Numerical investigation of ventilation performance of different air supply devices in an office environment

Building and Environment 90 (2015) 37e50

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Building and Environment journal homepage: www.elsevier.com/locate/buildenv

Numerical investigation of ventilation performance of different air supply devices in an office environment Huijuan Chen a, b, Setareh Janbakhsh a, b, Ulf Larsson b, Bahram Moshfegh a, b, * a b

€ping University, SE 581 83 Linko €ping, Sweden Department of Management and Engineering, Linko €vle, SE 80176 Ga €vle, Sweden Department of Building, Energy and Environmental Engineering, University of Ga

a r t i c l e i n f o

a b s t r a c t

Article history: Received 29 January 2015 Received in revised form 11 March 2015 Accepted 17 March 2015 Available online 26 March 2015

The aim of this study was to compare ventilation performance of four different air supply devices in an office environment with respect to thermal comfort, ventilation efficiency and energy-saving potential, by performing numerical simulations. The devices have the acronyms: Mixing supply device (MSD), Wall confluent jets supply device (WCJSD), Impinging jet supply device (IJSD) and Displacement supply device (DSD). Comparisons were made under identical set-up conditions, as well as at the same occupied zone temperature of about 24.2
C achieved by adding different heat loads and using different air-flow rates. Energy-saving potential was addressed based on the air-flow rate and the related fan power required for obtaining a similar occupied zone temperature for each device. Results showed that the WCJSD and IJSD could provide an acceptable thermal environment while removing excess heat more efficiently than the MSD, as it combined the positive effects of both mixing and stratification principles. This benefit also meant that this devices required less fan power than the MSD for obtaining equivalent occupant zone temperature. The DSD showed a superior performance on heat removal, air exchange efficiency and energy saving to all other devices, but it had difficulties in providing acceptable vertical temperature gradient between the ankle and neck levels for a standing person. © 2015 Elsevier Ltd. All rights reserved.

Keywords: Ventilation performance Air supply devices Thermal comfort Energy-saving potential

1. Introduction The building sector accounts for 35% of energy use and 15% of CO2 emissions in Sweden [1]. Residential, public and commercial buildings account for approximately 90% of total energy use in the building sector. The numbers are even greater in a global context where they account for more than 40% each of primary energy use and CO2 emissions. In order to reach the agreed long-term goals at national and EU level energy use in buildings needs to be sharply reduced. The challenge is to build new residential and public buildings in a more energy-efficient and sustainable manner, but an even greater challenge is to significantly reduce energy use in existing buildings and find good solutions to cover the remaining energy needs with renewable energy. In Sweden the ventilation system is an important component and account for a large part of

€ping * Corresponding author. Department of Management and Engineering, Linko € ping, Sweden. University, SE 581 83 Linko E-mail address: [email protected] (B. Moshfegh). http://dx.doi.org/10.1016/j.buildenv.2015.03.021 0360-1323/© 2015 Elsevier Ltd. All rights reserved.

the total energy usage in the residential and service sector [2]. Sweden has energy goals to reduce energy use in the building sector by 20% by 2020 and 50% by 2050 [3]. In purposed premises such as office buildings and hospitals, ventilation systems account for a considerable portion of total energy use in the building. Energy use by the ventilation system can be divided into three different parts: heating and cooling requirements and power needs for transportation of ventilation air into and out of the building. The main purpose of the ventilation system is to maintain an appropriate indoor climate and Indoor Air Quality (IAQ) for the occupants. It is also important to mention that people's perception of the indoor climate has a strong influence on their productivity and well-being [4e7]. Thus, in order to fulfill the goals for indoor climate and energy use, it is necessary to design ventilation systems and room air distribution methods that create a healthy indoor climate in an energy-efficient manner. Generally, room air distribution methods can be classified into a few classes such as: Mixing, Displacement and Hybrid. Confluent jets and Impinging jet air distribution methods are classified under Hybrid ventilation systems.

38

H. Chen et al. / Building and Environment 90 (2015) 37e50

Mixed air-flow distribution is the traditional method for supplying air to ventilated spaces. Cool air is blown in from the ceiling or wall and dilutes the room air in an attempt to provide a uniform temperature and contaminant level throughout the space. With the mixed air-flow distribution method the flow is driven by the inertia of the supply air. The ventilation efficiency is around 50%, which means a well-mixed condition into the room. Displacement air-flow distribution is a concept for supplying conditioned air to buildings. It uses the natural buoyancy of warm air to provide improved ventilation and comfort inside the room in which it was installed. A displacement system is effective in reducing cross contamination, since there is virtually no mixing in the space. In the displacement air-flow distribution system, supply air is introduced to the space at or near the floor level, at a low velocity, at a temperature slightly below the desired room temperature. The cooler supply air displaces the warmer room air, creating a fresh cold air zone at the occupied level. Heat produced in the space moves air, including contaminants, to the ceiling level where it is removed from the space by exhaust [8]. Displacement air-flow distribution systems can be designed to fulfill requirements for sustainable and energy-efficient buildings that provide healthy and productive indoor climate conditions. This system is able to realize excellent indoor climate conditions in terms of air quality and also thermal conditions. It also operates well when air-flow rate is controlled according to the demand. The ventilation efficiency is usually greater than 50% [9]. Disadvantages with displacement air-flow distribution systems can be highlighted as follows: uncomfortable air flow can be created over the floor; it cannot be used for heating; movement and obstacles in the room destroy the stratification [10]. A possible way to improve both the energy performance and ventilation efficiency in office premises is to explore new ideas on the design of supply devices that can combine the advantages from both mixing and displacement strategies. Presently, there are two on-going research projects on the design of new supply devices with appellation confluent jets and impinging jet and the authors’ intention is to investigate the performance of these devices in different indoor environments. So far, the outcomes from these ongoing studies have been presented in the following references [11e23]. Confluent jets can be described as multiple jets issuing from different nozzles with parallel axes in the same plane flow. An airflow distribution device based on wall confluent jets can be described as a united jet issued from the confluent jets supply device in the vicinity of a wall, along which it moves downwards attached to the wall due to Coanda effect [20]. Cho et al. [24] studied experimentally the characteristics of the flow field of wall confluent jets below the supply device; in addition, flow behavior of the wall confluent jets was numerically compared with the displacement system. The result showed that the wall confluent jets spread over the floor more than the displacement jet. The numerical investigation was limited to a specific configuration of wall confluent jets and displacement supply devices under one non-isothermal case. The investigation was limited to velocity and temperature. The study was done for a specific configuration of supply devices. Ventilation performance of a certain types of confluent jets and displacement supply devices was studied in a classroom by Karimipanah et al. [25]. The Confluent jets under consideration was positioned vertically at the corner in which the jets were directed against a target wall. In their study, the confluent jets supply device showed slightly better air quality and thermal comfort with displacement system. To the author's knowledge, wall confluent jets system was studied for a specific configuration of the supply device and was limited to one non-isothermal case [24,26]. Recently the flow behavior of the wall confluent jets system was

investigated experimentally in the region close to the supply device and for different indoor environments [19e21]. Janbakhsh and Moshfegh have also studied numerically in detail thermal behaviors of the wall confluent jets in a special test room and the result is going to be published in the near future. In their study, the numerical predictions for wall confluent jets of varying design configuration show that different device designs provide acceptable thermal environments and ventilation efficiency. The impinging jet air-flow distribution method works based on the principle that a high momentum air jet is discharged downwards, strikes the floor and spreads over it, thus distributing the fresh air along the floor in the form of a very thin shear layer. The supplied air momentum from an impinging jet device is in between that from mixing and displacement devices, and therefore they are also called medium momentum supply devices [27]. In the previous study of Karimipanah and Awbi [28], the impinging jet ventilation was compared with displacement ventilation in terms of ventilation effectiveness, velocity and temperature distributions in the room. However, draft discomfort has not been compared. The impinging jet air-flow distribution method retains the strength of displacement strategy while overcoming some drawbacks, thus it could provide efficient ventilation for occupied space by overcoming buoyancy forces from heat sources [28], and also be used for space heating. Detailed studies regarding flow and thermal behaviors of the impinging jet principle have been carried out by e.g. Chen et al. [11,12,29], and performance comparisons with other air-flow distribution methods can be found in Refs. [26,27,30]. These studies showed that impinging jet air-flow distribution method has a promising application for different indoor environments. Karimipanah and Moshfegh [31] numerically studied ventilation performance of four supply devices in a mock-up of an office room. In order to achieve the same value of air distribution index, the results revealed that confluent jets requires the lowest fan power compared to the mixing, displacement, impinging jet supply devices, which this study was limited to one specific case. The numerical investigation was limited to the certain type of confluent jets supply device. The aim of this study is to compare ventilation performance of different air supply devices in a small office environment with respect to thermal comfort, ventilation efficiency and energysaving potential, by performing numerical simulations. The current study is undertaken to compare the ventilation performance of the proposed wall confluent jets and impinging jet supply devices with that of the mixing and displacement supply devices. Four airflow distribution devices which involve mixing and/or stratification principles were considered: Mixing supply device (MSD), Wall confluent jets supply device (WCJSD), Impinging jet supply device (IJSD) and Displacement supply device (DSD). 2. Numerical method In this study numerical simulation was used to predict the velocity and temperature in an office environment for four types of air supply devices. Different turbulence models, i.e., the renormalization group (RNG) kε model, shear stress transport (SST) ku model, the v2  f model and the Reynolds stress model (RSM) were employed for investigation of the studied air supply devices. The chosen turbulence models were verified with the experimental results in the previous studies. The RNG kε turbulence model was used successfully for prediction of the MSD in an office room [26,32]. The SST ku model showed a good agreement with the experimental data in the near field of confluent jets [15] and in a room with a wall confluent jets ventilation system [20]. For the IJSD, the v2  f model has been validated for both jet region and various

H. Chen et al. / Building and Environment 90 (2015) 37e50

39

zones inside an office environment, see Ref. [13]. For the DSD, the Reynolds Stress Model (RSM) showed the best agreement with the measured temperature and velocity profiles near the DSD, see Ref. [33].

varies with temperature. The radiation heat is taken into account by using a discrete ordinates (DO) model; for more details see ANSYS [34]. Based on these assumptions, the continuity, Reynoldsaveraged NaviereStokes (RANS) and energy equations are given by:

2.1. Physical models

vUi ¼0 vxi

(1)


   v rUj Ui vP v 0 0  rui uj ¼ þ mV2 Ui þ ðr  r0 Þgi þ vxi vxj vxj

(2)


   v rCp Uj T v 0  rCp ui q ¼ lV2 T þ vxj vxj

(3)

The room under consideration has the dimensions 4.2 m length, 3.6 m width, and 2.5 m height. The mannequin, computer (PC) and lighting were simulated as internal heat loads. The layout of the physical models can be seen in Fig. 1. The air enters the room through the air supply devices, and room air is extracted at an exhaust placed at the upper corner of the same wall. The MSD is a slot with a length of 0.4 m and a width of 0.02 m, and installed close to the ceiling, i.e., 2.36 m above the floor. In the WCJSD the air issues from the 192 round jets that are placed in an array of 24  8. Each nozzle has a diameter of d ¼ 10 mm and spacing center to center between each nozzle in the same row and between two parallel rows is 14 mm. The studied WCJSD is installed at height 1.6 m above the floor. The IJSD is semi-elliptical with an opening area of 0.0166 m2 and installed at a height 0.8 m above the floor. The DSD is a flat diffuser with the dimensions length by height 0.49 m  0.4 m, located at 0.09 m above the floor. The schematic of the supply devices, except for the MSD the geometry of which is simple can be seen in Fig. 1.

where ui uj is Reynolds stresses, ui q is turbulent heat fluxes, and these two terms must be modeled in order to close the system of 0 0 0 equations. By using the Boussinesq hypothesis, ui uj and ui q are defined by:

2.2. Governing equations

where k is turbulent kinetic energy, mt eddy viscosity and Sij is the strain-rate tensor expressed as:

A steady state three-dimensional model is considered for analyzing the flow and temperature in the whole room. The buoyancy effect is included in the momentum equation, and the density is treated by the incompressible ideal gas law that it only

vUi vUj Sij ¼ 0:5 þ vxj vxi

0

0

0

0

rui uj ¼ 2mt Sij þ

0

2 d rk 3 ij

m vT 0 rCp ui q ¼  t st vxj

(4)

(5)

!

Fig. 1. Sketch of the office room with installed supply devices (MSD, WCJSD, IJSD and DSD) at the middle of the inlet wall.

(6)

40

H. Chen et al. / Building and Environment 90 (2015) 37e50

2.3. Turbulence modeling 2.3.1. The RNG kε model The RNG kε model was presented by Yakhot and Orszag [35]. The model uses the mathematical technique called renormalization group method [34]. An additional term in ε equation improves the accuracy of prediction of k in the RNG kε model. Details about the RNG kε model can be found in Moshfegh and Nyiredy [36]. The turbulence kinetic energy k and its dissipation ε are given by Ref. [34]:


 v rUj k v ¼ vxj vxj




 v rUj ε v ¼ vxj vxj



mþ

 mt vk sk vxj 

mþ

mt vε sε vxj

!

!

þ Pk þ PB  rε

(7)

2 ε * ε r þ Cε1 ½Pk þ Cε3 PB   Cε2 k k

(8) 2

2 Cm r kε ,

Where eddy viscosity mt ¼ turbulence production Pk¼mtS , qffiffiffiffiffiffiffiffiffiffiffiffiffi 2Sij Sij , buoyancy production PB ¼ bgi ðmt =sT Þ vT vx and constants

elliptical relaxation function (f). This model solves the governing equation of k and ε all the way down to wall surface without using wall functions [38]. The velocity fluctuation normal to the streamlines is used as the velocity scale to calculate near-wall turbulence eddy viscosity. The turbulence kinetic energy k and its dissipation ε, the wall normal stress, v2 and the elliptic relaxation function, f are given by Ref. [39]:


 v rUj k v ¼ vxj vxj




 v rUj ε v ¼ vxj vxj



  v rUj n2 vxj

S≡

are listed below [34]:

Cε1 ¼ 1:42; Cε2 ¼ 1:68;

Cm ¼ 0:0845;

b ¼ 0:012;

sk ¼ sε ¼ 1:39;

h ¼ Sk=ε;

2

h0 ¼ 4:38;

2.3.2. The SST ku model The SST ku model is formulated based on the standard ku and the kε model. The blending functions are designed for switching between the standard kε model in the bulk flow and the standard ku model near the wall [34,37]: The transport equation of the k and u are described below [34]:

 ! mt vk þ Pek  rb* ku sk vxj




 v rUj u v ¼ vxj vxj

 !  m vu mþ t þ Pu  rbu2 sε vxj

mþ

þ 2ð1  F1 Þrsu2 where

eddy

viscosity

(13)

2

(14)

3

2

3

turbulence 3

2 6k3=2 p1ffiffiffi 4 ε ; 3

0

3=2 kp ffiffiffiffiffiffiffiffiffi7 5; Ch n2 Cm 2Sij Sij

 1=4 n3 ε

length

scale

6 L ¼ CL max4min

3 7 5 and constants are listed below [39]:

sffiffiffiffiffiffiffiffiffiffiffiffi 



Cε1 ¼ Cε1 1 þ 0:045

k

n2 ;

Cε1 ¼ 1:4; Cε2 ¼ 1:9;

C1 ¼ 1:4; C2 ¼ 0:3; Cm ¼ 0:22; sε ¼ 1:3m sk ¼ 1:0; a ¼ 0:6

CL ¼ 0:23;

Ch ¼ 70;

(10) 1 max½1=a* ;SF2 =a1 u

,

~ ¼ min P k

e 4=15þðRet =Rb Þ4 , F1 is the blending ðPk ; 10rb* kuÞ; Pu ¼ ramPt k , b* ¼ b*∞ 4 1ðRet =Rb Þ

function and constants are [35]:

a∞ ¼ 0:52: Rb ¼ 8; a*∞ ¼ 1; sk1 ¼ 1:176; su1 ¼ 2:0; sk2 ¼ 1:0; su2 ¼ 1:168; a1 ¼ 0:31; bi;1 ¼ 0:075; bi;2 ¼ 0:0828;

(12)

0 1   v @ mt vn2 A ε ¼ mþ þ rkf  6rn2 vxj k sk vxj

k2ffiffiffiffiffiffiffiffiffi7, paffiffiffi p 5 3 n2 Cm 2Sij Sij

(9)

1 vk vu u vxj vxj mt ¼ rk u

  ! mt vε Cε1 Pk  Cε2 rε mþ þ T sε vxj

where eddy viscosity mt ¼ rCm n2 T; turbulence production Pk¼mtS2 2   1  qffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 6 (S≡ 2Sij Sij Þ, turbulence time scale T ¼ min4max Kε ; 6 nε ;

st ¼ 0:85


 v rUj k v ¼ vxj vxj

(11)

0 1 ðC1  1Þ @2 v2 A Pk v2  þ C2 þ 5 f L V f ¼ T 3 k rk kT

Cm h3 ð1  h=h0 Þ ¼ Cε2 þ ; 1 þ bh3

* Cε2

 ! mt vk mþ þ Pk  rε sk vxj

b* ¼ 0:09

2.3.3. The v2  f model One of the developed eddy-viscosity models, named the v2  f model, was proposed by Durbin [38], aimed at improving modeling of wall effects on the turbulence. In the v2  f model, additional transport equation of v2 is solved together with an equation for

2.3.4. The RSM model Transport equation for the Reynolds stresses and turbulence dissipation are solved in the RSM model [40]. The low Re-stress omega model was set up for the RSM model [34]. The equations for transport of Reynolds stresses can be written as:

v vxk





rUk ui uj

0

! ¼ Pij þ

1   vui uj

C v B Bm C vxk @ vxk A

3 2     v 4     rui uj uk þ p0 dkj ui þ dik uj 5 vxk 0 1     vu vui vuj vu j i A þ PB;ij  2m þ p0 @ þ vxk vxk vxj vxi (15)

H. Chen et al. / Building and Environment 90 (2015) 37e50 



where Pij ¼ r ui uk

vUj vxk





þ uj uk

41

! vUi vxk

is stress production, and the 

production due to buoyancy PB;ij ¼ rbðgi u0 j q þ gj ui q Þ The transport equation for the dissipation rate ε is given by:




   v rUj ε v m vε ε ε2 ¼ mþ t þ Cε1 ½Pk þ Cε3 PB   Cε2 r vxi k sε vxi k vxj where PB ¼

eddy

mt vT bgi Pr t vxi

Cm ¼ 0:09; Cε2 ¼ 1:92;

viscosity

2

mt ¼ Cm r kε ,

Pk ¼ mt

h

vUi vxj

þ

i

vUj vUi vxi vxj ,

(16)

and

and the following constants were used:

sk ¼ 0:82;

sε ¼ 1:0;

Cε1 ¼ 1:44;

Prt ¼ 0:85

Cε3 ¼ tan h½upar =uper , where upar is the velocity vector parallel to the gravitational vector and uper is the velocity vector perpendicular to the gravitational vector. Fig. 2. Example of computational grid (WCJSD).

2.4. Boundary conditions The velocity components, u and v, in x and y-direction provided for the MSD were based on measurements by Cho et al. [32], and those for the DSD were from validated CFD simulations [33]. The velocity profiles in y-direction used in the IJSD were based on measurements by Chen and Moshfegh [13]. In the WCJSD, uniform mass flow rate was provided at the inlet. For all four air-distribution devices, a uniform temperature was used at the inlet. Turbulence intensity and hydraulic diameter were used to determine k, ε and v2 . At the inlet, turbulence intensity of 10% was set for the simulations with the IJSD, MSD and DSD. For the WCJSD, turbulence intensity of 5% was provided at the inlet. Pressure outlet was chosen for outlet boundary condition. Constant heat fluxes were provided for the wall surfaces, and zero heat flux was set to the internal surfaces except for the floor, window and ceiling. All the surfaces were assumed to be grey with an emissivity of 0.9. Boundary conditions for case studies are summarized in Table 1. A total number of six cases were studied to find out the effect of the total heat load and supplied air conditions (i.e., supply air-flow rate and temperature) on the ventilation performance for the four supply devices. The studied heat loads were selected in order to find the ventilation performance without disturbing the flow by cooling load from ceiling. 2.5. Numerical aspects The finite-volume solver Fluent 14.5 [34] was used to numerically solve the governing equations with a segregated scheme. The pressureevelocity coupling was controlled by the SIMPLE algorithm. For most cases, the governing equations were discretized spatially with second-order upwind scheme for non-linear terms and second-order central scheme for viscous terms. While for the

case of the DSD where the RSM turbulence model was used, firstorder upwind scheme was chosen for the Reynolds Stresses due to the difficulties in obtaining converged solutions. Solutions were considered converged when the residuals for continuity and RANS equations were below 103 while for energy falling less than 106. The errors for the overall mass flow and heat flux balances were less than 0.03% and 0.3%, respectively. The predictions were performed on a high performance cluster with two processors, each of which consists of four cores and 32 GB system memory. 2.6. Mesh strategy In this study AIRPAK 3.0.16 [41] was used to construct the model and generate three-dimensional hexahedral mesh. Non-conformal mesh strategy was adapted and placed around the air supply devices and heat sources, in order to reduce the total number of mesh through the whole domain. The mesh strategy in the room for supply devices was based on the one which was validated previously [13]. Example of computational grid for the WCJSD can be seen in Fig. 2. The mesh sizes are reported in Table 2. The nonconformal interface was deleted for calculating the mean age of air, due to problems encountered using a proposed developed UserDefined Scalar (UDS) equation. As a result, a large mesh consisting of structured hexahedral cells was generated, see Table 2. Thus, the mean age of air calculation was only focussed on a few cases due to affordable computation source and time. The grid was refined enough near the solid walls to solve all boundary layers. In this study the near-wall modeling used the enhanced wall treatment (EWT) for the MSD. In the case of WCJSD, IJSD and DSD, no additional wall treatments were required, as the v2  f model, and the RSM stress-omega model with low-Reynolds

Table 1 Case studies. Case

Total heat load of floor area (W/m2)

Total heat load (W)

Internal heat load (W)

Heat load from floor (W)

Heat load from window (W)

Q (m3/s)

Tsupply (
C)

Cooling load from ceiling (W)

1 2 3 4 5 6

17 34 51 34 34 34

256 514 771 514 514 514

256 256 256 256 256 256

0 154 309 154 154 154

0 103 206 103 103 103

0.025 0.025 0.025 0.025 0.025 0.021

16.0 16.0 16.0 14.0 18.0 16.8

0 256 513 194 317 317

42

H. Chen et al. / Building and Environment 90 (2015) 37e50

4. Results and discussions

Table 2 Mesh size (in million). Supply device

Non-conformal mesh

Conformal mesh

MSD WCJSD IJSD DSD

3.7 6.6 4.2 4.3

6.5 8.1 7.1 7.9

correction was used, respectively. yþ of the first grid was less than 1.0 in most regions of the computation domain for all the cases.

2.7. Measures of performance The ventilation performance of the four air-distribution devices was examined with the predicted mean vote (PMV), predicted percentage dissatisfied (PPD), and draught (DR). The PMV was determined based on equation, defined in ISO 7730 [42]. The PPD was determined based on equation (17), see Ref. [42]. The effect of draught in the room was calculated using Fanger's equation (18). These equations were implemented into Fluent via User-Defined Function (UDF). The comfort level was set at metabolic rate 1.0 met and clothing 0.8 clo,

  PPD ¼ 100  95 exp  0:03353:PMV4  0:2179:PMV2 (17)  0:62   DR ¼ ð34  Tl Þ Ul  0:05 0:37Ul Tu þ 3:14

(18)

where Tl is the local air temperature (
C), Ul is the local mean air velocity (m/s) and Tu is the turbulence intensity (%). The DR value was assumed to be zero for the velocity less than 0.05 m/s, according to ISO 7730 [6]. Furthermore, it was of interest to evaluate the ventilation's effectiveness. In this study, heat removal effectiveness, εt, and air exchange efficiency, εa, based on mean age of air, t [43] were analyzed. εt is a measure of how effective the air-distribution method was in removing the heat produced in the occupied zone, defined by Equation (19). εa was obtained using Equation (20).

εt ¼

Te  Ti Toz  Ti

(19)

εa ¼

tn 2

(20)

where subscripts e, i and oz denote exhaust, inlet and mean value for the occupied zone. tn is the nominal time constant for a given room volume, V (m3), and air flow rate, Q (m3/s), expressed as tn ¼ V=Q , and is the average age of air in the occupied zone. The User-Defined Scalar (UDS) was implemented into Fluent for calculation of the mean age of air via UDF.

3. Data collection Values of the velocity, temperature, PMV, PPD and DR and mean age of air were average values over specific planes at heights of 0.1, 0.6, 1.1 and 1.7 m within the occupied zone. Based on the ASHRAE standard [44], the occupied zone was defined as a volume between the floor and height 1.8 m from the floor with the distance 1.0 m from the supply device and external wall (opposite wall). The occupied zone had the distance 0.3 m from the internal walls.

4.1. Validation study The model used in the case with WCJSD and IJSD has been validated with the detailed measurement, and the predicted mean velocity and temperature distributions were compared in the jet region and in the various zones in the room, for more detail see, e.g. Ref. [13]. In this paper, velocity and temperature profiles in the jet region (below the inlet and in the room) at the mid-plane are shown in Fig. 3 and Fig. 4. At all locations, the numerical predictions of the temperature and velocity profiles by SST and v2  f models are in good agreement with the measured results for both WCJSD and ICJSD, respectively. 4.2. Room air-flow patterns for MSD, WCJSD, IJSD and DSD Fig. 5 showed room air-flow patterns from the four studied devices, represented by the iso-velocity ¼ 0.25 m/s together with the corresponding temperature contour plot. These figures were generated based on the results from Case 1 where only internal heat sources and the four air-distribution systems were considered. According to Fig. 5, the high-level supply devices, i.e., the MSD generated a much larger area with the iso-velocity ¼ 0.25 m/s, compared to the WCJSD, IJSD and DSD where flows with velocities higher than 0.25 m/s only existed in a small region close to air supply devices. Since the high-level supply devices involved strong mixing between the jet and room air on its way down to the floor, supply air flows were heated rapidly to an appropriate temperature before entering occupied spaces, i.e., about 23
C. The WCJSD, IJSD and DSD directly delivered supply air to the occupied zone. Although the high-level supply devices created more favorable room temperatures for occupants, they might generate higher draught risks due to strong air motions provoked in the room. The difference in terms of air-flow patterns as well as temperature contour plots between the four air supply systems were essentially related to characteristics of air supply devices, such as the type and position, as well as air supply velocities and momentum. In the current configuration of MSD, positions of heat sources (i.e., occupant, etc.) played an important role in determining the room air-flow pattern, because buoyancy from the heat sources tended to force the flow attached to the ceiling to the other side of the room. As a result, the supply air-flow seemed to be directed from one corner to the room and consequently covered a large area near the floor, which could be seen from Fig. 5. In the WCJSD, the wall confluent jets remained attached to the wall due to Coanda effect and moved downward until reach to the floor. The jet across the floor is also attached to its adjacent boundary (i.e., the floor) after the stagnation zone. In the IJSD, the air flow was supplied with a high momentum which was lower than that in the MSD, at a certain height above the floor, and then distributed radially after impingement. Although high momentum was supplied with this devices (WCJSD and IJSD), velocities near the floor (e.g. at the level of 0.1 m) could be reduced to an acceptable level but still great for providing effective air distribution. Entrainment involved in WCJSD and IJSD are less than in the mixing. In the case of the DSD, air was supplied from a wall-mounted plane diffuser with a low momentum and discharged forwards to the room. The air-flow pattern behaved somewhat like a two-dimensional flow and involved the minimum amount of mixing between the jet and room air among the four studied air supply devices. 4.3. Vertical temperature distributions in the room Vertical temperature distributions in the mid-plane of the room for all supply devices were presented in Fig. 6. Again, these results

H. Chen et al. / Building and Environment 90 (2015) 37e50

WCJSD

y = 1.32 m

y = 1.46 m

y = 1.6 m

0.2

43

y = 1.6 m

y = 1.46 m

y = 1.32 m

M SST

0.15

x (m)

0.1

0.05

0 0

0.75 1.5 2.25

0

0.75 1.5 2.25

U (m/s)

IJSD

U (m/s)

x = 1.0 m

x = 0.6 m

0.15

0 0.75 1.5 2.25 U (m/s)

17

19 21 T (°C)

23

17

19 21 T (°C)

x = 0.6 m

x = 1.5 m

23

17 19 21 23 T (°C)

x = 1.5 m

x = 1.0 m

M

y (m)

0.1

0.05

0 0

0.5 1 U (m/s)

1.5

0

0.5 1 U (m/s)

1.5

0

0.5 1 U (m/s)

1.5

17

19

21

23

17

19

21

23

17

19 21 T (°C)

T (°C)

T (°C)

23

Fig. 3. Measured and predicted velocity and temperature profiles at different downstream distances (Top) from the inlet, WCJSD (bottom) from the inlet wall, IJSD.

were based on Case 1. It could be seen that the MSD provided quite uniform thermal conditions in the room with a temperature of about 24.5
C. As the mixing effect is reduced while thermal stratification is enhanced, the vertical temperature difference will

WCJSD

x = 0.7 m

1.8 1.6

M

1.4

SST

increase and as a result, a lower room air temperature will be provided. This could be noted from the WCJSD and IJSD. When the buoyancy force becomes the dominant force driving room air motions, a full or nearly full thermal stratification could be

x = 3.5 m

x = 2.1 m

x = 2.1 m

x = 3.5 m

20 21 22 23 T (°C)

20 21 22 23 T (°C)

x = 0.7 m

y (m)

1.2

1 0.8 0.6 0.4

0.2 0 0

IJSD

0.2 0.4 U (m/s)

0.6

0

x = 0. 7 m

1.8

0.2 0.4 0.6 U (m/s)

0

x = 2.1 m

0.2 0.4 0.6 U (m/s)

20 21 22 23 T (°C)

x = 0. 7 m

x = 3.5 m

x = 3.5 m

x = 2.1 m

M

1.6 1.4

y (m)

1.2 1 0.8 0.6 0.4

0.2 0 0

0.2 0.4 0.6 U (m/s)

0

0.2 0.4 0.6 U (m/s)

0

0.2 0.4 0.6 U (m/s)

20

21

22

T (°C)

23

20

21

22

T (°C)

23

20

21

22

23

U (m/s)

Fig. 4. Measured and predicted velocity and temperature profiles at different downstream distances from the inlet wall in the vertical middle plane (Top) WCJSD (bottom) IJSD.

44

H. Chen et al. / Building and Environment 90 (2015) 37e50

Fig. 5. Iso-surface of the velocity of 0.25 m/s together with its corresponding temperature contour plots for Case 1 in (a) MSD, (b) WCJSD, (c) IJSD and (d) DSD.

Fig. 6. Vertical temperature distributions for Case 1 in (a) MSD, (b) WCJSD, (c) IJSD and (d) DSD.

H. Chen et al. / Building and Environment 90 (2015) 37e50

achieved, which was the case for the DSD. A more stratified temperature field implies a more efficient heat removal by an air supply device. However, too strong temperature stratification, i.e., over 3
C [42] should be avoided from the aspect of local discomfort. 4.4. Comparisons of ventilation performance under identical conditions at different heat loads (Case 1, 2 and 3) Fig. 7 compared ventilation performance of the four air supply devices in terms of thermal comfort and thermal removal efficiency (εt) at three heat loads. Since exhaust temperatures for all supply devices were pre-designed to be 24.3
C in energy balance calculations, and supplementary cooling was provided by cooling the ceiling in all cases (see Table 1), differences regarding overall thermal comfort became less evident at higher heat loads. At the lowest heat load of 17 W/m2, a distinct thermal environment was observed from different air supply devices. The MSD provided the highest occupied zone temperature (24.1
C), followed by the IJSD (22.5
C), WCJSD (22.3
C), and DSD (21.6
C). As more heat was imposed on the floor and window, together with the internal heat sources reaching a total heat load of 34 and 51 W/m2, the MSD was observed to provide the same occupied zone temperature as could be seen from Fig. 7. This was due to the fact that the air-flow supplied from the MSD (i.e., slot) was cooled down by the cooling ceiling through convection rather than being heated up by the floor. Although higher occupied zone temperatures (i.e., greater than that in the MSD) had not been observed in the WCJSD, IJSD and DSD, one can surmise that this effect will be continued later on based on the reason explained above. Overall thermal comfort assessed by the index of PPD is a direct reflection of the occupied zone temperature, i.e., the less appropriate the room temperature, either (too) warm or (too) cool, the greater the values of PPD. Based on Fig. 7, it could be seen that the thermal environment of the small office for most cases was acceptable, as PPD values were controlled well below the threshold of 10% for thermal comfort [42]. However, this was not the case for the DSD at the lowest heat load of 17 W/m2, because the PPD value exceeded the permitted level. Higher air supply temperature would

45

be used to remedy this condition, especially for situations where incoming solar radiation is not considered on the floor. Local thermal discomfort is another important aspect for evaluating ventilation performance. In this study, draught discomfort was assessed at a height of 0.1 m above the floor, as recommended by ASHRAE [44] and ISO [42]. Fig. 7 showed plots of the average DR within the occupied zone at this height. The MSD investigated in this study had the highest average value of DR (about 18%), followed by the WCJSD and IJSD (about 12%), and DSD (about 9%). This trend was closely related to the area of the air supply device and the associated air supply velocity. The larger the area of the supply device, the smaller the air supply velocity and the lower the DR. For four devices, the draught risks at a height of 0.1 m might not be critical as the average values of DR were lower than 20% [44]. It is worth mentioning that even though high DR values were generated by the MSD near the floor level, it might not cause problems in regions close to the occupant, because buoyancy from the internal heat sources tended to direct the supplied flow to the opposite side of the room. In addition to draught discomfort, local discomfort caused by a large vertical temperature difference was also evaluated, and the results will be presented later in this section. In this study, heat removal effectiveness (εt) was assessed for evaluating ventilation efficiency of the four devices. According to Fig. 7, the DSD provided the highest εt (about 1.1e1.5) and MSD provided the lowest εt (about 1.0). The value of εt provided by the WCJSD was slightly higher than the IJSD, i.e., varied from about 1.0 to 1.4. The values of εt were related to thermal stratification levels created by each air supply device within the room. It could be also noted that the difference of εt was reduced with increasing heat loads, due to the presence of more mixing effects and the less stratified thermal field, and thus a relatively similar heat removal effectiveness was observed for all the devices at the highest heat load of 51 W/m2; see Fig. 7. Fig. 8 plotted temperatures at different room heights of 0.1, 0.6, 1.1 and 1.7 m for the four air-distribution systems. The temperature was an average temperature over certain planes at those heights within the occupied zone. According to Fig. 8, the vertical temperature distributions nearly followed straight lines in the MSD,

Fig. 7. Performance comparisons of the four air supply devices at three heat loads.

46

H. Chen et al. / Building and Environment 90 (2015) 37e50

MSD IJSD WCJSD DSD

1.8 1.5

Height (m)

1.2 0.9 0.6 0.3 0

16

18 20 22 24 T (°C) 17 W/m2

16

18 20 22 24 T (°C) 34 W/m2

16

18 20 22 24 T (°C) 51 W/m2

Fig. 8. Vertical temperature distributions along four heights in the room with different heat loads.

while they showed certain gradients in the WCJSD and IJSD in which DT1.70.1 varied approximately from 1
C to 3
C and 1
Ce2
C, respectively and particularly in the DSD from 3
C to 4
C. Since DT1.70.1 generated by the DSD was greater than 3
C, occupants might complain about local thermal discomfort. Increasing the air supply temperature, or enhancing the mixing between the supply air flow and room air by using a radial diffuser, could be used to reduce the excessive vertical stratification in the current investigated DSD. Moreover, the temperature gradients in the cases of the WCJSD, IJSD and DSD were noted to decrease with increasing heat loads, particularly in the higher region (i.e., above a height of 0.6 m) so that the gradient was significantly smoothed out. This was expected due to the reduced positive buoyancy forces from thermal plumes while negative buoyancy effects from the cooling ceiling increased. 4.5. Effects of different supply air temperatures

4.6. Effects of different combinations of supply air-flow rate and temperature Reducing supply air-flow rates could contribute to a better thermal environment in a room while saving fan power. Effects of different combinations of the supply air-flow rate and temperature were discussed in this session, based on the results from Case 5 (0.025 m3/s with 18
C) and Case 6 (0.021 m3/s with 16.8
C).

25

25

24

20 Dr at 0.1 m (%)

Toz (°C)

The supply air temperature is an important parameter, which not only affects thermal comfort but also influences energy usage for cooling or heating supply air to desired values. For example, under a cooling mode, a higher air temperature would allow an airdistribution system to use free cooling energy for a longer period. In this study, we investigated effects of different air supply temperatures on ventilation performance at the heat load of 34 W/m2, i.e., Case 2, 4 and 5 (see Table 1), and the results were presented in Fig. 9. It could be seen that the average occupied zone temperatures increased with increasing air supply temperatures for all supply devices. By analyzing the results of PPD, thermal comfort provided by the DSD improved significantly: the PPD value was reduced from a high level (about 11%) to an acceptable level (7%), as the occupied zone temperature increased from about 22
C to 23
C. However, the effect was less significant for other devices and PPD values were

maintained almost constant (about 6%) and acceptable. However, plots of PPD are not presented in this paper due to space limitations. Increasing the supply air temperature could also help to improve local thermal comfort. According to Fig. 9, draught discomfort was noted to be largely reduced using the MSD: the average value of DR at the 0.1 m plane within the occupied zone was reduced from 22% to 17%. This effect was not as significant concerning the WCJSD, IJSD and the DSD, in which velocities near the floor were still too high to be compensated by higher temperatures (i.e., 16 and 18
C) in this study. Even so, the maximum values of DR in the WCJSD, IJSD and DSD were below the maximum permitted value of 20% for draught discomfort. By analyzing results from thermal removal effectiveness, no significant effects were caused. Values of εt provided by all the devices were similar to those presented in Fig. 7 at the heat load of 34 W/m2. Fig. 10 showed vertical temperature distributions at different air supply temperatures. The temperature gradient was observed to be reduced with increasing the air supply temperature using the WCJSD, IJSD and DSD, in which DT1.70.1 decreased about 1
C for WCJSD and IJSD and for DSD decreased about 2
C. However, the reduction of DT1.70.1 was mainly concentrated in the region below the height of 0.6 m of the room. For the MSD, this effect could be neglected due to strong mixing between supply airflows and room air.

23 22 21

MSD IJSD WCJSD DSD

15 10 5

20

0 14 °C

16 °C

18 °C

14 °C

16 °C

Fig. 9. Effects of supply air temperatures on occupied zone temperature and DR.

18 °C

H. Chen et al. / Building and Environment 90 (2015) 37e50

MSD IJSD WCJSD DSD

1.8 1.5 1.2 Height (m)

47

0.9 0.6 0.3 0 16

18 20 22 24 26 T (°C) at the supply temperaure of 14 °C

16 18 20 22 24 26 T (°C) at the supply temperaure of 16 °C

16

18 20 22 24 T (°C) at the supply temperaure of 18 °C

26

Fig. 10. Effects of air supply temperatures on vertical temperature gradients.

By analyzing data, no considerable variations on the occupied zone temperature and PPD were caused by different air supply conditions, and the results were similar to those shown in Fig. 7 at a heat load of 34 W/m2. However, draught discomfort was greatly reduced by lowering the supply air-flow rate in the MSD, where DR decreased approximately from 17% to 12%; see Fig. 11. According to Fig. 12, a small impact on vertical temperature gradients could be observed which the largest variation of DT1.70.1 using the DSD was less than 0.5
C, and the value was even smaller for the other supply devices. 4.7. Comparisons of air exchange efficiency Ventilation performance of the four devices with respect to indoor air quality was further evaluated by the index of air exchange efficiency (εa) based on the mean age of air. The comparison of εa was only made for Case 1 (17 W/m2), due to affordable computation effort and time for the meshes used which were without nonconformal interfaces. Results of εa for the occupied zone were shown in Fig.13. As shown, the DSD provided the highest value of εa in which εa provided by MSD was greater than 0.5 for the perfect mixing, because of certain stratifications created within the room. Again, this trend was closely related to thermal stratification level created within the room. 4.8. Comparisons of ventilation performance at the same occupied zone temperature In this section, performance of the four air supply devices was compared at the same occupied zone temperature of about 24.2
C, achieved by using different heat loads or air supply flow rates,

35 MSD IJSD WCJSD DSD

30

Dr at 0.1 m (%)

25 20 15 10 5 0 Q = 0.025 m3/s and T = 18°C

Q = 0.021 m3/s and T = 16.8°C

Fig. 11. Effects of different air supply conditions on DR.

based on a reference condition specified in Case 2 (34 W/m2). Table 3 showed comparisons at the same supply air-flow rate of 0.025 m3/s and the same supply air temperature of 16
C. The WCJSD, IJSD and DSD required slightly higher heat loads (i.e., 36.4, 35.5 and 38.0 W/m2, respectively) than the MSD (34.0 W/m2), for achieving the same occupied zone temperature, which corresponds to about 7%, 4% and 12% higher cooling capacity. Table 4 compared the air-flow rate as well as fan power required for each supply device for achieving the same occupied zone temperature, at the heat load of 34 W/m2 and with the same supply air temperature of 16
C. The DSD required the lowest air-flow rate of 0.0208 m3/s among the four devices. The MSD required 1.2 times higher air-flow rate and needed 1.74 times higher fan power than the DSD, by using the relations between the flow rate, Q, pressure difference, Dp, and the fan power, E, i.e. Dp fQ2, E fQ3 [26]. For the WCJSD and IJSD, it required 1.06 and 1.09 times higher air-flow rate respectively, which corresponds to 1.18 and 1.32 times more fan power but was still less compared with the MSD. The results presented in Tables 3 and 4 for the MSD was based on the case 2. For the WCJSD, IJSD and DSD, the results of thermal comfort and εt between the cases used in Tables 3 and 4 were very similar. In this part, only the comparisons concerning different supply air-flow rates were presented in Fig. 14. Since all air supply devices provide a similar occupied zone temperature of about 24.2
C, PPD values were thus very close to each other and below the maximum value of 10% required for thermal comfort. For draught discomfort, the MSD generated the highest value of DR (19%) among the four devices, which was in the permitted level of 20% [44], see Fig. 14. For local discomfort due to the excessive vertical temperature gradient, the DSD might cause complaints by occupants as DT1.70.1 was greater than 3
C. At the same time DT1.70.1 was reduced to an acceptable level in the WCJSD and IJSD, but still great for enabling effective heat removal. The DSD and WCJSD provided the highest thermal removal effectiveness (about 1.2), which was slightly higher than that in the IJSD (about 1.1), and the lowest value was observed using the MSD (about 1.0). The same trend was observed for the air exchange efficiency for the occupied zone. it should be noted that these findings from the study were limited to certain types of air supply devices and under specific conditions. Further studies with a wider scope are required to make these conclusions applicable for general conditions. The above presented results indicated that the WCJSD and IJSD perform similarly to the DSD but superior to the MSD in the ventilation of a small office room, which is consistent with the findings observed by Karimipanah et al. [26]. In the study of Karimipanah et al. [26], the wall confluent jet was capable to give a slightly better performance than displacement ventilation: it requires a slightly lower air flow rate for achieving the same air

48

H. Chen et al. / Building and Environment 90 (2015) 37e50

MSD IJSD WCJSD DSD

1.8 1.5 1.2

1.8 1.5 1.2 Height (m)

Height (m)

0.9 0.6 0.3 0 16 18 20 22 24 26 T (°C) at 0.025 m3/s with 18 °C

0.9 0.6 0.3

0 16 18 20 22 24 26 T (°C) at 0.021 m3/s with 16.8 °C

Fig. 12. Effects of different flow supply combinations on vertical temperature gradients.

MSD IJSD WCJSD DSD

1 0.8

5. Conclusion

εa

0.6 0.4 0.2 0 17 W/m2 Fig. 13. Comparisons of εa at the heat load of 17 W/m2.

distribution index and gives slightly higher ventilation effectiveness. However, in this study DSD was observed to be somewhat more effective than the WCJSD and IJSD in terms of energy performance and thermal removal effectiveness. This could be explained by the reason that different types of air supply devices, room configuration and different strategies for designing case studies were considered in these two studies. Overall, this study confirmed that the WCJSD and IJSD have a great potential to achieve a good indoor environment and use energy efficiently.

The present study compared ventilation performance of four different air supply devices for an office environment, with a focus on thermal comfort, ventilation efficiency as well as energy-saving potential related to fan power. Under identical conditions, the results indicated that WCJSD and IJSD behaved like a combined mixing and displacement system: it could provide an acceptable thermal environment, while removing excess heat more efficiently as compared to the conventional mixing system. Although DSD was shown to be superior to all other devices in terms of heat removal and fan power reduction, it seemed to have difficulties fulfilling the accepted criteria for the vertical temperature gradient between the ankle and neck levels (for a standing person), i.e., below 3
C. By supplying warmer air, i.e., 18
C with the DSD, a considerable improvement on the overall and local thermal comfort levels was observed. Comparisons for achieving the same occupied zone temperature showed that the WCJSD and IJSD required slightly higher air-flow rate (1.06 and 1.09 times) and fan power (1.18 and 1.3 times) than DSD, but it was more energy efficient than the MSD. Meanwhile, this device could provide an acceptable overall and local thermal comfort. Therefore, the WCJSD and IJSD have a potential to provide a good indoor environment and save energy. Since this study is limited to certain types of air supply devices that were installed at the specific location in the room, therefore further investigation needs to be carried out to derive more general

Table 3 Performance comparisons in terms of cooling capacity. Supply device

Heat load (W/m2)

Difference of cooling capacity compared with the MSD based on a floor area of 15.12 m2 (W)

Increased cooling capacity compared with the MSD

Mixing Impinging jet Wall confluent jets Displacement

34.0 35.5 36.4 38.0

e 23 36 60

e 4% 7% 12%

Table 4 Performance comparisons in terms of fan power. Supply device

Supply air-flow rate (m3/s)

Air-flow rate compared with the DSD

Fan power ratio compared with the DSD

Mixing Impinging jet Wall confluent jets Displacement

0.025 0.0228 0.022 0.0208

1.2 1.09 1.06 e

174% 132% 118% e

35

1.6

30

1.4

1

20 15

0.8 0.6

10

0.4

5

0.2

0

0

1.8

MSD WCJSD IJSD DSD

1.5 1.2

1 0.8 0.6

0.9 εa

Height (m)

49

MSD IJSD WCJSD DSD

1.2

25

εt

Dr at 0.1 m (%)

H. Chen et al. / Building and Environment 90 (2015) 37e50

0.6 0.3

0.4 0.2

0 16

18

20 22 T (°C)

24

26

0

Fig. 14. Performance comparisons at the same occupied zone temperature by using different supply air-flow rates.

conclusions of these supply devices. It is interesting to study the WCJSD and the IJSD with most common MSD and DSD at different locations in the room (e.g., corner or other walls) and at different heights. Acknowledgments The authors gratefully acknowledge the financial support received from Formas (Contract number 242-2008-835), KK Foundation (Contract number 2007/0289), University of G€ avle €vle, Sweden) and Linko €ping University (Linko € ping, Sweden). (Ga The authors sincerely thank Dr. Mathias Cehlin from University of €vle, Sweden for valuable suggestions on this study. Ga References [1] Sayigh A. Sustainability, Energy and architecture: case studies in realizing green buildings. Elsevier Science; 2013. [2] Energy in Sweden 2010, Swedish Energy Agency (SEA), Eskilstuna, Sweden. [3] National board of housing building and planning, energy use in buildings, Subgoal 6-Report for the detailed evaluation of good built environment, Boverket, Karlskrona, 2007, 978e91-875751-59-4. [4] Fanger PO. Thermal comfort. New York: McGraw-Hill; 1973. [5] Wargocki P. Human perception, productivity and symptoms related to indoor air quality. Denmark: Technical University of Denmark; 1998, ISBN 87-7475201-4. [6] Fanger PO. Provide good air quality for people and improve their productivity. Air distribution in room (Roomvent). 2000. [7] Fisk WJ, Rosenfeld AH. Estimates of improved productivity and health from better indoor environments. Indoor Air 1997;7:158e72. [8] Etheridge D, Sandberg M. Building ventilation: theory and measurement. 1996. €vle [9] Sandberg M, Stymne H, Blomqvist C, Mattsson M. Ventilation i funktion. Ga €r byggnadsforskning; 1993. (in Swedish). Statens institut fo [10] Sandberg M, Mattsson M. The effect of moving heat sources upon the stratification in rooms ventilated by displacement ventilation. Proceeding Roomvent. Int Conf Air Distribution Rooms 1992;3:33e52.

[11] Chen H, Moshfegh B, Cehlin M. Computational investigation on the factors influencing thermal comfort for impinging jet ventilation. Build Environ 2013;66:29e41. [12] Chen HJ, Moshfegh B, Cehlin M. Numerical investigation of the flow behavior of an isothermal impinging jet in a room. Build Environ 2012;49:154e66. [13] Chen HJ, Moshfegh B, Cehlin M. Investigation on the flow and thermal behavior of impinging jet ventilation systems in an office with different heat loads. Build Environ 2013;59:127e44. [14] Ghahremanian S, Moshfegh B. Evaluation of RANS models in predicting low Reynolds, free, turbulent round jet. J Fluids Eng 2013;136:011201e13. [15] Ghahremanian S, Moshfegh B. A study on proximal region of low Reynolds confluent jets, part 2: numerical prediction of the flow field. ASHRAE Transactions 2014;120. Part 1, NY-14-022. [16] Ghahremanian S, Moshfegh B. A study on proximal region of low Reynolds confluent jets Part 1: evaluation of turbulence models in prediction of inlet boundary conditions. ASHRAE Transactions 2014;120. Part 1, NY-14-021. [17] Ghahremanian S, Svensson K, Tummers M, Moshfegh B. Near-field development of a row of round jets at low Reynolds numbers. Exp Fluids 2014;55:1e18. [18] Ghahremanian S, Svensson K, Tummers MJ, Moshfegh B. Near-field mixing of jets issuing from an array of round nozzles. Int J Heat Fluid Flow 2014;47: 84e100. [19] Janbakhsh S, Moshfegh B. Experimental investigation of a ventilation system based on wall confluent jets. Build Environ 2014;80:18e31. [20] Janbakhsh S, Moshfegh B. Numerical study of a ventilation system based on wall confluent jets. HVAC&R Res 2014;20:846e61. [21] Janbakhsh S, Moshfegh B, Ghahremanian S. A newly designed supply diffuser for industrial premises. Int J Vent 2010;9:59e68. [22] Svensson K, Ghahremanian S, Moshfegh B, Tummers M. Numerical and experimental investigation of flow behaviour in a confluent jet ventilation system for industrial premises. In: Proceedings of the 10th international conference on industrial ventilation, Paris; 2012. [23] Svensson K, Rohdin P, Moshfegh B, Tummers MJ. Numerical and experimental investigation of the near zone flow field in an array of confluent round jets. Int J Heat Fluid Flow 2014;46:127e46. [24] Cho Y, Awbi HB, Karimipanah T. Theoretical and experimental investigation of wall confluent jets ventilation and comparison with wall displacement ventilation. Build Environ 2008;43:1091e100. [25] Karimipanah T, Awbi HB, Sandberg M, Blomqvist C. Investigation of air quality, comfort parameters and effectiveness for two floor-level air supply systems in classrooms. Build Environ 2007;42:647e55. [26] Karimipanah T, Awbi HB, Moshfegh B. The air distribution index as an indicator for energy consumption and performance of ventilation systems. J Human-Environmental Syst 2008;11(20):77e84.

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[27] Karimipanah T, Sandberg M, Awbi HB. A comparative study of different air distribution systems in a classroom. Proceeding Roomvent. In: International conference on air distribution in rooms 9e12 July, Reading, UK, 2; 2000. p. 1013e8. [28] Karimipanah T, Awbi HB. Theoretical and experimental investigation of impinging jet ventilation and comparison with wall displacement ventilation. Build Environ 2002;37:1329e42. [29] Chen HJ, Moshfegh B. Comparing k-ε models on predictions of an impinging jet for ventilation of an office room. In: Proceedings of Roomvent 2011, 11th international conference on air distribution in rooms 19e22 June 2011 Trondheim, Norway; 2011. [30] Rohdin P, Moshfegh B. Impinging jet: a new ventilation strategy for industries e a case study of a light alloy foundry. In: Proceedings of the 9th international conference on industrial ventilation, ETH Zürich Switzerland; October 18e21. 2009. [31] Karimipanah T, Moshfegh B. On the performance of confluent jets ventilation system in office space. In: Proceedings of Roomvent, 10th international conference on air distribution in rooms. Helsinki, Finland; 2007. [32] Cho Y, Awbi HB, Karimipanah T. A comparison between four different ventilation systems. In: Proceedings of the 8th international conference on air distribution in rooms; 2002. p. 181e4. [33] Cehlin M, Moshfegh B. Numerical modeling of a complex diffuser in a room with displacement ventilation. Build Environ 2010;45:2240e52. [34] ANSYS. Fluent theory guide. ANSYS Inc.; 2010. [35] Yakhot V, Orszag S. Renormalization group analysis of turbulence. J Sci Comput 1986;1:3e51. [36] Moshfegh B, Nyiredy R. Comparing RANS models for flow and thermal analysis of pin fin heat sinks. In: Proceedings of the 15th Australasian fluid mechanics conference, Sydney, Australia; 2004. [37] Menter FR. Two-equation eddy-viscosity turbulence models for engineering applications. AIAA J 1994;32:1598e605. [38] Durbin PA. Separated flow computation with k-e-v2 model. AIAA J 1995;33: 659e64. [39] ANSYS fluent v2f turbulence model manual. Ansys Inc; 2009. [40] Launder BE, Reece GJ, Rodi W. Progress in the development of a Reynolds stress turbulence closure. J Fluid Mech 1975;68(3):537e66. [41] Airpak. 3.0. User's guides. Fluent Inc; 2007. [42] ISO7730. Ergonomics of the thermal environment-analytical determination and interpretation of thermal comfort using calculation of PMV and PPD indices and local thermal comfort criteria. 2005. [43] Awbi HB. Ventilation of buildings. London: Spon Press; 2003.

[44] ASHRAE. Thermal environmental conditions for human occupancy. American Society of Heating, Refrigerating and Air-Conditioning Engineers; 2004.

Nomenclature d: diameter of the nozzle (mm) Q: supply airflow rate (m3/s) T: temperature (
C) Tu: turbulence intensity (%) U: velocity (m/s) V: room volume (m3) x, y, z: Cartesian coordinates (m) Greek symbols εa: air exchange efficiency () εt: Heat removal effectiveness () tn: nominal time constant (s) : average age of air in the occupied zone (s) Subscripts e: exhaust i: inlet l: local o: mean value for the occupied zone Abbreviations DR: draught rate M: measurement PC: personal computer PPD: predicted percentage dissatisfied PMV: predicted mean vote DSD: displacement supply device IJSD: impinging jet supply device MSD: mixing supply device WCJSD: wall confluent jets supply device