Thermal uniformity in an open plan room with an active chilled beam system and conventional air distribution systems

Accepted Manuscript Title: Thermal uniformity in an open plan room with an active chilled beam system and conventional air distribution systems Author: Kyu-Nam Rhee Mi-Su Shin Sun-Ho Choi PII: DOI: Reference:

S0378-7788(15)00095-X http://dx.doi.org/doi:10.1016/j.enbuild.2015.01.068 ENB 5674

To appear in:

ENB

Received date: Revised date: Accepted date:

19-7-2014 16-1-2015 31-1-2015

Please cite this article as: K.-N. Rhee, M.-S. Shin, S.-H. Choi, Thermal uniformity in an open plan room with an active chilled beam system and conventional air distribution systems, Energy and Buildings (2015), http://dx.doi.org/10.1016/j.enbuild.2015.01.068 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.

*Highlights (for review)

HIGHLIGHT Active chilled beam system was evaluated in terms of thermal uniformity.

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The evaluation was conducted in a full-scale test bed with 100m2 floor area.

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Air diffusion performance, draught and thermal stratification were evaluated.

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Active chilled beam system achieved an acceptable comfort with less air flow rate.

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*Manuscript

Thermal uniformity in an open plan room with an active chilled beam

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system and conventional air distribution systems

Kyu-Nam Rhee1, Mi-Su Shin2, Sun-Ho Choi3 1

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Technology Research Division, Construction and Engineering Group, Samsung C&T Corporation, Korea

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Address: Daeryung Bldg., 362 Gangnam-daero, Yeoksam-dong, Gangnam-gu, Seoul, 135081, Korea

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e-mail: [email protected] 2

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Department of Architecture, Graduate School, College of Engineering, Seoul National University, Korea

e-mail: [email protected] 3

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Address: Dept. of Architecture, 1 Gwanak-ro, Gwanak-gu, Seoul, 151-744, Korea

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Buiding Environment Research Division, Construction and Engineering Group, Samsung C&T Corporation, Korea Address: Shindeok Bldg., 343, Gangnam-daero, Seocho-dong, Seocho-gu, Seoul, 137-858, Korea

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e-mail: [email protected]

Corresponding author: Kyu-Nam Rhee ([email protected]) Address: Daeryung Bldg., 362, Gangnam-daero, Yeoksam-dong, Gangnam-gu, Seoul, 135081, Korea Tel +82-2-3669-0573 Fax +82-2-2145-6456

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Keywords air distribution system, active chilled beam system, uniformity, thermal comfort, air diffusion

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performance, draught

Abstract

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An active chilled beam (ACB) system is known to be not only energy-efficient but also

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advantageous to high indoor environment quality. For this reason, its application for indoor climate control is increasing in Europe, North America, and even Asia. The objective of this

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study is to evaluate the performance of the ACB system in terms of uniformity for the indoor thermal environment. To investigate the thermal uniformity of the ACB system, comparative

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experiments were conducted in a full-scale test bed, where conventional air distribution systems as well as the ACB system can be selectively operated. The thermal uniformity was

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evaluated with three performance indices: ADPI to evaluate the air diffusion performance, air velocity to evaluate the local discomfort due to cold draught, and vertical air temperature

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difference to evaluate temperature stratification. The result shows that the ACB system is successful in providing the acceptable thermal uniformity, even with less air flow rate than

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other conventional air distribution systems.

1. INTRODUCTION

Chilled beam systems are widely applied in Europe and North America, due to low energy consumption, convenient maintenance, silent operation, less space requirement, and suitable integration with architectural design [1-3]. Especially, an active chilled beam (hereinafter

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ACB) system is considered as an alternative air distribution system, which can provide a high level comfort and energy saving potentials, because it can reduce air flow rate from an air handling unit, by making the most use of induced room air [3, 4]. Moreover, its energy efficiency can be more improved when it is configured with a dedicated outdoor air system

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(DOAS), which can handle the latent load while supplying the correct amount of ventilation

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air [5, 6]. In order to verify the performance of the system, previous studies on the ACB system were mainly focused on the validation of fundamental performances such as cooling

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capacity [7, 8], energy saving effect [3, 9-11], humidity control in hot and humid climate [12, 13], and air distribution pattern [14-19].

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Chen et al [7] investigated the cooling capacity of chilled beams, with different water flow

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rate and circuit design for a chilled beam terminal unit. A proper method of circuitry arrangement was proposed, in order to improve thermal and hydraulic performance of active

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chilled beams. De Clercq et al [8] proposed a cooling and heating capacity model for heat exchangers in active chilled beams, which accounts for nozzle types, circuitry design, and

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heat exchanger length. The parameters of the model were calibrated with capacity measurement tests and a modified model was suggested for the implementation to capacity simulation programs.

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As for energy saving effects of the ACB system, Roth et al [3] argued that the active chilled beam systems can achieve as large as 10~20 % per year, compared with conventional VAV systems. Betz et al [9] compared the capabilities of energy simulation programs for the analysis of active chilled beam system performance. It was argued that the simulation can represent induction ratio, flexible HVAC system configurations, multiple chilled water loops, and humidity control in order to enhance the accuracy of dynamic behavior or energy consumption of active chilled beam systems. Livchak et al [10] proposed an empirical formula for coil heat transfer coefficient, which can be used to represent active chilled beams

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in energy simulation programs. In addition, it was also suggested that primary air flow should be minimized, while the use of water coil should be maximized, in order to develop more energy-efficient active chilled beam systems. Stein et al [11] argued that an active chilled beam system might be more energy-efficient than variable air volume with reheat systems, if

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it is designed with low primary air flow rate and with medium temperature chilled water. It

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was also claimed that the system is efficient for a building with high sensible load in a climate where outdoor air economizers are not effective.

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When it comes to humidity controls, Loudermilk et al [12] suggested alternatives to air handling unit configurations and design conditions to overcome condensation risk of chilled

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beams in the humid climate. It was shown that the humidity and condensation can be

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successfully controlled by applying the secondary moisture removal such as desiccant air handling unit, or relaxing space design humidity level. Kosonen et al [13] conducted a

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demonstration of condensation controls for active chilled beams in hot and humid climate. It was shown that room design conditions could be attainable with properly determined supply

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water temperature and primary air temperature, and it was verified that a dry cooling was possible if the indoor air is dehumidified prior to the operation of the water-side of active chilled beam systems.

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With regard to air distribution patterns, Fredriksson et al [14] investigated the flow from the chilled beam in a mock-up room, where air flow pattern was investigated comprehensively with a visualization method and air velocity measurement. It was found that the chilled beam generates a kind of thermal plume and air velocity fluctuations, which might cause a sensation of draught, especially for occupants near the beam. True et al [15] conducted a study on the flow pattern by the active chilled beam in 5.4 x 4.2 x 2.5 m mock-up chamber, where air flow and the consequent draught risks were investigated with different heat loads, primary air flows and induced air flows. It was suggested that the maximum room height

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should be designed within 3.5 m, in order to prevent the draught due to downward airflow, and it was also proposed that the primary and/or induced air flow should be limited, in order to mitigate excessive air velocities and draught risks. Zbořil et al [16] investigated draught rate and vertical air temperature difference in a mock-up room with 4 chilled beams, with

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different heat loads and induced air flow rates. The result showed that draught rate was lower

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than 20% and vertical temperature difference was lower than 3 °C for all cases, which complies with present comfort standards. It was also claimed that thermal flows by occupants

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or internal heat load can significantly affect the air flow pattern and the consequent draught rates. Cao et al [17] employed a PIV (Particle Image Velocimetry) method in order to

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investigate the pattern of air flow, which is attached to the ceiling surface after being

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discharged form chilled beam slots. Through the observation of the instantaneous flow structure, the study clearly provided a visualization of the attached jet, which is important for

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the efficient mixing of room air by the active chilled beam. Koskela et al [18, 19] investigated the flow pattern and thermal conditions in an open-plan test room, having a floor area of 33.6

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m2 and three active chilled beams. The experiments were conducted with different internal heat loads assuming summer, winter, and spring/autumn conditions, and thermal environment was assessed in terms of draught and air speeds. It was found that the internal heat loads

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significantly affect the air flow pattern and draught risks. It was also claimed that local maximum air speed is caused by downfall of chilled beam air jets, especially at the head level, and high air speeds are caused by large-scale circulation, especially at the floor level of the open-plan office. In terms of thermal comfort, it was shown that the ACB system could provide equivalent or more satisfactory comfort levels, compared with other conventional air distribution systems [1]. It can be regarded that the ACB system can guarantee a suitable comfort level from the viewpoint of general comfort index such as PMV (Predicted Mean Vote). However, it should

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also be noted that the thermal comfort can be attained by considering not only general comfort but also local discomfort, e.g. draught or vertical air temperature difference [20-22]. The local discomfort can be prevented by implementing an adequate air distribution, which can mitigate excessive air movement or stagnant air in the conditioned space [23].

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When applying the ACB system, the indoor room air might not be sufficiently mixed,

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because it utilizes the smaller air flow rate than conventional air distribution systems. If the room air is not sufficiently mixed throughout the space, thermal uniformity might be

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degraded and occupants’ thermal sensation can be widely varied, depending on their location in the space. On the other hand, local discomfort can be induced by low temperature and/or

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high air velocity, because supply air merges in the middle of two adjacent chilled beams and

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falls down to the floor, making occupants feel discomfort due to descending cold air [4, 18]. In addition, draught or vertical air temperature difference can be generated when induced

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indoor air is raised to the chilled beam, resulting in the local discomfort and the deterioration of indoor environmental quality. Therefore, this study aims at investigating the performance

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of the ACB system in terms of the thermal uniformity, which can be evaluated by air diffusion performance, draught rate, and vertical air temperature difference. Moreover, most of the previous studies were based on the experiments conducted in a small

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mock-up room [14-19], even though it is important to investigate how uniform thermal environment can be achieved in a full-scale space, usually larger than the mock-up test room. In addition, the comparative studies on the thermal uniformity were not sufficiently conducted when different air distribution systems, including the ACB system, conventional overhead mixing ventilation system (e.g. CAV system), underfloor air distribution system, are applied to the same space. Therefore, this study also aims at investigating the thermal uniformity in the open-plan space where multiple chilled beams are installed. To do this, a test bed facility was constructed so

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that the performance of the ACB system can be compared with that of different air distribution systems such as conventional overhead mixing ventilation systems and an underfloor air distribution (UFAD) system. Using this test bed facility, comparative experiments were conducted, in order to evaluate the thermal uniformity when applying each

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air distribution system.

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2. DEVELOPMENT OF A TESTBED

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In order to investigate the thermal uniformity in the open-plan space, a test bed was constructed so that four different air distribution systems can be selectively operated. Fig. 1

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shows the dimensions and boundary conditions of the test bed, which was developed to conduct experiments on the thermal uniformity. The experiment space of the test bed was 8.5

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m (W) x 11.8 m (L) x 2.7 m (H) and was constructed with a underfloor plenum for the UFAD system, and a ceiling plenum for air duct, chilled beams, water pipe, and so on. The envelope

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of the test bed was finished with curtain walls in North, East and South façade.

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[Fig. 1 about here]

For the comparative experiments, CAV system, CAV+EHP (electrical heat pump) system, UFAD system and ACB system were installed in the test bed, as shown in Fig. 2. The CAV and CAV+EHP system can be classified as conventional overhead mixing ventilation system. The CAV system was designed so that conditioned air could be supplied through evenly-placed ceiling diffusers, by one air handling unit for the test bed space, as described in Fig. 2(a). On the while, room air was extracted to the return slots which were installed at the interior zone and along the perimeter zone. The CAV+EHP system is a

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combined system, by which the interior zone is provided with constant air volume by a central air handling unit, while the perimeter zone is controlled individually by electrical heat pumps. This combined system is widely used for commercial buildings, because it is efficient when an individual cooling control is necessary without a central air conditioning, especially

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during holidays or overtime hours [24]. In the CAV+EHP system of this study, the location of

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supply and return diffusers was almost identical with that of CAV system, except the perimeter supply diffusers, which were connected to the electrical heat pump, as shown in Fig.

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2(b). For the UFAD system, underfloor plenum space was split into four separate zones with isolation baffles and each zone was pressurized by supply air from an air handling unit, as

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shown in Fig. 2(c). The pressure difference between the underfloor plenum and occupied

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space was adjusted so that each floor diffuser can discharge about 120 m3/h of supply air. In order to enhance the air diffusion, swirl type diffusers were deployed on the floor. The

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diffuser was designed with radial air slots and perforated air outlet in the center. In this system, supply air flow was reflected by the slots and slid along the floor, which leaded to a

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low-turbulence horizontal, radial air flow at low velocity. In order to examine the ACB system performance, 2-way discharge chilled beams were applied to the test space, as described in Fig. 2(d). Primary air was provided by a dedicated

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outdoor air system (DOAS), by which fresh outdoor air was cooled down and dehumidified. The return air was directly exhausted to the outdoor by a separate exhaust fan, which means the ACB system operated as a 100 % outdoor air system. The DOAS was equipped with a desiccant wheel, which is effective for dehumidifying outdoor air and reducing latent load. With regard to the water side, Chilled water was produced by an air source heat pump and delivered to the water side of each chilled beam, after its temperature was modulated with a plate heat exchanger.

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[Fig. 2 about here]

In order to cope with the cooling load, the capacity of CAV, CAV+EHP, and UFAD systems were determined assuming that they are designed and operated as all-air systems. Thus, the

(1)

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)

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supply air flow rate was calculated with the following steady-state heat balance equation.

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The cooling capacity of each system was decided, based on the cooling load of the test room. External design conditions are 31.2 °C and 63 % RH, and indoor design conditions are 24 °C,

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50% RH, which represent the design conditions of Seoul, Korea. By considering the heat transmittance through the envelope (curtain walls, roof, and floor), solar heat gain, internal

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heat gains, and infiltration, the total cooling load was estimated to be 7,361 W, or 73.4 W/m2, which can be regarded as cooling load of a typical office building. Based on this design load,

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the cooling capacity of each system was designed as 7,500 W. This cooling capacity resulted in approximately 2,400 m3/h of supply air flow rate for all-air systems (CAV, CAV+EHP and UFAD systems), when assuming design supply air temperature as 14 °C.

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As for ACB system, the primary air flow rate, which is delivered by the air handing unit, was determined considering the cooling capacity of a chilled beam. The chilled beam was selected so that the primary air flow rate can achieve the cooling capacity, which can process the cooling load of the test space. In this case, the primary air flow rate was 60 m3/h per chilled beam, which can deliver cooling capacity of 750 W per chilled beam (total capacity by 10 chilled beams was 7,500 W). Additionally, the determined air flow rate was also examined if it can satisfy the required ventilation rate for typical office space. The ventilation rate was determined according to ASHRAE Standard 62.1 [25], where outdoor air flow rate was

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recommended as 8.5 L/s/person. The floor area was 100 m2 and occupancy density was assumed to be 15 person/m2, therefore the ventilation rate was estimated as 450 m3/h, which can be satisfied by primary air flow rate of ten chilled beams (600 m3/h). Table 1 summarizes

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the brief specification of four air distribution systems, heat source and air handling unit.

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[Table 1. about here]

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In a room with an ACB system, supply air flow rate in the room can be estimated as sum of primary and induced air flow rate, as in Equation (2), where induced air flow rate can be

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quantified with an induction ratio. The induction ratio could be known directly, by measuring

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primary air and induced air flow rate, as defined in Equation (3).

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(2)

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(3)

[Fig. 3 about here]

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However, it is very difficult to measure the induced air flow rate because the combined (primary + induced) air is discharged from the chilled beam, while the induced air is drawn into the chilled beam simultaneously [26], as depicted in Fig. 3. For this reason, the induction ratio was estimated by applying an equation on the mixing of two airstreams of different flow rates and temperatures. As the supply air can be regarded as the mixture of primary air and induced air, supply air temperature can be calculated by applying Equation (4). This equation can be rearranged with regard to this study,

, , and

, induced air flow rate, as formulated in Equation (5). In

were measured by thermocouples at primary air plenum, discharge

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side, and induction side of chilled beam, respectively. In addition,

was measured by a flow

meter at the supply air duct, which is connected to the primary air plenum.

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(4)

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(5)

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[Fig. 4 about here]

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Fig. 4 shows the estimated induction ratio, under the condition that the primary air flow rate is controlled at 60 m3/h per each chilled beam. It is shown that the induction ratio of the

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chilled beam ranges approximately from 3 to 4, and thus at least air volume of 180 m3/h can be induced, which results in the air circulation of 2,400 m3/h (primary air plus induced room

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air by 10 chilled beams) in the entire test space. As a result, the ACB system requires

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approximately 25 % of air flow rate from an air handling unit, compared with other three air distribution systems. It should be noted that the ACB system can be applied to the space with much less air flow rate than conventional air distribution systems, which may lead to the

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consequent reduction of fan energy consumption.

3. EVALUATION METHOD

3.1 Performance evaluation index In order to investigate the thermal uniformity when applying each air distribution system, air diffusion performance index (ADPI), air velocity for draught assessment, and vertical air

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temperature difference (VATD) were adopted in this study. ADPI is useful in describing the air diffusion performance for air distribution devices [27], which is used to quantify the uniform distribution of temperature and air velocity in the occupied zone. ADPI is also one method of rating the uniformity of a given thermal environment, which is closely related with

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thermal comfort. Therefore, ADPI was designated as a performance index for thermal

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uniformity, quantifying a HVAC system performance that deliver or distribute the conditioned air at different locations. ADPI is defined as a percentage of points measured in a

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room which are within both effective draught temperature and velocity range for comfort, as

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described in following equations.

(6)

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and v < 0.35 m/s

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(7)

If the air distribution system can realize the fully-mixed condition in a room, the ADPI can be

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up to 100 %, which means that the air temperature in every point approaches average room air temperature and air velocity is maintained at an appropriate level. In practical applications, however, the air distribution system is regarded as acceptable in terms of air mixture, when

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the ADPI is higher than 80 % [28, 29].

In addition, it is also required to prevent the excessive air motion (draught) throughout the space, in order to attain the uniform thermal environment. Therefore, the draught risk needs to be minimized in the conditioned space.

According to ISO 7730 [30], the local thermal comfort is categorized into three levels (A, B, and C), and the draught risk can also be categorized by mean air velocity, as tabulated in Table 2 [18, 19]. In this study, the air velocity at ankle (0.1 m) and neck (1.1 m) levels were

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measured, and the mean air velocity were considered for the assessment of draught risks. Additionally, in order to investigate the percentage objecting to draughts, this study examined the distribution of temperature differences (between local air temperature and room average temperature), as well as air velocities, in accordance with ASHRAE Handbook [29].

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In addition to ADPI and air velocity, vertical air temperature difference (VATD) needs to be

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considered, because the thermal uniformity can be achieved by preventing excessive room air stratification in the conditioned space. Although the general comfort is satisfied, local

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discomfort at the neck or the feet can occur, if a temperature gradient is sufficiently large in the vertical direction. In this study, air temperatures at neck level (1.1 m when sitting) and

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ankle level (0.1 m) were measured, in order to evaluate the VATD. In addition, percentage

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[Table 2. about here]

(8)

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dissatisfied due to the VATD was estimated, by using following equation.

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3.2 Measurement method

In this study, physical measurement was conducted in the occupied zone of the test bed, in accordance with ASHRAE Standard [21]. Thermal uniformity was evaluated for the occupied space, between the floor and 1.8 m level above the floor, more than 1.0 m from each curtain wall, and 0.5 m from internal wall. In order to obtain thermal distribution data throughout the space, air temperature and velocity were measured at every 1 m in the horizontal plane, as described in Fig. 5(a). As for vertical axis, air temperature and velocity were measured at the height of 0.1 m, 0.6 m, 1.1 m, and 1.7

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m, 2.0 m, and 2.3 m, respectively. The measurement data at 2.0 m and 2.3 m, which is out of the occupied zone, was utilized only to figure out the vertical air temperature profile from the floor to the ceiling. In the end, total 280 points (7(x) x 10(y) x 4(z)) were included in the analysis of thermal uniformity in the occupied zone. For the measurement of temperature and

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velocity, T-type thermocouples and omni-directional hot wire anemometers were attached to a

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vertical pole at the automatic 2-axis traverse system, as described in Fig. 5(b).

The 2-axis automatic traverse system was designed to move in the direction of x (East-West)

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and y (South-North) axis, respectively, at the speed of 0.05 m/s. Thus, it took 20 seconds to move from one measurement point to the next measurement point, because the distance

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between adjacent two points is 1m. Considering the moving time of the traverse system and

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time constant of the sensors, as shown in Table 3, the data for the first 40 seconds (20 seconds for traverse moving plus 20 seconds for sensor stabilizing) was not collected by the data

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acquisition device. In order to minimize the disturbance of the measurement data at each point, the averaged measurement data for 30 seconds, after the first 40 seconds passed, was

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put to use in the analysis.

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[Fig. 5 about here]

[Table 3. about here]

4. EVALUATION RESULTS

4.1 Air temperature distribution Before evaluating the performance with regard to thermal uniformity, the distribution of air

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temperature was examined, as described in Fig. 6 and 7, which were intended to present an overall thermal condition by the operation of different air distribution systems. Fig. 6 shows the horizontal distribution of air temperature at the plane z=1.1 (1.1 m above the floor), which was selected to show temperature distributions at the breathing level. Fig. 7 shows the

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vertical distribution of air temperature at the plane y=7, which was selected because the plane

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passes through supply/return diffusers or chilled beam terminal units, thus it is suitable for showing the difference of temperature distributions with each air distribution system.

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When the CAV system was operated, it was observed that the room air was well mixed and there was not much temperature difference throughout the space (22.6~23.8 ºC, averagely

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23.1 ºC), as plotted in Fig. 6(a). Fig. 7(a) shows that there was little temperature difference in

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the vertical plane. This tendency can also be found in the previous studies, arguing that the temperature distribution in the occupied zone is quite uniform with properly designed mixing

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ventilation system [31, 32].

As for the CAV+EHP system, the interior zone is controlled by a central air handling unit,

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while the perimeter zone is controlled by separate air conditioners powered by electrical heat pumps, as mentioned earlier. For this reason, the measurement result shows that the region near the East façade resulted in relatively high temperature, and thus the uniformity was

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reduced when compared with the CAV system, as plotted in Fig. 6(b). This result is also observed in the vertical plane, where there is a temperature gradient between the interior and perimeter zone, as shown in Fig. 7(b). It seems that the operation of two separate air distribution systems (CAV for interior zone and EHP for perimeter zone) caused the slight difference of air temperature in the test space (21.6~24.5 ºC, averagely 23.0 ºC), which can lead to the decrease of thermal uniformity. When the UFAD system was operated for the test space, room air moved from floor diffusers to ceiling return diffusers at a low velocity, much similar with the pattern of displacement

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ventilation. Therefore, the points at the same horizontal plane showed almost same temperature, as described in Fig. 6(c). However, the temperature increased gradually with the height, resulting in the thermal stratification, which is usually observed in the displacement ventilation system [31, 33-35], as shown in Fig. 7(c). As a result, the temperature range

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became relatively wide (22.8~27.0 ºC, averagely 25.3 ºC), which can lead to the decrease of

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thermal uniformity.

As for the ACB system, the perimeter zone along the East façade resulted in relatively high

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temperature, as shown in Fig. 6(d), because the entire space was designed to be controlled by chilled beams, without a separate system for the perimeter zone. Fig. 7(d) also shows the

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temperature increase in the perimeter zone, from the viewpoint of vertical plane. For this

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reason, the range of room air temperature was 22.4~25.4 ºC, averagely 23.7 ºC, and the temperature difference between interior and perimeter zone was at most 1.5 ºC in the same

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horizontal plane, which might result in the decrease of thermal uniformity. Nonetheless, for more quantitative analysis of the thermal uniformity, the performance index such as ADPI, air

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velocity, and VATD needs to be investigated, in order to consider air mixing or local thermal discomfort in the entire space.

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[Fig. 6 about here]

[Fig. 7 about here]

4.2 ADPI (Air Diffusion Performance Index) In utilizing ADPI to evaluate the thermal uniformity, effective draught temperature (Ted) and air velocity were taken into account. Based on the measurement data, air velocity with

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temperature difference were plotted as in Fig. 8. The x-axis represents the difference between air temperature at each local point and an average room air temperature, while the y-axis means air velocity at each local point. ADPI can be estimated by counting the number of points within the comfort range: the region under ‘v=0.35 m/s’ line and the region between

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‘Ted=-1.5 ºC’ and ‘Ted=+1.0 ºC’ lines.

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[Fig. 8 about here]

The temperature and air velocity of the CAV system were distributed along the y-axis, which

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means the temperature throughout the space was almost uniform. In addition, air velocity

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rarely exceeded 0.35 m/s, and thus most of effective draft temperatures were located between -1.5 ºC and +1.0 ºC, resulting in ADPI of 98.6 %. As for the CAV+EHP system, the range of

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temperature difference became wider than that of the CAV system, because the perimeter and interior zone were separately operated, as discussed earlier. The ADPI was 83.2 %, which

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was lower than that of the CAV system; however, it was acceptable for the comfort criteria (80 %) recommended by previous studies [28, 29]. As for the UFAD system, there were no points where air velocity exceeded 0.35 m/s because

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it was operated similarly as low velocity displacement system. However, the temperature distribution became very wide, because the temperature near the floor was relatively low, while the temperature of upper zone was quite higher than average temperature, due to the heat absorption by gradually rising room air. As a result, the ADPI of the UFAD system was estimated to be 42.0 %. In case of the ACB system, air temperature showed more uniform distribution than CAV+EHP or UFAD system; however, there were points where air velocity was higher than 0.35 m/s. It is thought that the high velocity was caused in the region that was directly

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affected by air flow from and/or toward chilled beams (e.g. the region in the middle of adjacent two chilled beams or the region below chilled beams.) In this region, supply air from adjacent two chilled beams merges and falls down on the floor, or room air can be raised by the help of induction effect, resulting in the increased air velocity. The ADPI of the ACB

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system was estimated to be 80.7 %, which was relatively lower than those of CAV and

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CAV+EHP systems. However, it should be noted that the ACB system satisfied the recommended level (80 %), even though it is operated with much less air flow rate (600

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m3/h) than other three systems (2,400 m3/h).

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4.3 Draught

Fig. 9 shows the distribution of air velocity levels at the height of 1.1 m, which is closely

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related with the draught assessment. The CAV and CAV+EHP system resulted in slightly high air velocity levels in some regions near perimeter zone. It is thought that the higher air

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velocity was caused in the perimeter zone, because the perimeter zone was designed with higher supply air flow rate than interior zone, in order to deal with skin loads of the test bed. On the other hand, the UFAD system showed very low air velocity levels throughout the

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space, resulting in low possibilities of draught. However, a temperature gradient might increase with the height due to low air velocity, causing the increased thermal stratification. The probable thermal discomfort due to this stratification will be analyzed in the following section on vertical air temperature difference. In case of the ACB system, high air velocities were observed in the regions near chilled beams, where the air flow descends from chilled beams, or room air rises toward the chilled beam due to induction effect. Therefore, high air velocity might cause local thermal discomforts due to draught, for occupants near the chilled beams.

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[Fig. 9 about here]

Fig. 10 shows the maximum and mean air velocity at the height of 0.1 m and 1.1 m, by which

ip t

the possibilities of draught can be assessed, in accordance with ISO 7730 standard. Based on

cr

the mean air velocity, it can be known that all systems leaded to mean air velocity lower than 0.19 m/s, which means the four systems could satisfy at least Category B of ISO 7730, in

us

terms of draught. The ACB system showed similar levels of draught risks with the CAV and

an

CAV+EHP systems, while its risk was higher than that of UFAD system.

M

[Fig. 10 about here]

ed

Meanwhile, the air velocity level has a limitation that it does not consider the human body’s sensitivity to the draught. In general, the neck region is more sensitive to the low air

ce pt

temperature and high air velocity, than the ankle region [29]. In order to investigate the draught risk at the ankle and neck regions, the distribution of air temperature and velocity was overlapped with equal-percentage curves of occupants objecting to draught, in

Ac

accordance with ASHARE Handbook Fundamental [29], as shown in Fig. 11.

[Fig. 11 about here]

The thermal sensation line (diagonal dashed line) in Fig. 11 represents the temperature-air velocity conditions that make effective draught temperature become 0, which means neutral thermal sensation at the ankle and neck levels. This line can be obtained by replacing Ted (effective draught temperature) with 0 in Equation (7), as follows:

Page 20 of 49

(9)

ip t

This equation can be rearranged with regard to v, yielding following equation.

cr

(10)

us

This equation can be expressed as a thermal sensation line, a slanted dashed line in Figure 11.

an

Therefore, the right side of this line results in Ted value higher than 0, causing feeling of warmth at the ankle or neck levels. On the contrary, the left side of this line results in Ted

M

value lower than 0, causing feeling of coolness, which can lead to local discomfort due to the

ed

draught.

Fig. 11(a) shows that the four distribution systems, including the ACB system, could cause

ce pt

the feeling of coolness in the ankle region, however, the objecting percentage did not exceed 10 % in all systems. In the neck region, on the contrary, there were some points where the objecting percentage became higher than 10 % in the CAV+EHP and ACB system, as

Ac

described in Fig. 11(b). The CAV+EHP system is considered to cause high draught risk in some region, due to the high air velocity in the perimeter zone, as discussed earlier. As for the ACB system, the objecting percentage higher than 10 % is thought to be caused by descending supply air or rising induced air at the neck region. However, the objecting percentage higher than 20 % was found in very limited region, which means that most of the space could satisfy ISO 7730 category B, in terms of local discomfort due to draught. Nonetheless, it should be noted that chilled beam arrangement, primary air flow rate and/or air temperature need to be carefully selected, in order to minimize the draught risk at the neck

Page 21 of 49

region.

4.4 Vertical Air Temperature Difference

ip t

Fig. 12 shows the dimensionless vertical air temperatures, by which vertical air temperature

cr

profiles, or temperature stratification with four air distribution systems can be compared. The vertical air temperature profile was normalized with the Equation (11), considering that it

us

cannot be directly compared with each other, because the condition of supply and/or exhaust air temperature of each system would be different, in order to achieve an acceptable thermal

M

an

comfort in the conditioned room.

ed

(11)

CAV, CAV+EHP, and ACB systems showed the vertical air temperature profile that is

ce pt

commonly observed in the mixing ventilation system: there is little temperature difference between the top and bottom region of the test space. The profiles imply that the room air was well mixed in the vertical direction, although there were slight increases in air temperature

Ac

with the height. However, the UFAD system resulted in the different vertical profile, where there was more increase in the vertical air temperature than other three systems.

[Fig. 12 about here]

Fig. 13 shows the range of vertical air temperature difference and PD throughout the space. It is shown that CAV, CAV+EHP and ACB system resulted in the vertical air temperature difference less than 0.9 ºC and PD less than 0.7 %, which means ISO 7730 Category A could

Page 22 of 49

be satisfied by those three systems. On the other hand, the UFAD system resulted in the vertical air temperature difference more than 2 ºC due to temperature stratification, which belongs to ISO 7730 Category B.

cr

ip t

[Fig. 13 about here]

It can be said that the ACB system did not cause thermal discomfort due to thermal

us

stratification, because the air induction effect leaded to the good air mixture, even in vertical

M

terms of vertical air temperature difference.

an

direction. As a result, the ACB system could achieve an acceptable thermal uniformity in

ed

5. DISCUSSION

Table 4 shows the overall comparison with regard to three performance indices in this study:

ce pt

ADPI, air velocity, and VATD. Based on these results, it is thought that the thermal uniformity of the ACB system could be investigated and compared with other conventional air distribution systems. ADPI of the ACB system was estimated as 80.7 %, which means a

Ac

little lower air diffusion performance than CAV or CAV+EHP; however, it could satisfy the recommended ADPI level (higher than 80 %), in spite of much smaller supply air than other conventional air distribution systems. As for the draught risk assessment, the application of the ACB system resulted in mean air velocity of 0.17 and 0.14 m/s, at the ankle and neck levels, respectively, which showed a similar level with conventional overhead mixing ventilation systems (CAV and CAV+EHP systems). However, the ACB system performance in terms of draught prevention was a little inferior to the UFAD system, which operated like a low velocity displacement ventilation.

Page 23 of 49

VATD of the ACB system was -1.24~0.71 ºC (averagely 0.2 ºC), which implied the wellmixed condition in the vertical direction as CAV or CAV+EHP systems. Percentage dissatisfied (PD) due to the VATD also showed very low level (0.11~0.58 %), resulting in ISO 7730 category A. It is expected that the ACB system will not cause discomfort due to

cr

ip t

temperature stratification.

us

[Table 4 about here]

In summary, it was shown that the ACB system can achieve an acceptable thermal uniformity

an

with relatively small air flow rate, compared with other conventional air distribution systems.

M

However, it should be noted that the thermal uniformity could be slightly diminished in perimeter zones, where cooling load can increase due to heat gains through the envelope. To

ed

cope with this, it is required that the envelope is designed with high thermal performance, such as low heat transmittance and solar heat gain coefficients. It might be another solution to

ce pt

install separate chilled beams, which are dedicated to handle the increased heat gains in the perimeter zones. If the separate chilled beams cannot be installed, the system configuration needs to be implemented with a zone control, by which the cooling capacity for perimeter

Ac

zones can be adjusted higher than interior zones. For example, primary air flow rate to perimeter zones can be increased by control devices such as modulation dampers at air duct or variable nozzles of chilled beams. The increased air flow rate will result in the higher induction ratio, which means the augmented heat transfer in the heat exchanger, and the consequent increase in the overall cooling capacity. As for water side, chilled water temperature for perimeter zones can be lowered by a mixing valve, which makes it possible to increase the water side cooling capacity. However, it should be considered that the cooling capacity should be increased on the condition that the draught is not induced in the

Page 24 of 49

conditioned room. Thus, a further study needs to be conducted on the proper selection of design parameters, such as primary air flow rate and chilled water temperature, in order to find out a balance point that can maximize the cooling capacity, while preventing thermal discomfort. In addition, this study has a limitation that the simultaneous measurement for the

ip t

entire space was not available, because it took some time for the automatic traverse system to

cr

acquire stabilized values, when the system moves from one place to another. Therefore, it needs to take account of the slight error in the spatial distribution, which is caused by the time

us

difference of the measurement. A future study needs to be carried out, in order to establish a method to minimize the error, which can improve the evaluation accuracy of thermal

M

an

uniformity.

ed

6. CONCLUSIONS

The objective of this study was to investigate the thermal uniformity in the space, when an

ce pt

active chilled beam (ACB) system is applied to the open-plan office space. To do this, comparative experiments were conducted in the test bed, which was equipped with four different air distribution systems. In evaluating the thermal uniformity of the ACB system, air

Ac

diffusion performance index (ADPI), air velocity for draught assessment, and vertical temperature difference (VATD) were compared with conventional air distribution systems such as CAV, CAV+EHP and underfloor air distribution system. The ADPI of the ACB system was 80.7 %, satisfying comfort criteria with regard to air diffusion performance. The mean air velocity at ankle and neck level was 0.17 and 0.14 m/s, respectively, corresponding to ISO 7730 category B in terms of draught. It was found that the draught could be caused in the limited region near chilled beams; however, it could be mitigated by proper selection of design parameters such as primary air flow rate. The VATD was less than 1 ºC, resulting in

Page 25 of 49

the very low PD (percentage dissatisfied) due to thermal stratification. It can be concluded that the ACB can achieve an acceptable thermal uniformity, with less air flow rate from an air handling unit than conventional air distribution systems. The ACB system can be an alternative air distribution system that provides a uniform thermal

ip t

environment at reduced air flow rate, which can lead to the saving of fan energy and the

us

cr

consequent reduction of building energy consumption.

an

NOMENCLATURE : induction ratio [-]

M

: specific heat of air [J/(kg°C)] : vertical air temperature difference [°C] : dimensionless air temperature [-]

PD

: percentage dissatisfied due to vertical air temperature difference [%]

Q

: sensible cooling load [J]

ce pt

ed

k

: induced air flow rate [m3/s] : primary air flow rate [m3/s]

Ac

: supply air flow rate, by ACB system [m3/s] : supply air flow rate, by CAV, CAV+EHP, UFAD systems [m3/h]

r

: specific weight of air [kg/m3] : standard deviation of air velocity [m/s]

T

: air temperature at local point [°C]

Tavg

: average room air temperature [°C] : entering air temperature [°C]

Page 26 of 49

Ted

: effective draught temperature [°C] : induced air temperature [°C] : primary air temperature [°C] : return air temperature [°C]

ip t

: design room air temperature [°C]

cr

: supply air temperature [°C] : design supply air temperature [°C]

us

: air velocity at local point [m/s]

an

v

M

REFERENCES

[1] M. Virta, D. Butler, J. Gräslund, J. Hogeling, E. Kristiansen, M. Reinkainen, G. Svensson,

ed

Chilled Beam Application Guidebook, REHVA, Brussels, 2006. [2] D. Alexander, M. O’Rourke, Design considerations for active chilled beams, ASHRAE

ce pt

Journal 50 (9) (2008) 50–58

[3] K. Roth, J. Dieckmann, R. Zogg, J. Brodrick, Chilled beam cooling, ASHRAE Journal 49 (7) (2007) 84 –86

Ac

[4] K. Loudermilk, Designing chilled beams for thermal comfort, ASHRAE Journal 51 (10) (2009) 58 – 64

[5] J. Jeong, S. Mumma, W. Bahnfleth, Energy conservation benefits of a dedicated outdoor air system with parallel sensible cooling by ceiling radiant panels, ASHRAE Transactions 109 (2003) 627 –636 [6] S. Mumma, Dedicated outdoor air systems: air diffusion performance, ASHRAE IAQ Applications 5 (3) (2004) 16 –18 [7] C. Chen, W. Cai, Y. Wang, C. Lin, Performance comparison of heat exchangers with

Page 27 of 49

different circuitry arrangements for active chilled beam applications, Energy and Buildings 79 (2014) 164 – 172 [8] De Clercq B., Deltour B., Van Overloop J. Measuring and Modelling Heat Exchange Capacity of Active Chilled Beams, Proceedings of Clima 2013; 11th REHVA World Congress

ip t

[9] F. Betz, J. McNeill, B. Talbert, H. Thimmanna, N. Repka, Issues arising from the use of

cr

chilled beams in energy models, Proceedings of the 5th national conference of IBPSA-USA 2012, 655 – 667

us

[10] A. Livchak, C. Lowell, Don't turn active beams into expensive diffusers, ASHRAE Journal 54(4) (2012) 52 – 60

an

[11] J. Stein, S. Taylor, VAV reheat versus active chilled beams and DOAS, ASHRAE Journal

M

55(5) (2013) 18 – 32

[12] K. Loudermilk, D. Alexander, Efficient space humidity control with active chilled beam

ed

systems, ASHRAE Journal 54(1) (2012) 28 – 38

[13] R. Kosonen, F. Tan, A feasibility study of a ventilated beam system in the hot and humid

ce pt

climate: a case-study approach, Building and Environment 40(9) (2005) 1164–1173 [14] J. Fredriksson, M. Sandberg, B. Moshfegh, Experimental investigation of the velocity field and airflow pattern generated by cooling ceiling beams, Building and Environment 36

Ac

(7) (2001) 891 –899

[15] J. True, V. Zboril, R. Kosonen, A. Melikov, Consideration for minimising draught discomfort in rooms with active chilled beams, Proceedings of Clima 2007 Wellbeing Indoors [16] V. Zbořil, L. Bozhkov, B. Yordanova, A. Melikov, R. Kosonen, Airflow distribution in rooms with chilled beams, Proceedings of 10th International Conference on Air Distribution in Rooms (Roomvent 2007), Helsinki, Finland, June 13–15. [17] G. Cao, M. Sivukari, J. Kurnitski, M. Ruponen, O. Seppänen, Particle Image Velocimetry (PIV) application in the measurement of indoor air distribution by an active chilled beam,

Page 28 of 49

Building and Environment 45 (2012) 1932 –1940 [18] H. Koskela, H. Häggblom, R. Kosonen, M. Ruponen, Air distribution in office environment with asymmetric workstation layout using chilled beams, Building and Environment 45(9) (2010) 1923–1931

ip t

[19] H. Koskela, H. Häggblom, R. Kosonen, M. Ruponen, Flow pattern and thermal comfort

cr

in office environment with active chilled beams, HVAC&R Research18(4) (2012) 723–736 [20] B. Olesen, Guidelines for comfort, ASHRAE Journal 42(8) (2000) 40 –45

us

[21] ASHRAE. ANSI/ASHRAE Standard 55, Thermal Environmental Conditions for Human Occupancy, Atlanta: American Society of Heating, Refrigerating and Air Conditioning

an

Engineers, 2009

M

[22] F. Causone, F. Baldin, B. Olesen, S. Corgnati, Floor heating and cooling combined with displacement ventilation: Possibilities and limitations, Energy and Buildings 42(2010) 2338 –

ed

2352

[23] K. Chung, C. Lee, Predicting air flow and thermal comfort in an indoor environment

ce pt

under different air diffusion models, Building and Environment 31(1) (1996) 21 – 26 [24] T. Hwang, J. Kim, Assessment of indoor environmental quality in open-plan offices, Indoor and Built Environment 22(1) (2013) 139 –156

Ac

[25] ASHRAE. ANSI/ASHRAE Standard 62.1, Ventilation for Acceptable Indoor Air Quality, Atlanta: American Society of Heating, Refrigerating and Air Conditioning Engineers, 2013 [26] M. Ruponen, J. Tinker, A Novel Method to Measure the Air Entrainment Ratio of an Active Chilled Beam, International Journal of Ventilation, 7(4) (2007) 299 – 308 [27] W. Chow, L. Wong, Experimental studies on air diffusion of a linear diffuser and associated thermal comfort indices in an air-conditioned space, Building and Environment 29(4) (1994) 523 –530 [28] D. John, Designing air-distribution systems to maximize comfort, ASHRAE Journal,

Page 29 of 49

54(9) (2012) 20–26 [29] ASHRAE. ASHRAE Handbook of Fundamentals, Atlanta: American Society of Heating, Refrigerating and Air Conditioning Engineers, 2009 [30] ISO. ISO Standard 7730, ergonomics of the thermal environment—analytical

ip t

determination and interpretation of thermal comfort using calculation of the PMV and PPD

cr

indices and local thermal comfort criteria, Geneva, Switzerland: International Organisation for Standardisation, 2005

us

[31] G. Cao, H. Awbi, R. Yao, Y. Fan, K. Sirén, R. Kosonen, J. Zhang, A review of the performance of different ventilation and airflow distribution systems in buildings, Building

an

and Environment 73 (2014) 171 – 186

M

[32] H. Awbi, Energy efficient ventilation for retrofit buildings, Proceedings of 48th AiCARR international conference on energy performance of existing buildings 2011, 23–46

ed

[33] C. Chao, M. Wan, Airflow and air temperature distribution in the occupied region of an underfloor ventilation system, Building and Environment 39 (2004) 749 –762

ce pt

[34] T. Akimoto, T. Nobe, S. Tanabe, K. Kimura, Floor-supply displacement air-conditioning: laboratory experiments. ASHRAE Transactions 105(2) (1999) 739 –751 [35] S. Ho, L. Rosario, M. Rahman, Comparison of underfloor and overhead air distribution

Ac

systems in an office environment, Building and Environment 46 (2011) 1415 –1427

FIGURE CAPTIONS

Fig. 1 Developed test bed for the evaluation of thermal uniformity

Fig. 2 Air supply and return schemes of each system (a) CAV system (b) CAV+EHP system (c) UFAD system (d) ACB system

Page 30 of 49

Fig. 3 Active chilled beam for the test bed

Fig. 4 Induction ratio of the tested chilled beam

ip t

Fig. 5 Physical measurement system

us

Fig. 6 Temperature distribution at horizontal plane (z=1.1m)

cr

(a) Measurement grid (b) 2-axis traverse system

an

(a) CAV system (b) CAV+EHP system (c) UFAD system (d) ACB system

M

Fig. 7 Temperature distribution at vertical plane (y=7.0m)

ed

(a) CAV system (b) CAV+EHP system (c) UFAD system (d) ACB system

Fig. 8 ADPI, temperature and velocity distribution

ce pt

(a) CAV system (b) CAV+EHP system (c) UFAD system (d) ACB system

Fig. 9 Air velocity distribution at horizontal plane (z=1.1m)

Ac

(a) CAV system (b) CAV+EHP system (c) UFAD system (d) ACB system

Fig. 10 Maximum and mean air velocity at 0.1 m and 1.1 m height (a) 0.1m height (ankle region) (b) 1.1 m height (neck region)

Fig. 11 Percentage of occupants objecting to draught (a) Ankle region (b) Neck region

Page 31 of 49

Fig. 12 Vertical air temperature profiles

Fig. 13 Vertical air temperature differences

Ac

ce pt

ed

M

an

us

cr

ip t

(a) Temperature difference (b) PD

Page 32 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.01

Page 33 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.02

Page 34 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.03

Page 35 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.04

Page 36 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.05

Page 37 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.06

Page 38 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.07

Page 39 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.08

Page 40 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.09

Page 41 of 49

Ac ce p

te

d

M

an

us

cr

ip t

Fig.10

Page 42 of 49

Ac ce p

te

d

M

an

us

cr

ip t

Fig.11

Page 43 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.12

Page 44 of 49

Ac

ce

pt

ed

M

an

us

cr

i

Fig.13

Page 45 of 49

Table(s) with Caption(s)

Table 1. Operation conditions of air distribution systems CAV

CAV+EHP

UFAD

Temperature: 24°C, Relative humidity: 50%

Supply/Return

Ceiling/Ceiling

Floor/Ceiling

Ceiling/Ceiling

AHU: 600

AHU: 2,400

AHU: 600(1)

supplied by AHU

(interior zone)

(120 m3/h

or EHP [m3/h]

EHP: 1,800

per outlet)

chilled beam)

DOAS

volume AHU: 2,400

(perimeter zone) Air-cooled

Air-cooled AHU

AHU(2)

+Air

source AHU

with

desiccant dehumidification

M

EHP

Air-cooled

an

Plant type

(60 m3/h per

cr

Air

ip t

Ceiling/Ceiling

us

Set point

ACB

(1) Induced air flow rate is not included in the supply air volume. The volume means air flow rate which is supplied by an air handling unit

Ac

ce pt

ed

(2) Cooling coil is a DX(direct expansion) type, of which refrigerant is cooled by an air source heat pump

Page 46 of 49

Table 2. Comfort category of ISO 7730

PMV

Vertical air temperature

velocity [m/s]

difference

Summer

Winter

PD%

°C

ip t

Category PPD

Maximum mean air

<6

-0.2
0.12

0.10

<3

<2

B

<10 -0.5
0.19

0.16

<5

<3

C

<15 -0.7
0.24

0.21

cr

A

<4

Ac

ce pt

ed

M

an

us

<10

Page 47 of 49

Table 3. Sensor specification for the physical measurement type

accuracy

response time

hot wire omni-directional

0.05~4.99m/s

±0.1 m/s

1.0 sec

T type

-200~350°C

±0.5 °C

0.5 sec

Ac

ce pt

ed

M

an

us

cr

Thermocouple

ip t

Anemometer

range

Page 48 of 49

Table 4 Summary of evaluation results

Maximum PD [%] ISO 7730 Category

ACB 80.7 0.17 0.14 B

0.47

0.86

2.26

0.71

0.47 A

0.65 A

2.13 B

0.58 A

ip t

[°C]

UFAD 42.0 0.09 0.03 A

Ac

ce pt

ed

M

an

us

Vertical air temperature

CAV+EHP 83.2 0.18 0.16 B

cr

ADPI [%] ankle level (0.1m) Draught (Mean air neck level (1.1m) velocity [m/s]) ISO 7730 Category Maximum VATD

CAV 98.6 0.18 0.13 B

Page 49 of 49