiator-side wall) [°C] Max. radiant asymmetry (floor-ceiling, except in MVRC radiator-manikin) [°C] Average air velocity [m/s] Average of 3 highest velocities [m/s] Highest velocity [m/s] Std. dev. of air velocity [m/s] Average turbulence intensity [%] Average of 3 highest turb. intensities [%] Std. dev. of turbulence intensity [%] Average draught rate [%] Average of 3 highest draught rates [%] Highest draught rate [%] Std. dev. of draught rate [%]

IN USUAL HEAT LOAD (38 W/m2) CASE

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OFFICE ROOM

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Table 3. Summary of measurement results in the meeting room cases

IN DESIGN HEAT LOAD (86 W/m2) CASE CBR

CCMV

CB

CBR

25.6 24.2 27.0 0.7

25.9 24.6 27.0 0.6

25.9 25.2 26.9 0.4

25.9 24.6 27.0 0.6

25.9 24.6 26.7 0.5

25.9 24.4 27.2 0.8 0.3 0.0 0.7 0.2

26.1 24.8 27.2 0.7 0.2 0.0 0.5 0.2

26.0 25.1 26.8 0.4 0.1 -0.2 0.9 0.2

26.1 24.9 27.2 0.6 0.2 -0.1 0.4 0.1

26.0 24.8 27.0 0.6 0.2 -0.1 0.5 0.1

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CB

26.2

26.3

26.3

26.5

26.3

24.8 1.3 26.8 25.2 1.6 0.6 0.4

25.2 1.1 26.9 25.5 1.4 0.5 0.3

25.6 0.7 26.6 25.7 0.9 0.3 0.2

25.5 1.0 26.8 25.7 1.1 0.2 0.2

25.3 1.0 26.7 25.6 1.2 0.5 0.3

0.4 0.5 3.2 1.3 1.9 0.16 0.32 0.34 0.07

0.4 0.4 2.4 0.7 2.4 0.16 0.32 0.33 0.07

0.5 0.4 No meas. No meas. No meas. 0.13 0.26 0.28 0.05

0.7 0.7 2.8 1.3 2.0 0.14 0.33 0.37 0.07

0.7 0.7 1.9 -1.0 3.4 0.14 0.32 0.38 0.07

41 74 13

42 69 12

42 71 14

42 66 12

41 67 13

11.2 24.2 25.0 6.4

11.0 24.7 25.7 6.6

7.5 16.6 16.9 4.4

8.6 23.0 23.7 6.1

8.6 22.3 24.7 6.4

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Measurement results in occupied CCMV zone at heights 0.1, 0.6, 1.1 and 1.7 m 25.9 Average air temperature [°C] 24.8 Min. air temperature [°C] 27.2 Max. air temperature [°C] 0.6 Std. dev. of air temperature [°C] Average operative temperature [°C] 26.0 Min. operative temperature [°C] 24.8 Max. operative temperature [°C] 27.5 Std. dev. of operative temperature [°C] 0.7 Average operative - air temperature [°C] 0.1 Min. operative - air temperature [°C] -0.2 Max. operative - air temperature [°C] 0.5 Std. dev. of operative-air temperature [°C] 0.2 At height 1.1 m: Avrg. air temperature of window side [°C] 26.5 Avrg. air temperature of door side [°C] 25.4 Avrg. horizontal air temp. diff. [°C] 1.1 Avrg. oper. temp. of window side [°C] 26.9 Avrg. oper. temperature of door side [°C] 25.5 Avrg. horizontal oper. temp. diff. [°C] 1.4 Avrg. oper. - air temp. of window side [°C] 0.4 Avrg. oper. - air temp. of door side [°C] 0.1 At heights 0.1 m - 1.7 m: Avrg. vertical air temperature diff. [°C] 0.3 Avrg. vertical oper. temperature diff. [°C] 0.1 Max. radiant asymmetry (window-door) [°C] No meas. Max. radiant asymmetry (side-side wall) [°C] No meas. Max. radiant asymmetry (floor-ceiling) [°C] No meas. Average air velocity [m/s] 0.14 Average of 3 highest velocities [m/s] 0.28 Highest velocity [m/s] 0.29 Std. dev. of air velocity [m/s] 0.06 Average turbulence intensity [%] 40 Average of 3 highest turb. intensities [%] 62 Std. dev. of turbulence intensity [%] 12 Average draught rate [%] 8.8 Average of 3 highest draught rates [%] 18.2 Highest draught rate [%] 19.2 Std. dev. of draught rate [%] 4.7

IN USUAL HEAT LOAD (71 W/m2) CASE

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MEETING ROOM

4. Discussion

Overall there was a small difference in the generated thermal environment obtained by cooling with the chilled ceiling and chilled beam systems. This was similar to findings in earlier studies [3, 4].

The average air and globe (operative) temperature levels were very similar (less than 0.3 °C difference) with all four systems. Average operative temperature with the CCMV system was 0.1 °C higher than air temperature and near the simulated windows about 0.5 °C higher. The portion of radiation heat transfer from CCMV in cooling mode of operation was about 50% [20]. Part of cooling was delivered into the room as convective cooling of supply air. Due to the 12

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compensating heat loads in the room transferring heat by free convection and radiation, the operative (globe) temperature also with CCMV system was above air temperature. This is opposite to the referred in some radiant cooling system design guidelines and technical manuals suggesting 2-3 °C lower operative temperature, when calculating the cooling power output of the chilled ceiling. Because of this difference significant energy savings due to the constant operation under higher room air temperature in cooling mode compared to systems based on convective cooling only is claimed to

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be achieved [21-25]. Based on present results this would lead to 2-3 °C warmer room air and operative temperature in design conditions (heavy heat load) than the design criteria for indoor temperatures. The cooling power measurements of radiant panels and cooling dimensioning data according EN 14240 [26] has reduced possibility for under-sizing the cooling power, but still this is presented in the product data as significant benefit and it is advised to take into account in

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design of such radiant cooling systems. The results (Tables 2 and 3) show small benefit of integrating radiant panels into chilled beams (CBR) in order to decrease the operative temperature into the occupied zone. The energy efficiency aspects

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will be discussed in a separate paper.

The measured radiant asymmetry under all studied systems was reasonable with the used cooling water temperature and heat load distribution. Based on ISO 7730 recommendations this would result in less than 1.5 % dissatisfied due to

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radiant asymmetry with all four systems.

The major problem with the studied systems was the non-symmetrical heat load distribution in both the office room and the meeting room cases. The strong buoyancy flow generated by the warm window surface and the located nearby occupants and computers pushed the supplied ventilation cool air jets towards the wall opposite to the simulated

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windows. This could be resisted by increasing the momentum of the supply air jets that was seen in the MVRC case, where the supply air flow rate was nearly doubled. The airflow pattern was most probably affected also by the low

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momentum convection flow from the cool chilled surface in both CCMV and CBR cases and the induction airflow in CB and CBR cases, which generated negative buoyant flow near the wall opposite to the simulated windows.

Manikin based equivalent temperatures (MBET) gave additional information on the thermal conditions. It was determined for different body segments and for the whole body by calculating weighted average based on the area of the different body parts [27]. The effect of radiant cooling, mostly to the upper body especially in the CCMV cases, and very locally to the front surface of the body in the MVRC cases, could be seen clearly especially under the design heat load. Exception to that was the office room layout under the usual heat load where for some reason the MBET was higher in the CCMV case. This deviation should be researched further preferably by CFD-simulations. The most probable reason could be the existing difference in the natural convection at the windows due to less circulation of the room flow 13

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compared to the design heat load cases. Still the MBET for the top of head should be less for the CCMV case than the CB cases, which was seen clearly in the other measured cases. The effect of the convective air flow is still significant with all systems, as it covers about 50% of the heat transfer from the chilled surface also in the CCMV and the MVRC cases [20]. Systems based on radiant transfer made conditions more uniform by compensating for the effect of the heat loads, but also those systems delivered too much cooling in the locations not affected much by the heat loads. This was

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seen in the CCMV cases, where the manikin, located in the perimeter zone i.e. in the area close to the window, in office room and meeting room cases had the MBET of the head closer to the room design temperature compared to that measured under the fully convective CB system. However, when the manikin was located in the other end of the room under the meeting room cases the calculated MBET was below 26 oC. Under the MVRC case the upper part of the body

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had the MBET near the room design temperature of 26 oC, but the lower part had too low equivalent temperature, which might result in draught issues. According to this it would be preferable to use radiant panel in the perimeter zone or in a

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local radiant panel set-up for the upper body (more affected by the heat loads compared to lower body) for a more uniform thermal sensation of the occupants. However it is not recommended to use the same cooling water circuit and extend that practice to the locations not affected by the heat loads.

The postures of the thermal manikins in the office room and meeting room cases are presented in the Figure 3. The

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inclination angle of seated manikins was a bit different, inclination angle was smaller in the meeting room case. The posture of the manikin/dummy was important because it generated thermal plume that interacted with the ventilation flow. The posture of the manikins, seated and standing affected to the boundary layer around the body and the thermal plume developed above the body. The boundary layer and the thermal plume were wider in seated posture than in

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standing posture [15, 28]. The thermal plume generated by the dummy used in the present study was almost the same as the thermal plume generated by a thermal manikin with complex body shape [15]. A backward inclination of a seated

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body increased the velocity and air temperature in the breathing zone [29]. The impact of the body inclination on the thermal plume above the body was not studied in deep, however it was expected that its shape and strength did not change much when compared to the upward posture. Thus the airflow interaction changed little and had small effect on the results obtained, including the MBET.

A bit higher velocities and draught rates in CB and CBR cases were caused most probably by the bigger supply air volume by the chilled beam systems due to the induction air circulation especially in the area near the door (Fig. 1). This could have been slightly increased by the use of exposed chilled beam in this study where supply air jets attach a bit weaker to the ceiling surface and might resist convection flows a bit less than ceiling integrated chilled beam. The conditions in the room where the occupants were located were quite similar for the studied systems especially in the 14

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office room cases. The effect of delivering part of the cooling from radiant panels in the CBR cases lowered the room air velocity level only slightly compared to purely convective CB system. One reason for this was that the induction air flow rate in CB and CBR cases was the same. The highest velocities and draught rates measured under the MVRC system were lower when compared with the other three systems. This was probably caused by the increased air supply rate leading to better attachment of the supply jets along the ceiling surface due to the higher initial momentum of the jets.

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Furthermore its interaction with the buoyancy flow led to less turning of the supplied flow towards the door side and circulation towards the window side, and led to better diffusion of the supply air into the room air. Overall the increase of the heat load level in the room increased the average velocities and draught rates. This can be clearly seen in the presented measurements especially when looking at the draught rates results, where also the highest draught rates are

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lower under the lower heat load level. The draught rate gives more holistic value from local thermal conditions by taking into account in addition to the velocity the local temperature and turbulent intensity as well. Nevertheless considering that

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the relatively high air temperatures and that highest velocities occurred at the low levels (feet level) it may be expected that draught will not be a problem with the studied systems. This will be discussed in a following paper presenting results from human subject experiments.

Horizontal temperature gradient existed in the room due to the non-symmetrical heat load distribution that could not be

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avoided with any of the studied cooling systems. For this reason specific perimeter cooling system (cooling system for the area near window) or workstation installed cooling system, preferably controlled by the occupant, could provide the most optimal thermal conditions for the office room or meeting room especially within the perimeter zone. This system could be based on both radiant and convective cooling, by making the conditions more uniform with the radiant part of

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the system, and preventing too strong circulation of the convection flows with the convective part of the system by

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increasing the momentum of the supply air jet like in the MVRC case.

The four systems were further tested in a follow-up human subjective experiment with 24 participants (12 male and 12 female) in the office room case with design heat load in order to document the subjective evaluation of the generated thermal indoor environment. Detailed results will be presented in a following paper. An important result is that all four systems managed to provide a whole body thermal sensation with a median close to the “neutral” level based on the EN 15251-2007 seven-grade thermal sensation scale [2]. However for the CBR and CCMV systems the majority of subjects voted their thermal sensation closer to the median value compared to CB and MVRC, i.e. more subjects sensed and accepted the thermal environment in a similar way when CBR and CCMV were operated. This confirms the finding that occupants in office buildings perceive better and assess higher the thermal comfort provided by radiant cooling systems (chilled ceiling systems) compared to convective cooling (conventional mixing ventilation) [5]. Nevertheless, this 15

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resulted in subjective whole body thermal sensation acceptability close to “clearly acceptable” (based on the EN 15251-

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2007 acceptability scale [2]) with all four tested systems.

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Table 4. Office room layout in design and usual heat load cases with the 23-body segment thermal manikin in WS1 (left side) and WS2 (right) WS1 (left side)

Office room case

R. Chest L.-R. Chest Back of neck

Back

L.Foot

R.Foot L.-R. Foot Standard dev. of indiv.parts All

CBR 26.7 26.4 0.3 26.0 25.8 0.3 26.4 25.9 0.6 -0.4 -0.1 26.6 25.9 0.7 26.7 26.6 0.1 27.5 25.9 1.6 26.7 25.4 1.3 0.8 0.4 0.4 0.4 26.9 26.6 0.3

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L.Chest

CB 27.0 26.5 0.6 26.4 26.0 0.4 26.4 26.4 0.0 -0.1 -0.4 26.6 26.5 0.2 26.8 26.4 0.4 27.0 25.9 1.2 26.6 25.2 1.4 0.5 0.7 0.3 0.5 26.9 26.5 0.4

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Top of head

Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Difference Design-Usual Design heat load Usual heat load Design heat load Usual heat load Design heat load Usual heat load Difference Design-Usual

CCMV 26.2 27.0 -0.9 25.6 25.9 -0.3 25.9 26.9 -1.0 -0.3 -1.0 25.4 26.6 -1.3 26.0 26.8 -0.7 27.5 26.0 1.5 26.5 26.0 0.5 1.1 0.1 0.7 0.5 26.5 27.0 -0.5

MVRC 27.3 26.2 1.1 25.2 25.4 -0.2 26.1 26.0 0.1 -0.8 -0.6 25.6 26.8 -1.2 26.7 26.4 0.3 25.1 23.8 1.3 24.9 24.1 0.8 0.2 -0.2 0.9 1.2 26.2 26.1 0.1

WS2 (right side) CB 27.2 26.6 0.6 25.5 25.6 0.0 27.3 26.6 0.7 -1.7 -1.0 26.8 26.6 0.2 27.0 26.7 0.3 27.0 25.2 1.7 27.5 25.7 1.9 -0.6 -0.4 0.7 0.6 27.2 26.7 0.5

CBR 27.0 25.8 1.2 25.7 25.8 -0.1 27.1 26.6 0.5 -1.4 -0.8 26.7 26.5 0.2 27.0 26.7 0.3 27.0 25.3 1.7 27.6 25.8 1.8 -0.5 -0.5 0.6 0.5 27.0 26.6 0.4

CCMV 26.2 25.5 0.7 24.8 25.5 -0.7 26.6 27.3 -0.7 -1.8 -1.8 26.0 26.7 -0.7 26.0 26.6 -0.6 26.8 25.7 1.1 27.5 26.3 1.2 -0.7 -0.5 0.8 0.7 26.5 26.9 -0.3

MVRC 26.7 24.9 1.7 24.7 24.9 -0.2 26.7 26.5 0.1 -2.0 -1.6 26.9 26.8 0.1 26.6 26.3 0.3 24.8 24.0 0.8 24.9 24.0 0.9 -0.1 -0.1 1.0 1.2 26.3 26.1 0.2

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side) CCMV 22.7 24.7 -2.0 24.7 24.7 0.0 24.7 24.7 0.0

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25.8 25.9 -0.1 26.6 26.7 -0.1 24.9 25.7 -0.8 25.4 26.0 -0.6 -0.5 -0.3 0.7 0.7 26.0 26.6 -0.6

parts (door CBR 25.1 25.5 -0.4 25.6 25.5 0.1 25.6 25.5 0.1

25.8 25.8 0.0 26.7 26.7 0.0 25.2 25.4 -0.1 25.9 26.0 -0.1 -0.7 -0.7 0.6 0.5 26.1 26.5 -0.4

25.8 25.8 0.0 26.8 26.8 0.0 24.8 25.7 -0.9 25.1 26.2 -1.1 -0.3 -0.6 1.4 0.8 26.1 26.6 -0.6

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5. Conclusions

17 body CB 24.6 24.6 0.0 25.5 25.7 -0.2 25.5 25.7 -0.2

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Design heat load Usual heat load Difference Design-Usual Design heat load L. Chest (23 body parts) / Usual heat load Chest (17 body parts) Difference Design-Usual Design heat load R. Chest (23 body parts) / Usual heat load Chest (17 body parts) Difference Design-Usual Design heat load L.-R. Chest Usual heat load Design heat load Back of neck (23 body parts) / Usual heat load Head (17 body parts) Difference Design-Usual Design heat load Back Usual heat load Difference Design-Usual Design heat load L.Foot Usual heat load Difference Design-Usual Design heat load R.Foot Usual heat load Difference Design-Usual Design heat load L.-R. Foot Usual heat load Standard deviation Design heat load of individual body parts Usual heat load Design heat load All Usual heat load Difference Design-Usual Top of head

23 body parts (window side) CB CBR CCMV 28.5 28.2 26.5 27.7 27.9 26.3 0.8 0.3 0.3 26.0 26.0 26.0 26.0 26.0 26.0 0.0 0.0 0.0 26.5 26.5 26.3 26.5 26.4 26.3 0.0 0.1 0.0 -0.4 -0.4 -0.3 -0.4 -0.4 -0.4 27.8 26.6 26.9 27.0 26.9 27.3 0.8 -0.3 -0.4 27.7 27.6 27.5 27.8 27.7 27.4 -0.1 -0.1 0.1 28.1 27.7 29.4 27.4 27.3 27.1 0.7 0.4 2.3 28.4 28.7 28.5 26.4 26.8 26.8 2.0 1.8 1.7 -0.3 -1.0 0.8 1.0 0.4 0.3 1.0 1.0 1.2 0.7 0.7 0.6 25.9 25.8 25.4 26.1 25.8 25.5 -0.3 0.0 -0.1

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Meeting room case

The results revealed small differences in the thermal environment generated in simulated office and meeting rooms with purely convective system (chilled beam - CB) and combined convective and radiant systems including chilled

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ceiling combined with mixing ventilation (CCMV), chilled beam with integrated radiant panel (CBR) and local desk cooling radiant panel combined with mixing ventilation (MVRC) under cooling (summer) conditions. The air

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temperature and globe (operative) temperature differed up to 0.2°C. Vertical temperature in all systems was less than 1 °C and radiant temperature asymmetry less than 5 °C in all systems. The air velocities and draught rates were within the ranges recommended in the standards with slightly higher (respectively less than 0.05 m/s and 5 %) maximum levels measured in single points of the occupied zone in the CB and CBR cases compared with CCMV cases. •

The results of this study do not support the product data and design guidelines presenting up to 2-3 K lower operative temperature in rooms with radiant cooling systems. The use of such assumption may lead to elevated room air temperatures under cooling design heat load.

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The integration of the radiant panel into chilled beam did not give clear benefit to the thermal conditions.

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The heat load distribution played major role for the airflow pattern and characteristics in all studied systems. Due to the interaction of the strong upward buoyancy flow generated by the heat sources near the window the supplied ventilation air was pushed toward the wall opposite to the window. This resulted in up to 2.0 K higher globe (operative) temperature at window side and door side of the room. Under these conditions specific perimeter cooling systems or workstation installed cooling systems based on both radiant and convective cooling and preferably

•

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controlled by occupant [2] could provide the most optimal thermal conditions. The lowest velocity and draught rate levels were measured under the MVRC cases, but challenge with local radiant panel system was the lowest MBET level of certain body parts affected less by heat loads. This effect was seen also

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in meeting room cases with CCMV with thermal manikin located at the door side.

Acknowledgment

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The study is supported by Technology Agency of Finland (TEKES) in RYM-SHOK research program and the International Centre for Indoor Environment and Energy, Department of Civil Engineering, Technical University of

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Denmark.

References

[1] ISO 7730: 2005, Ergonomics of the thermal environment - Analytical determination and interpretation of thermal

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comfort using calculation of the PMV and PPD indices and local thermal comfort criteria. International Organization for Standardization, Geneva, Switzerland.

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[2] EN 15251: 2007, Indoor Environmental Input Parameters for Design and Assessment of Energy Performance of Buildings Addressing Indoor Air Quality, Thermal Environment, Lighting and Acoustics. European Committee for Standardization, B-1050 Brussels.

[3] Kosonen R, Mustakallio P, Melikov A, Duszyk M. Comparison of the Thermal Environment in Rooms with Chilled Beam and Radiant Panel Systems. Proceedings of Roomvent 2011 conference. Trondheim, Norway. [4] Duszyk M, Melikov A, Kosonen R, Mustakallio P. Comparison of Temperature and Velocity Field in Rooms with Chilled Beams and Radiant Panel Systems Combined with Mixing Ventilation. Proceedings of Roomvent 2011 conference. Trondheim, Norway.

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[5] Imanari T, Omori T, Bogaki K. Thermal comfort and energy consumption of the radiant ceiling panel system, comparison with the conventional all –air system. Energy and Buildings 1999, 30 (2), p. 167-175. [6] Babiak J, Olesen B W, Petras D. (2007) Low temperature heating and high temperature cooling, REHVA Guidebook no 7, Finland, 2007.

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[7] J Woollett, J Rimmer et al. Active and Passive Beam Application Design Guide. REHVA-ASHRAE Guidebook no 21, United States, 2015.

[8] Müller D, Kandzia C, Kosonen R, Melikov A, Nielsen P. Mixing ventilation: Guide on mixing air distribution design. REHVA Guidebook no 19, Finland, 2013. p. 10, 70.

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[9] Rhee K-N, Kim K W. A 50 year review of basic and applied research in radiant heating and cooling systems for the built environment. Building and Environment 2015, 91, p. 166-190.

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[10] Corgnati S, Perino P M, Fracastoro G V, Nielsen P V. Experimental and numerical analysis of air and radiant cooling systems in offices. Building and Environment 2009, 44, p. 801-806. [11] Chiang W-H, Wang C-Y, Huang J-S. Evaluation of cooling ceiling and mechanical ventilation systems on thermal comfort using CFD study in an office for subtropical region. Building and Environment 2012, 48, p. 113-127. [12] Lipczynska A, Kaczmarczyk J, Melikov A K. Thermal environment and air quality in office with personalized

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ventilation combined with chilled ceiling. Building and Environment 2015, 92, p. 603-614. [13] Koskela H, Häggblom H, Kosonen R, Ruponen M. Air distribution in office environment with asymmetric workstation layout using chilled beams. Building and Environment 2010, 45, p. 1923-1931. [14] Rheea K-N, Shinb M-S, Choic S-H, Thermal uniformity in an open plan room with an active chilled beam system

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and conventional air distribution systems. Energy and Buildings 2015, 93, p. 236-248. [15] Zukowska D, Melikov A, Popiolek Z. Impact of geometry of a sedentary occupant simulator on the generated

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thermal plume: Experimental investigation. HVAC&R Research 2012, 18 (4), p. 795-811. [16] Tanabe S, Zhang H, Arens E.A, Madsen T.L, Bauman F.S. Evaluating thermal environments by using a thermal manikin with controlled skin surface temperature. ASHRAE Transactions 1994, 100, 39–48. [17] Simone A, Babiak J, Bullo M, Landkilde G, Olesen B.W. Operative temperature control of radiant surface heating and cooling systems. Proceedings of Clima 2007 congress, Helsinki, Finland. [18] ISO 7726: 2008, Ergonomics of the thermal environment - Instruments for measuring physical quantities, International Organization for Standardization, Genève, Switzerland. [19] ISO 14505: 2004, Ergonomics of the thermal environment — Evaluation of thermal environment in vehicles — Part 2: Determination of Equivalent Temperature, International Organization for Standardization, Geneva, Switzerland.

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[20] Andrés-Chicote M, Tejero-González A, Velasco-Gómez E, Javier Rey F. Experimental study on the cooling capacity of a radiant cooled ceiling system. Energy and Buildings 2012, 54, p. 207-214. [21] Fraccaro s.r.l. PLAFORAD radiant ceiling product data. Italy, 2011, p. 5. [22] Schako Ceiling Cooling System Alpety product data. Germany, 2009. p. 3. [23] SPC Thermatile Plus Heating & Cooling Radiant Ceiling Panels. United Kingdom, 2014. p. 3.

[25] ECOPHIT Graphite Construction Material News. Germany, 2013. p. 7.

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[24] TROXCellenceCenter for chilled ceilings. Germany, 2010. p. 4, 6.

[26] EN 14240: 2004, Ventilation for buildings. Chilled ceilings. Testing and rating, European Committee for Standardization, B-1050 Brussels.

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[27] Thermal manikin product data and controlling software internet page. http://www. manikin.dk.PT-Teknik. Denmark. [28] Licina D, Pantelic J, Melikov A K, Sekhar C, Tham K W. Experimental investigation of the human convective

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boundary layer in a quiescent indoor environment, Building and Environment, 2014, 75, p. 79-91. [29] Licina D, Melikov A, Sekhar C, Tham K W. Air temperature investigation in microenvironment around a human

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body, Building and Environment, 2015, 92, 39-47.

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Measured thermal environment of office and meeting rooms in two simulated cooling situations with four cooling systems based on advanced convective, radiant or combined heat exchange are compared. Studied systems include chilled beam and chilled ceiling with mixing ventilation, which are used often in modern office buildings. The results revealed that the differences in the thermal conditions achieved with the four systems were not significant and chilled beam provided slightly higher velocity level in the occupied zone.

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