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Abstract

The air supply of the variable air volume (VAV) system could decrease with indoor load, resulting in reduced indoor comfort. An active diffuser is a good solution, but its adjustment process is often neglected. In most cases, the diffuser maintains the air supply velocity constant by controlling the air supply area. This traditional control method based on constant velocity (CMV) often fails to reach jet throw expectation due to the continuous change of airflow diffusion performance during the adjustment process. In this study, a new control method based on constant jet throw length (CMT) is proposed and applied in two typical rooms (office and conference). The results show that the CMT could increase the throw length in the spreading zone and enhances the airflow at the wall zone and ground zone, providing a more uniform temperature and more comfortable indoor environment. In 40% air volume condition, the air diffusion performance index (ADPI) in the office using CMT has improved from 89.66% to 95.69% compared to CMV, while ADPI in the conference improved from 92.11% to 97.37%. This study can help designers determine the optimal adjustment scheme to improve the adaptability of the active diffuser to the VAV system.

Keywords

HAVC environment active diffuser control method jet throw CFD airflow distribution VAV systems

Introduction

HVAC accounts for a 30%∼60% of the total building energy consumption.^1,2 Reducing the energy consumption of air conditioning system has become an important part of building carbon reduction.³ Variable air volume air conditioning (VAV) is a system developed to achieve energy-saving operation by reducing the energy consumption of the fan when the indoor load decreases. It has the advantages of strong flexibility, low energy consumption and high air quality.⁴ Compared with the constant air volume system, the power consumption of the VAV system can be reduced by 40%.⁵ The minimum air volume setting value can affect the total energy consumption of the system.^6,7 At present, the minimum airflow of VAV is usually set to 30%∼50% of the maximum. The research results of Kang et al.⁸ provide a reference for setting minimum air volume below 20%. Due to the good energy-saving benefits caused by the reduction of air supply volume, the VAV system is widely used in various buildings such as shopping malls, offices and houses.^9,10,11

Reduction of air supply can lead to lower comfort and indoor pollutant accumulation, directly affecting the health and work efficiency of personnel.¹² For cooling conditions, the reduction of air momentum caused by decreased air volume would lead to cold air subsidence. On the contrary, excessive air volume would cause draught problems in the work area. For heating conditions, thermal air floating at low wind velocity would lead to vertical stratification of room temperature, resulting in ‘warm head-cool feet’ discomfort.¹³ Furthermore, some studies have shown that a lower air supply would reduce indoor pollutant removal efficiency. Indoor health problems caused by insufficient air supply exist in hospitals,^14,15 office buildings,^16,17 residences¹⁸ and public places.¹⁹

Reasonable air supply terminal can improve the indoor air distribution. Active supply diffuser is a type of air outlet with adjustable air supply area or variable geometry, which has unique advantages in coping with load changes. Previous studies have evaluated the indoor environmental performance of some terminals. Wu et al.²⁰ experimentally tested an active ceiling diffuser that can control the air supply area through a central baffle. The test results show improved indoor ventilation efficiency and comfort advantages compared with commonly used diffusers. Nina²¹ proposed a nozzle controlled by three rings. When the air supply decreases, closing some rings can improve the jet velocity at low air volume. The computational fluid dynamics (CFD) simulation results show that it is suitable for large space environment construction. Rabani et al.²² tested a radial diffuser that a sleeve coupling module can control to ensure constant air velocity when the air volume decreases. The test results show that when the air conditioner operates at the lowest air volume, the diffuser can provide 75% comfort of the scope.

The regulation mode of active diffusers is mainly divided into step adjustment (SA) and continuous adjustment (CA). SA means that the diffuser has several fixed states, and the air volume range is divided into several intervals to maintain the air outlet state unchanged in different intervals. The most basic step adjustment is divided into three operating states of maximum, minimum and partial air volume, and has three opening forms corresponding to them.²¹ CA means that with the change in air supply, the air supply area can be continuously adjusted to ensure constant velocity. Few existing studies have focused on the optimization of the operation of active diffusers. The adjustment process of the outlet is accompanied by the change in the geometric shape and the outlet area, which leads to significant differences in airflow diffusion performance for different opening states.

Conventional control methods to keep the discharge velocity constant could cause the throw length to deviate from the design value. This problem has not received sufficient consideration. The control method of the active diffuser should be discussed to further improve its indoor environmental construction ability. This paper proposes an optimization method considering the full operation state diffuser to ensure the constant throw length. This method comprehensively considers the variation in the diffusion performance of the diffuser under different opening states. Taking into account the variation in the VAV system and set throw length, the optimal control strategy of active diffuser with air volume variation can be found.

The purpose of this paper is to study the air distribution of a new type of ceiling diffuser under two kinds of tuyere adjustment strategies: control method based on constant velocity (CMV) and constant jet throw length (CMT). Both full-scale experiments and numerical simulation were conducted, including two parts: (1) Jet characteristics tests of diffusers under different opening conditions. (2) Thermal comfort test of office and conference room. The indoor velocity and temperature distribution, Air Diffusion Performance Index (ADPI) and ventilation efficiency were used to evaluate the performance of CMV and CMT.

Methodology

Thermal comfort and jet throw

ADPI is defined as the ratio of the total number of points where the Effective Draught Temperature (EDT) in the workplace meets the requirements; EDT is a calculated temperature difference that combines air temperature difference and air velocity.

The diffuser selection method based on ADPI has been used by designers since 1970.²³ This method considers the diffuser jet throw length $s$ affected by jet deflection state and the characteristic length $L$ representing the geometric size of the room. $s$ is defined as the distance of airflow decaying to 0.5 m/s or 0.25 m/s along the jet direction under isothermal conditions. The characteristic length $L$ is defined as the horizontal distance from the diffuser to the nearest vertical plane. The ratio $s / L$ can be used to establish a connection with ADPI. The correspondence between $s / L$ and ADPI for several common diffusers at different loads is provided in ASHRAE-Handbook Fundamentals,²⁴ which can be used by designers for diffuser selection and airflow design. In this study, the jet throw length $s$ was used as the primary indicator to evaluate whether the diffuser performance meets the thermal comfort requirements.

Xiao et al.²⁵ and Sun et al.²⁶ studied the flow state of the ceiling diffuser jet. The common research results show that for the cooling condition of low outflow angle, the jet flow of the ceiling diffuser can be divided into three stages: free jet zone, transition zone and adherent jet zone. After the airflow is discharged from the diffuser at a specific angle, the development of the jet boundary is limited by the ceiling. The low pressure arising from air entrainment would cause the bending of the jet axis, and finally hit the wall at a certain angle. The airflow would gradually develop into an adherent jet after a brief transition. The jet throw length of the airflow $s$ consists of three parts, as shown in equation (1).

\begin{array}{c} s = s_{r} + s_{t} + s_{c} \end{array}

(1)

where

s_{r}

s_{t}

and

s_{c}

are the effective distance of free jet section zone, transition section zone and adherent jet section zone, respectively, m.

Wu²⁷ combined with Gortler solution of free turbulent jet and obtained the distance expression of the free jet zone of ceiling diffuser. Equation (2) shows that the length of free jet zone is related to the geometric parameters of the diffuser. It is not affected by the change in the air supply during operation.

\begin{array}{c} s_{r} = \frac{σ_{1} b_{0} \sin α}{3 α} (\frac{1}{4 \cos^{2} \frac{α + π}{3}} - 1) \end{array}

(2)

where

σ_{1}

is the diffusion coefficient of free jet section;

b_{0}

is the diffuser width, m; and

α

is the angle between the air supply direction and the ceiling, rad.

The throw length of transition zone $s_{t}$ is short and can be simplified to a constant value as expressed by equation (3).²⁸ Furthermore, the free jet has little effect on the decay of the axial velocity after the impact point.²⁷ The horizontal velocity before and after the transition zone can be considered consistent.²⁹

\begin{array}{c} s_{t} = c o n s t a n t \end{array}

(3)

The fully developed zone of the flow can be described by attached plane jet, as shown in equation (4).^27,30

\begin{array}{c} \frac{u_{x c}}{u_{0}} = \frac{σ_{2} \sqrt{2 F_{0}}}{x_{c} + x_{0 c}} \end{array}

(4)

where

σ_{2}

is the diffusion coefficient of the adherent jet zone;

F_{0}

is the effective area of diffuser, m²;

x_{c}

is the distance from starting point of the adherent jet zone to calculated zone, m;

x_{0 c}

is the constant related to jet pole, m;

u_{x c}

is the velocity at

x_{c}

, m/s; and

u_{0}

is the discharge velocity of the diffuser, m/s.

$x_{c}$ is calculated from the starting point of the adherent jet zone. It is not convenient to directly determine the actual distance from the diffuser. The air outlet is generally used as the starting point and equation (4) is rewritten as equation (5).

\begin{array}{c} \frac{u_{x}}{u_{0}} = \frac{σ_{2} \sqrt{2 F_{0}}}{x - (s_{r} + s_{t}) + x_{0 c}} \end{array}

(5)

where

x

is the distance from diffuser to calculated zone, m, and

u_{x}

is the velocity at

x

, m/s.

To describe the variation of the jet throw with the air supply volume, the average air velocity $u_{0}$ at the air outlet is expressed as the ratio of the air supply volume $Q$ (m³/s) to the effective area $F_{0}$ . The total throw length can be calculated from the terminal velocity of 0.5 m/s. Equation (6) shows that the reduction of the air supply volume in the VAV system would reduce the throw length. It also indicates that it is an effective method to ensure the throw length by adjusting the air supply area during partial airflow operation of the air conditioning system. However, not considering the varying $K$ in the adjustment process would cause throw deviation.

\begin{array}{c} s = s_{r} + s_{t} + 2 \sqrt{2} σ_{2} \frac{Q}{\sqrt{F_{0}}} - x_{0 c} \end{array}

(6)

Equations (5) and (6) can be organized into the simpler expression as shown in equations (7) and (8).

\begin{array}{c} \frac{u_{x}}{u_{0}} = \frac{K \sqrt{F_{0}}}{x + x_{0}} \end{array}

(7)

\begin{array}{c} s = 2 K \frac{Q}{\sqrt{F_{0}}} - x_{0} \end{array}

(8)

where

K = \sqrt{2} σ_{2}

represents velocity decay coefficient related to type of diffuser;

x_{0} = x_{0 c} - (s_{r} + s_{t})

represents the constant not affected by the change of air supply volume, m. The attenuation law of jet axial velocity can be accurately described by experimentally fitting the values of the two parameters.

Equation (8) shows that the jet throw length is mainly influenced by two dynamically varying parameters, the air supply area $F_{0}$ and the velocity decay coefficient $K$ . This paper defines the transient diffusion coefficient (TDC) to characterize the effect of different opening states on throw enhancement, as shown in equation (9).

\begin{array}{c} T D C = \frac{K}{\sqrt{F_{0}}} \end{array}

(9)

Studied cases

This paper proposes an active ceiling diffuser with variable air supply area, which can be adjusted through the sub-tuyere between the outlet blades. As shown in Figure 1(a). The outer edge of each sub-tuyere is equipped with a rotating shaft connected to the baffle. The opening and closing of sub-tuyere can be controlled by rotating the baffle, so as to adapt to the changing air volume of the VAV system. The sub-tuyeres can be divided into four categories according to the location (Figure 1(b)). The airflow characteristics under different sub-tuyere combinations are different, which are discussed in the following.

Figure 1.

Introduction to the diffuser: (a) geometric dimensions and operating modes and (b) sub-tuyere number and corresponding opening state.

Two common room types, office and conference room, were studied and reported in this paper. Each room type considers two operating conditions, design load and partial load. The indoor heat sources in this study were considered only for the sensible heat load of occupants and equipment. The walls were considered to be adiabatic while the heat transfer from the external windows was taken into account. The power of all the heat dissipation equipment involved in the experimental process is controllable. Considering the lower limit of air supply volume commonly used for variable air volume systems, the partial load was set to 40% of the maximum value. The heat release³¹ and the location of the heat source in the room are shown in Figure 2-S1. The arrangement of the indoor equipment is shown in Figure 2(a)-S2 and (b)-S2. Due to the limitations of the experimental conditions, the conference conditions were simulated in FLUENT. The models created in the program are shown in Figure 2(c)-S2 and (d)-S2.

Figure 2.

The appearance of room setting in cases: (a) office under design load; (b) office under partial load; (c) conference room under design load and (d) conference room under partial load. (The left part (S1) represents the plan view of the room layout and the right part (S2) represents the equipment layout in experiments or models in simulations.).

This paper compares the ability of three control methods to create the environment for part-load conditions: no diffuser adjustment (denoted as Partial), CMV and CMT. In addition, the design condition (denoted as Design) was set as the control group. The parameters for each condition are shown in Table 1.

Table 1.

Parameters for different working conditions in two types of rooms.

Case office (experiment)
Case no.	Volume ratio	Air volume (m³/h)	Heat source layout	Load (W)	Temperature (°C)	Open sub-tuyere	Turbulence intensity
Design	100%	493.2	a-S1; a-S2	1050	20	ABCD	—
Partial	40%	197.3	b-S1; b-S2	420	20	ABCD
CMV	40%	197.3	b-S1; b-S2	420	20	BD
CMT	40%	197.3	b-S1; b-S2	420	20	C
Case conference (simulation)
Case no.	Volume ratio	Air volume (m³/h)	Heat source layout	Load (W)	Temperature (°C)	Open sub-tuyere	Turbulence intensity
Design	100%	640.8	c-S1; c-S2	1380	20	ABCD	2.1%
Partial	40%	256.3	d-S1; d-S2	520	20	ABCD	5.5%
CMV	40%	256.3	d-S1; d-S2	520	20	BD	4.4%
CMT	40%	256.3	d-S1; d-S2	520	20	C	7.3%

Experimentation

The experimental measurements were performed in a 10 m × 5 m × 2.8 m (L $\times$ W $\times$ H) thermal chamber equipped with heating, ventilation and air conditioning (HVAC) system. The suspended ceiling and walls are insulated. The thermal chamber is located inside an air-conditioned room with controlled indoor temperature, which ensures a constant temperature outside the laboratory during the experiment. The air supply volume is controlled by a fan, which can be adjusted from 0°r/min to 50°r/min by means of a frequency converter.

Two air supply outlets are installed at the centre axis of the laboratory ceiling (x = 2.5 m and 7.5 m, y = 2.5 m). The diffusers are connected to the static pressure box, which can ensure the stability of airflow and uniform air supply in all directions. Moreover, each branch duct is equipped with a damper, which can realize the balanced distribution of conditioned air between different diffusers. Figure 3 gives an overview of the laboratory.

Figure 3.

The layout of thermal charmer: (a) schematic illustration of the experimental system and (b) location of inlet and outlet.

The experiments in this study were divided into two parts: jet characteristics tests and indoor environmental parameters tests. The former aimed to obtain the parameters

K

and

x_{0}

in equation (7), while the latter is for the evaluation of airflow organization. The equipment characteristics are listed in Table 2.

Table 2.

Technical properties of measuring devices.

Model	Variables	Data range	Accuracy
Bczp-010(PT1000)	Air temperature	0∼50°C	±0.1°C
CTV114	Indoor air velocity	0∼2 m/s	±0.05 m/s
CTV115	Air supply velocity	0∼15 m/s	±0.2 m/s
DAQM901A	Data acquisition		Thermal offset 0 uV

(1) Jet characteristics tests

The jet characteristics tests were conducted under isothermal conditions, which consisted of two parts: (a) Determination of the jet axis position and (b) velocity attenuation tests under different sub-tuyere combinations. As the airflow is not affected by buoyancy during the movement, it can be considered that the jet axis is parallel to the ceiling.³²

Velocity tests are performed at four different distances of 0.01 m, 0.02 m,³³ 0.05 m³⁴ and 0.08 m from the ceiling to determine the position of the attachment axis. 12 measuring points were arranged every 0.3 m along the jet direction from the diffuser (as shown in Figure 4(a)). The velocity transducers were fixed by a triangular bracket, and the position of the measuring points can be adjusted by the position of the bracket and the height of the vertical rod. Based on the test results of the axis position, the velocity attenuation curves of the diffuser under different open states were tested at the corresponding heights.

Figure 4.

Schematic of measuring locations arrangement: (a) velocity attenuation tests; (b) indoor environmental parameters test and (c) pictures of the measuring devices.

(2) Indoor environmental parameters tests

As shown in Figure 5, experiments for office conditions were conducted in the thermal chamber. The electric blanket was used to represent the heat dissipation from the external windows.³⁵ Printers and laptops were replaced by heat-controlled radiators and heating panels, respectively. The heat dissipation of personnel was undertaken by the heating pad with a controllable opening state.

Figure 5.

Heat source arrangement in thermal chamber.

In the indoor environmental parameters’ tests, air temperature and velocity were measured. Each test condition of the thermal chamber was set in the previous evening. A data acquisition device (DAQM901A) was used to collect and monitor signals from different sensors. The thermal parameters were guaranteed to be stable for two hours before the test. After adjusting the sensor position each time, the next reading was performed after a 5-min interval to ensure the stabilization of sensor readings.³⁶ The data were read every 2°s and recorded continuously for 3°min, and the average value was taken as the final result.²⁰

Five temperature transducers were installed on the same vertical axis at heights of 0.6 m (sitting person, lower back level), 1.1 m (sitting person, head level), 1.6 m (standing person, head level), 2.1 m (upper space level 1) and 2.6 m (upper space level 2) as is shown in Figure 4(b), which represent the heights of different parts of the human body and the upper space of the room, respectively. Four sets of temperature transducers were arranged in the thermal chamber, and the position was controlled by moving the slide rail. There were 24 sampling axes in the room, evenly arranged along the slide rail direction, with a total of 120 sampling points.

Computational fluid dynamics (CFD) simulation

CFD prediction is an effective tool for spatial flow field analysis.^37,38 In this paper, the airflow simulation of the diffuser was performed in the FLUENT program and the RNG k-ε turbulence model was used to describe the air jet.^21,39 The SIMPLE algorithm was used to couple pressure and velocity.

The complex geometry of the diffuser created great difficulties for CFD model establishment and computer operations.⁴⁰ This study adopted the local definition method to establish the diffuser model. The velocity boundary condition was adopted for the air supply inlet. The outflow state of the diffuser was defined by the horizontal and vertical velocities in different areas. This method can be well adapted to the boundary condition settings of different sub-tuyere opening states and simplify the modelling process.⁴¹ The air outlet has little effect on the indoor airflow distribution, which provided a pressure outlet for exhausting air.

The turbulence intensity of the initial jet has an effect on the airflow distribution, which is related to the characteristics of the diffuser.⁴² In the velocity attenuation tests, the velocity and turbulent intensity (Tu) (as shown in equation (10)) of the diffuser were determined to achieve an accurate setting of the inlet boundary conditions.

\begin{array}{c} T u = \frac{S_{V}}{\bar{u}} \times 100 % \end{array}

(10)

where

\bar{u}

is the time average of the measured discharge velocity of diffuser and

S_{V}

is the standard deviation for velocity.

The research objects of this paper are steady state conditions. The occupants and equipment were set to have constant heat flux. The values and positions of these heat sources are shown in Figure 2. This study did not consider heat transfer through the maintenance structure. The walls and floor were set to adiabatic. Detailed information on the boundary conditions of the conference model is summarized in Table 3.

Table 3.

Boundary conditions in the model.

Geometry model	Boundary conditions	Additional description
Human body	426 W/m³	Total heat: 75 W
Computer	2037 W/m³	Heat source power: 55 W
Multimedia	160 W/m³	Heat source power: 80 W
Wall	Adiabatic
Floor	Adiabatic
Air supply outlet	Velocity inlet	The set value is related to the diffuser opening state
Air return inlet	Pressure outlet

The quality of the mesh division determines the convergence of the calculation and the accuracy of the simulation.⁴³ Unstructured tetrahedral cells were used to discretize the computational domain (Figure 6). The corresponding position of the mesh section is shown in Figure 7(b). The jet of ceiling diffuser would attach to the wall due to the Coanda effect, and the inflation mesh was set near the conference ceiling. In addition, the surface grid of the diffuser and the volume grid near the heat source were locally densified. Analyses were performed with four different tetrahedral mesh sizes (732026, 1438716, 3577393 and 5417328). The temperature in a typical axis was chosen as a parameter for the test (Figure 7). Results were observed to be unchanged more than 3577393 elements, but this could increase the computational cost. Therefore, the grid number of the conference room was set to 3577393.

Figure 6.

Sample of generated mesh: (a) bottom view of the mesh and (b) section view of the mesh.

Figure 7.

Grid independence test: (a) temperature distribution comparison between four meshes and (b) position of the selected axis and mesh sections.

Evaluation criteria

Healthy and comfortable environment is an essential factor for personnel productivity.⁴⁴ To compare different control methods, this study evaluated the airflow organization using the following indicators:

1. Temperature distribution;

2. Velocity distribution;

3. Effective Draught Temperature (EDT);

4. Air Diffusion Performance Index (ADPI);

5. Ventilation efficiency.

The velocity and temperature distribution were obtained directly from the measured (or simulated) data at each point. The $E D T$ , as a parameter reflecting the room uniformity, is expressed as shown in equation (11):⁴⁵

\begin{array}{c} E D T = (T_{x} - T_{a}) - 8 (v_{x} - v_{a}) \end{array}

(11)

where

T_{x}

and

v_{x}

are the temperature and velocity of the corresponding point, respectively.

T_{a}

and

v_{a}

represent the average temperature and average velocity of the room, respectively. A smaller

E D T

value means a more uniform indoor environment. In general, −1.7 < EDT < 1.1 is considered acceptable. The

A D P I

was calculated from the

E D T a s d e f i n e d b y E q u a t i o n

(12):⁴⁶

\begin{array}{c} A D P I = \frac{n}{N} \times 100 % \end{array}

(12)

where

n

represents the number of acceptable measurement points for

E D T

and

N

represents the total number of measurement points.

Ventilation efficiency $(ε_{c})$ measures the efficiency of an air conditioning system in removing foul air from a room. For cooling conditions, this parameter can be expressed directly in terms of temperature as defined by equation (13):⁴⁷

\begin{array}{c} ε_{c} = \frac{T_{e} - T_{S}}{T_{a} - T_{S}} \end{array}

(13)

where

T_{e}

is the exhaust air temperature and

T_{S}

is the supply air temperature. The larger the value of

ε_{c}

, the more efficient the air conditioner is in removing indoor air.

Results

The results are organized as follows: (a) Validation of the numerical model in this study, (b) experimental results of airflow characteristics, including the determination of $K$ and $x_{0}$ in equation (7), (c) the proposal and feasibility verification of CMT, and (d) application of CMT in two typical scenarios.

Validation

The validation model used in this study is the office design scenario in Table 1. Figure 8(a) shows the experimental and simulated values of jet velocity at L1 position. The two are in good agreement in most positions, while the simulated value is relatively low in a position close to the diffuser (50% lower than the experimental value). This may be due to the inaccurate modelling of the experimental boundary condition at the simulation inlet.⁴⁸ Since the area near the diffuser was not the focus of this study, this deviation was ignored. Figure 8(b) shows the temperature distribution of L2 and L3, which indicates that the experiment agrees well with the simulation. The simulation method (turbulence model, mesh density and tuyere modelling) described in this paper can be considered reasonable.

Figure 8.

Comparison of simulated and experimental values: (a) velocity distribution of the horizontal line L1 at different air volumes (all sub-tuyeres open); (b) temperature distribution of vertical lines L2 and L3 and (c) the position of the selected lines.

Jet characteristics

The experiments were carried out under isothermal conditions without arranging heat sources in the room. The design air volume was 328 m³/h. A single diffuser was tested in the thermal chamber by controlling the damper.

(1) Axis position

The experiment was conducted at the measurement points shown in Figure 4(a) to find the axis of the jet. The order of velocity at different heights is $v_{0.01 m}$ = $v_{0.02 m}$ > $v_{0.05 m}$ > $v_{0.08 m}$ (see Figure 9). Considering that the attached plane jet can be affected by ceiling type (flat ceiling or grille ceiling) and the actual working condition of the diffuser installation (whether there are local components on the ceiling) when it is closer to the ceiling,⁴⁹ this study selected 0.02 m from the ceiling as the axial height of the diffuser jet.

Figure 9.

Velocity attenuation curves at different distances from the ceiling.

The velocity distribution at point 1 was different from others. This was because the position was in the free jet zone, there was a bend in the jet axis and a low-pressure air area. Velocity decayed regularly after point 2. The airflow was assumed to enter the adherent jet zone. From equation (7), the velocity decay coefficient $K$ is only related to the adherent jet zone, and the range of the free jet and transition zone are uniformly attributed to the constant $x_{0}$ . In order not to affect the overall data fitting accuracy, point 1 was removed when calculating the velocity decay curve.

(2) Velocity attenuation

To consider the diffusion ability of the diffuser under different opening conditions, the velocity attenuation tests were conducted for all possible states involved in the diffuser operation. Each sub-tuyere corresponds to two states of ‘open’ and ‘closed’. There were 15 operating states of the diffuser, as shown in Figure 1, without considering the case where all sub-tuyeres were closed. These states included cases where the sub-tuyeres were fully open (ABCD), three sub-tuyeres were open (BCD, ACD, ABD, ABD, ABC), two sub-vents were open (AC, AB, AD, BC, BD, CD) and a single sub-tuyere was open independently (A, B, C, D). The dimensionless velocity $u_{x} / u_{0}$ and dimensionless distance $x / \sqrt F_{0}$ were introduced and fitted to the test data. The results are shown in Figure 10. The velocity decay coefficient $K$ varied in the interval between the maximum value $K_{A D} = 4.093$ and the minimum value $K_{D}$ = 1.485, and the coefficient $x_{0}$ ranged from 0.144 to 0.541.

Figure 10.

Velocity curves of different opening states of sub-tuyeres.

Figure 11 summarizes the velocity attenuation coefficient $K$ and transient diffusion coefficient $T D C$ under different opening conditions. Overall, $T D C$ increases with the decrease of the opening area of the diffuser, which proves that the reduction of air supply area can improve the diffusion ability. The state of opening D sub-tuyere alone is the smallest but $T D C$ is significantly lower than expected, mainly because the D outlet is at the inner edge and the geometric shape is not conducive to airflow diffusion ( $K_{D}$ = 1.48). The diffuser should avoid running under this condition. Equation (8) and equation (9) show that the increase $T D C$ can ensure that the airflow range could reach the set position when the air volume is reduced. The operation strategy of the tuyere should ensure that $T D C$ is increased with the decrease of air volume.

Figure 11.

Velocity attenuation coefficient $K$ and transient diffusion coefficient $T D C$ under different sub-tuyeres opening states.

Proposal of CMT

The purpose of this section is to propose CMT and verify its feasibility, including the comparison of airflow jet throw and velocity contour. The CFD simulation process in this section does not involve heat transfer and there is no indoor heat source arrangement. The modelling is similar to Figure 2(c)-S2 and is not described in detail here. The air inlet and outlet were set as velocity inlet and pressure outlet, respectively. The turbulence density was calculated using equation (10).

(1) Jet throw evaluation

The diffuser states selection using CMV is included in Table 4. The discharge velocity was kept at about 4.2 m/s. The experimental test results show that when the airflow of the VAV system was reduced, keeping the air supply velocity constant cannot maintain the jet throw at the set position. This phenomenon was more evident in low airflow conditions, where the throw lengths of 60% and 40% airflow were decreased to 76% and 57% of the original values, respectively. When the airflow was reduced to 20%, the throw length was already reduced to half of the design condition.

Table 4.

The opening sequence of sub-tuyere based on CMV.

Air volume conditions, %	Air volume (m³/h)	Opensub-tuyere	Air supply areas ( $m^{2}$ )	Air supply velocity (m/s)	Jet throw length (m)
100	328	ABCD	0.0216	4.218	3.201
80	262	ABD	0.0171	4.262	2.900
60	197	AC	0.0126	4.339	2.436
40	131	BD	0.0090	4.049	1.834
20	66	C	0.0045	4.049	1.615

Equation (8) was used to calculate the range of throw length variation for each sub-tuyere opening condition. Figure 12 shows each working condition was divided into four ranges according to five air volumes. The red line represents the position of the rated distance 3.2 m. The points on the left/right side of it represent conditions that the jet throw length is less/greater than the set value.

Figure 12.

Variation of jet throw length with air volume for different sub-tuyere combinations.

This paper defined the state points within 10% error range of the set value to meet the design requirements. The area of the sub-tuyere to be opened for different air volumes was selected as shown in Figure 12. When the air supply volume fell to 80%, the combination of ABD or AC sub-tuyere could ensure that the airflow range was maintained at about 3.2 m. When the air supply volume continued to fall to 60%, the combination of BCD, A and D sub-tuyere produced an ideal airflow improvement effect. When the air supply volume was 40%, opening B or C sub-tuyere alone could meet the requirements. For 20% airflow conditions, the diffuser could only open the C sub-tuyere to raise the range to 1.615 m at maximum, but could not reach the design value.

Table 5 summarizes the opening status of the sub-tuyere determined by CMT. Since the

T D C

of the diffuser was increased when the air volume was reduced, all these methods can theoretically meet the requirements of jet throw enhancement. However, higher air supply velocity would imply a greater pressure decline and noise, and excessive air supply velocity should be avoided in the selection. In this study, the sub-tuyere control method of ABCD-ABD-BCD-B-C was chosen.

Table 5.

The opening sequence of sub-tuyere based on CMT.

Air volume conditions	Air volume (m³/h)	Open sub-tuyere	K	TDC	Jet throw (m)
100%	328	ABCD	2.89	19.66	3.201
80%	262	ABD	2.93	22.41	2.900
80%	262	AC	2.86	25.48	3.365
60%	197	BCD	3.53	30.38	2.922
		A	2.62	29.11	2.946
		D	1.48	28.48	2.970
40%	131	B	3.40	42.84	2.855
		C	3.46	51.58	3.494
20%	66	C	3.46	51.58	1.615

The results in Table 4 and Table 5 show that the increase in momentum caused by the reduction in air supply area illustrated that both methods could produce an improvement effect on the jet throw. The difference between the two was reflected in the 60% and 40% airflow operating conditions, where the CMT improved the throw length by 20% and 56%, respectively. Opening AC at 60% airflow produced a higher supply air velocity than BCD ( $u_{A C} = 4.339 m / s$ , $u_{B C D} = 4.053 m / s$ ), but the diffuser produced a better diffusion performance under BCD condition ( ${T D C}_{A C} = 25.48$ , ${T D C}_{B C D} = 30.38$ ), which shows the defect of taking velocity as a single factor. In addition, the superiority shown by CMT at 40% volume is also due to the method that selected the opening state of the diffuser with a higher $T D C$ ( ${T D C}_{B D} = 28.88$ , ${T D C}_{B} = 42.84$ ).

(2) Velocity contour evaluation

Figure 13 shows the velocity contour of different air volume. The airflow cover area was decreased significantly with the decline of air supply volume. When the air volume was reduced to 20%, the indoor air velocity was basically below 0.5 m/s. This means that fresh air could not reach most areas. In the absence of adjustment measures, indoor air quality and comfort under low airflow would be difficult to ensure.

Figure 13.

Velocity contours (plane 0.02 m from the ceiling) for different air volume ratio: (a) 100%; (b) 80%; (c) 60%; (d) 40% and (e) 20%.

Numerical simulations were performed under the conditions shown in Table 4 and Table 5. Results in Figure 14 show that the throw length and diffusion width were improved by using CMT under 80% and 60% airflow conditions. This is mainly because the CMV does not consider the change in diffusion capacity and air supply area, and the actual working condition deviated from the design value. Furthermore, the comparison results of Figures 13 and 14 show that the new diffuser proposed in this paper has a significant enhancement effect on the airflow coverage area and can effectively improve airflow distribution at low airflow in the VAV system.

Figure 14.

Velocity contours (plane 0.02 m from the ceiling) under different air volumes: (a) 60% volume (CMV left and CMT right) and (b) 40% volume (CMV left and CMT right).

Application of CMT in improving indoor airflow distribution

(1) Air temperature distribution

The indoor temperature distribution for different operating conditions is shown in Figure 15. Generally, the average room temperature was maintained at about 26°C for different conditions. For the partial working condition, the room temperature of the office showed a significant inhomogeneity, with a difference of 6°C between the maximum temperature of 30°C and the minimum temperature of 24°C. This also showed a maximum temperature difference of 5°C in the conference room. The average room temperature in the office with CMV operating conditions was 1°C higher than the set value, which means that the heat in the room was not replaced effectively. Overall, CMT was able to significantly improve the room temperature uniformity while maintaining set points.

Figure 15.

Indoor temperature distribution of two types of rooms: (a) office and (b) conference room.

(2) Effective draught temperature (EDT)

The environmental evaluation index, effective ventilation temperature (EDT), is shown in Figure 16. Considering that −1.7 < EDT < 1.1 is acceptable, the ADPI for different working conditions was calculated by equation (12) (as shown in Table 6). The results show that CMT can create a more satisfactory environment, which is reflected in both the office and conference room. For office conditions, compared to Partial and CMV working conditions 81.36% and 89.66%, CMT can achieve ADPI of 95.69%. In the conference room, ADPI was raised from 78.76% and 92.11% to 97.37%.

Figure 16.

Effective draught temperature of two types of rooms: (a) office and (b) conference room.

Table 6.

ADPI for different conditions.

Case	Design, %	Partial, %	CMV, %	CMT, %
Office	90.76	81.36	89.66	95.69
Conference room	88.89	78.76	92.11	97.37

(3) Air velocity distribution

As shown in Figure 17, the velocity distribution characteristics of the ceiling diffuser under different working conditions are similar. The velocity is concentrated in the low velocity section of 0∼0.1 m/s. This is related to the ceiling diffuser airflow pattern, where the airflow reaches the occupant zone after adhering along the ceiling and walls, where the velocity decays to approximately 0 m/s.⁵⁰ Although CMT increases overall indoor velocity, this level is still acceptable (<0.1 m/s).⁵¹ It is worth noting that for the air supply form (such as the corridor-wall grille) where the airflow directly reaches the work area, the draught sensation will be an important factor affecting the comfort.³⁷

Figure 17.

Indoor velocity distribution of two types of rooms: (a) office and (b) conference room.

(4) Ventilation efficiency

Figure 18 shows that CMT has a lower ventilation efficiency compared to the other three conditions. The ventilation efficiency of the latter was around 1.0, while the CMT fell to near 0.9. These results are reflected in both office ( $ε_{c}$ = 0.88) and conference room ( $ε_{c}$ = 0.91) conditions. In this study, the return air outlet was set on the side wall, part of the cold airflow descended along the wall and then directly entered the return air outlet, which was detrimental to the elimination of residual heat in the room. This effect was more significant for CMT due to the extended airflow range, which enhanced the airflow of the wall zone. This was verified in the return air temperature, which was about 1°C lower for CMT than other conditions. The optimization of the return air outlet arrangement should be considered in the future.

Figure 18.

Ventilation efficiency of two types of rooms: (a) office and (b) conference.

Discussion

(1) Analysis of airflow pattern

To determine the correlation between airflow distribution and indoor comfort, the velocity vectors and temperature contours under four typical scenarios in the conference were compared as shown in Figures 19 and 20.

Figure 19.

Velocity vector charts under four typical scenarios: (a) 100% air volume (without regulation); (b) 40% air volume (without regulation); (c) 40% air volume (CMV) and (d) 40% air volume (CMT).

Figure 20.

Temperature contour plots of the conference room: (a) 100% air volume (without regulation); (b) 40% air volume (without regulation); (c) 40% air volume (CMV) and (d) 40% air volume (CMT).

The airflow was discharged from the diffuser, forming a spreading zone near the ceiling. Then it turned vertically downward along the walls, and becomes horizontal again when it reached the ground. Earlier studies revealed that the jet velocity decay before and after the corner is correlated.²⁹ This means that the airflow in these three regions can be treated as a whole. Figure 19 indicates the positions where the jet velocity decayed to 0.5 m/s and 0.25 m/s, which are commonly used to quantify the range of the airflow.²⁴ When the diffuser was not used to make any adjustment to the reduced air volume, the jet almost decayed near the ceiling. When the airflow reached the wall, the velocity declined to 0.06 m/s (Figure 19(b)). CMT was shown to have the longest throw length under the same air supply volume. At the same location on the wall, the velocity of CMT was 60% higher than that of CMV. This method, therefore, would enhance the jet strength at each positions.

Figure 20 shows the strengthening of the air flow, which removed the heat in the room more effectively. Due to the heat dissipation of occupants and computers, the high temperature area in the room was located in the middle. Compared with the control group (Figure 20(b)), the use of CMV allowed the cold air to be delivered farther and the uniformity of the indoor temperature distribution was improved. However, there were still large areas near the personnel with temperatures reaching 30°C. CMT could improve the problem of excessive local temperature, which would meet the personnel thermal comfort requirements. Figure 20(a) is not discussed here because of its different heat source distribution.

The simulation results indicate that the airflow pattern of ceiling diffusers can be divided into two categories: full coverage model (Figure 21(a)) and partial coverage model (Figure 21(b)). The former represents the typical pattern of the ceiling diffuser.²⁹ The airflow entered the working area at a certain velocity after impinging and turning at the corner. The thermal plume generated by the indoor heat source would move upwards and entrain the airflow in the spreading zone.⁵² This loop can remove heat from the working area.

Figure 21.

Schematic diagrams of flow pattern on the middle section in the case of two different air volumes: (a) full coverage model and (b) partial coverage model.

The airflow pattern would gradually change to partial coverage model as the initial air supply momentum was decreased. The jet was completely mixed with room air before it reached the walls, and the airflow in the wall and the ground zone would be weaken. When the heat source in the room was concentrated and of high intensity, the heat dissipation in the corresponding area would not be effectively removed. Earlier studies have shown that excessive wind velocity near the ground may cause draught problems in rooms.⁵³ This study supplements that conclusion and shows the increase in the initial air supply momentum is necessary in order to improve the heat removal efficiency.

Compared with CMV, CMT would extend the distance of the spreading zone by increasing the momentum of the outlet airflow. The airflow range of the spread zone would be more easily assured and the airflow intensity of the wall zone and ground zone would be enhanced. Thus, it can maintain optimum heat removal and environment creation in the VAV system.

(2) Throw length of non-isothermal jet

The throw length distribution diagram (Figure 12) was tested under isothermal conditions. Air supply temperature difference was not considered. Influenced by inertial force and buoyancy, the velocity decay and behaviour of non-isothermal jets vary across different $A r$ (Archimedes number) ranges.³² ASHRAE-Handbook provides several typical indoor loads when recommending throw length to consider indoor heat source conditions.²⁴ The ADPI selection guide is different in cooling and heating operation modes.^13,54 In the future, in order to improve the applicability of CMT, the influence of supply air temperature difference on throw length needs to be corrected.

Conclusion

In this study, an active diffuser control method based on constant jet throw length has been proposed, which aims to replace the original control method of keeping the outlet air velocity constant. The main findings are summarized as follows:

(1) A ceiling diffuser based on sub-tuyere was tested. Each sub-tuyere area of the diffuser can be controlled by a baffle, so that the air supply area can be varied at different air volumes. The experimental results show that the new diffuser can adapt to a range of airflow variations from 40% to 100%, and the throw length can be increased to 50% of the design value under the condition of 20% air volume.

(2) Although CMV can have an improving effect on the indoor environment under partial load. It can still make the throw length deviate from expectations. The diffusion performance of the diffuser is dynamically changing during the adjustment process. The experimental results show that the actual throw length deviated to 76% and 57% of the original at 60% and 40% airflow conditions, respectively.

(3) The new control method CMT was proposed and applied to typical office and conference rooms. Results show that CMT can create a more uniform indoor temperature environment compared to CMV. This method can achieve ADPI values of 95.69% and 97.37% in offices and conference rooms, respectively.

This paper recommends that the active diffuser should be tested by the manufacturer for different opening conditions before leaving the factory, and provides typical reference airflow range distribution diagrams for designers. In the future, different kinds of tuyeres should be studied to verify the universality of CMT. Furthermore, it is time-consuming and labour-intensive to test the airflow of all the open states of the active tuyere. How to simplify this method to better apply it to actual production is also the focus of research.

Footnotes

Acknowledgements

Wei Xu is gratefully thanked for his assistance in the experiment. Thanks to Qian Wang, Cheng Guo and Qili Shi for their valuable comments on this paper.

Author Contributions

Linye Song: investigation, data analysis, methodology and writing – original draft preparation; Kaijun Li: methodology, writing – review and editing; Xinghui Zhang: supervision, project administration, writing – review and editing; and Jing Hua: writing – review and editing.

Declaration of conflicting interests

The author(s) declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article.

Funding

The author(s) received no financial support for the research, authorship, and/or publication of this article.

ORCID iDs

Linye Song

Kaijun Li

References

Kialashaki

. Energy and economic analysis of model-based air dampers strategies on a VAV system. International Journal of Environmental Science and Technology 2018; 16(8): 4687–4696.

Peng

Zou

Bai

Feng

. Building energy consumption prediction and energy control of large-scale shopping malls based on a noncentralized self-adaptive energy management control system. Energy Exploration & Exploitation 2020; 39(5): 1381–1393.

Song

Zhang

Hua

Zhang

. Differentiated Control of Large Spatial Environments: Air Curtain Grid System. Sustainability 2023; 15(6): 5489.

Okochi

Yao

. A review of recent developments and technological advancements of variable-air-volume (VAV) air-conditioning systems. Renewable and Sustainable Energy Reviews 2016; 59: 784–817.

Warsinger

. Energy savings of retrofitting residential buildings with variable air volume systems across different climates. Journal of Building Engineering 2020; 30: 101223.

Wang

Zhang

Brambley

Futrell

. Performance simulation and analysis of occupancy-based control for office buildings with variable-air-volume systems. Energies 2020; 13(15): 3756.

Nassif

Tahmasebi

Ridwana

Ebrahimi

. New optimal supply air temperature and minimum zone air flow resetting strategies for VAV Systems. Buildings 2022; 12(3): 348.

Kang

Kim

Cho

. A study on the control method of single duct VAV terminal unit through the determination of proper minimum air flow. Energ Build 2014; 69: 464–472.

Ligade

Razban

. Investigation of Energy Efficient Retrofit HVAC Systems for a University: Case Study. Sustainability 2019; 11(20): 5593.

10.

Hurnik

. Novel cylindrical induction controller and its application in VAV air conditioning system in an office building. Energ Build 2016; 130: 341–349.

11.

Ringo Ng

Liao

. Evaluating the need and potential of equipping north american houses with multi-zone VAV Systems. Advanced Materials Research 2011; 361-363: 22–30.

12.

Chenari

Dias Carrilho

Gameiro da Silva

. Towards sustainable, energy-efficient and healthy ventilation strategies in buildings: A review. Renewable and Sustainable Energy Reviews 2016; 59: 1426–1447.

13.

Kim

Lim

Moon

Park

Rhee

. Heating Performances of a Large-Scale Factory Evaluated through Thermal Comfort and Building Energy Consumption. Energies 2021; 14: 5617.

14.

Lin

Liu

Zhang

. Systematic summary and analysis of Chinese HVAC guidelines coping with COVID-19. Indoor Built Environ 2022; 31(5): 1176–1192.

15.

Liu

. Requirements for outdoor air in hospital wards in China. Science and Technology for the Built Environment 2018; 24(10): 1150–1155.

16.

Che

Tso

Sun

DYK

Lee

Chao

CYH

. Energy consumption, indoor thermal comfort and air quality in a commercial office with retrofitted heat, ventilation and air conditioning (HVAC) system. Energ Build 2019; 201: 202–215.

17.

CKH

Chan

Lai

ACK

. Influence of mechanical ventilation system on indoor carbon dioxide and particulate matter concentration. Build Environ 2014; 76: 73–80.

18.

Mak

Cui

Xue

. Ventilation of air-conditioned residential buildings: A case study in Hong Kong. Energ Build 2016; 127: 116–127.

19.

Zhu

Lei

Wang

Zhang

. Air quality and passenger comfort in an air-conditioned bus micro-environment. Environ Monit Assess 2018; 190(5): 276.

20.

Mustakallio

Kosonen

Kaukola

Chen

Liu

. Experimental study of five different VAV air terminal devices under variable heat gain conditions in simulated office and meeting rooms. Build Environ 2022; 209: 108641.

21.

Szczepanik-Scislo

Schnotale

. An Air Terminal Device with a Changing Geometry to Improve Indoor Air Quality for VAV Ventilation Systems. Energies 2020; 13(18): 4947.

22.

Rabani

Madessa

Nord

Schild

Mysen

. Performance assessment of all-air heating in an office cubicle equipped with an active supply diffuser in a cold climate. Build Environ 2019; 156: 123–136.

23.

Miller

. Room air distribution performance of four selected outlets. ASHRAE Transactions 1971; 77(2): 194.

24.

ASHRAE . ASHRAE Handbook — Fundamentals. Atlanta GA: American Society of Heating, Refrigerating and Air Conditioning Engineers, 1997.

25.

Xiao

Zhai

. Mathematical model and analysis of semi-restricted dispersed jet from a diffuser. Heating Ventilating & Air Conditioning 2007; 04: 12–14.

26.

Sun

Smith

. Air flow characteristics of a room with square cone diffusers. Build Environ 2005; 40(5): 589–600.

27.

. Study on isothermal jet characteristics of square ceiling diffuser. Master thesis. Shanghai, China: University of Shanghai for Science and Technology, 2011. (in Chinese).

28.

O’Donovan

Murray

. Fluctuating fluid flow and heat transfer of an obliquely impinging air jet. International Journal of Heat and Mass Transfer 2008; 51(25–26): 6169–6179.

29.

Cao

Ruponen

Paavilainen

Kurnitski

. Modelling and simulation of the near-wall velocity of a turbulent ceiling attached plane jet after its impingement with the corner. Build Environ 2011; 46(2): 489–500.

30.

Long

Simulation and experimental study of the cold air attached jet. Master thesis, Xi’an, China: Xi’an University of Architecture and Technology, 2007. (in Chinese).

31.

. Practical heating air conditioning design manual. Beijing: China Building Industry Press, 2007.

32.

Liao

Liang

Chiang

. Scale model study of airflow performance in a ceiling slot-ventilated enclosure: Non-isothermal condition. Build Environ 2007; 42(3): 1142–1150.

33.

Rajaratnam

. Turbulent jets. Amsterdam: Elsevier Scientific Publishing Company, 1976.

34.

Borowski

Jaszczur

Satoła

Karch

. Air flow characteristics of a room with air vortex diffuser. In: 11th International Conference on Computational Heat - ICCHMT, Cracow, Poland, 21-24 May 2018, pp240.

35.

Lyu

Feng

Cheng

Liao

. Experimental and numerical analysis of air temperature uniformity in occupied zone under stratum ventilation for heating mode. Journal of Building Engineering 2021; 43: 103016.

36.

ANSI/ASHRAE standard 55-2020: thermal environmental conditions for human occupancy. ASHRAE, Atlanta (USA), 2020.

37.

Zhao

Lestinen

Mustakallio

Kilpeläinen

Jokisalo

Kosonen

. Experimental study on thermal environment in a simulated classroom with different air distribution methods. Journal of Building Engineering 2021; 43: 103025.

38.

Yau

Poh

Badarudin

. A numerical airflow pattern study of a floor swirl diffuser for UFAD system. Energ Build 2018; 158: 525–535.

39.

Wei

Chen

Lai

Chen

. An improved displacement ventilation system for a machining plant. Atmospheric Environment 2020; 228: 117419.

40.

Zhao

Guan

Ren

. Air supply opening model of ceiling diffusers for numerical simulation of indoor air distribution under actual connected conditions, Part II: Application of the Model. Numerical Heat Transfer, Part A: Applications 2006; 49(8): 821–830.

41.

Song

Zhang

Wang

Hua

. Study of influence of boundary condition of diffuser with non-uniform velocity on the jet characteristics and indoor flow field. Energies 2023; 16(3): 1079.

42.

Cao

Kandzia

Müller

Heikkinen

Kosonen

Ruponen

. Experimental study of the effect of turbulence intensities on the maximum velocity decay of an attached plane jet. Energ Build 2013; 65: 127–136.

43.

Chen

Zhou

Feng

Cao

. Application of polyhedral meshing strategy in indoor environment simulation: Model accuracy and computing time. Indoor Built Environ 2021; 31(3): 719–731.

44.

Brager

Zhang

Arens

. Evolving opportunities for providing thermal comfort. Building Research & Information 2015; 43: 274–287.

45.

Cheng

Fong

Yao

Lin

Fong

. Uniformity of stratum-ventilated thermal environment and thermal sensation. Indoor Air 2014; 24 (5): 521–532.

46.

Miller

Nash

. A further analysis of room air distribution performance. ASHRAE Trans 1971; 77(2): 205.

47.

Yang

Zhou

Olofsson

. Influence of relayed fans and low level exhausts on performance of attachment ventilation under heating mode. Journal of Building Engineering 2021; 36: 102155.

48.

Draksler

Ničeno

Končar

Cizelj

. Large eddy simulation of multiple impinging jets in hexagonal configuration–Mean flow characteristics. International Journal of Heat and Fluid Flow 2014; 46: 147–157.

49.

Song

. Visualization of the airflow patterns of 12 typical diffusers and experimental investigation on the turbulent coefficient. Master thesis. Xi’an, China: Xian University of Architecture and Technology, 2005. (in Chinese).

50.

Mohammed

. A simplified method for modeling of round and square ceiling diffusers. Energ Build 2013; 64: 473–482.

51.

Khodakarami

Nasrollahi

. Thermal comfort in hospitals – A literature review. Renewable and Sustainable Energy Reviews 2012; 16(6): 4071–4077.

52.

Liu

Zhao

Liu

Luo

. Numerical investigation of the unsteady flow characteristics of human body thermal plume. Build Simul 2016; 9(6): 677–687.

53.

Cao

Kurnitski

Ruponen

Seppänen

. Experimental investigation and modelling of a buoyant attached plane jet in a room. Applied Thermal Engineering 2009; 29(14–15): 2790–2798.

54.

Liu

Clark

Novoselac

. Air diffusion performance index (ADPI) of overhead-air-distribution at low cooling loads. Energ Build 2017; 134: 271–284.

An active diffuser control method based on the constant jet throw for improving airflow distribution in variable air volume systems

Abstract

Keywords

Introduction

Methodology

Thermal comfort and jet throw

Studied cases

Experimentation

(1) Jet characteristics tests

(2) Indoor environmental parameters tests

Computational fluid dynamics (CFD) simulation

Evaluation criteria

Results

Validation

Jet characteristics

(1) Axis position

(2) Velocity attenuation

Proposal of CMT

(1) Jet throw evaluation

(2) Velocity contour evaluation

Application of CMT in improving indoor airflow distribution

(1) Air temperature distribution

(2) Effective draught temperature (EDT)

(3) Air velocity distribution

(4) Ventilation efficiency

Discussion

(1) Analysis of airflow pattern

(2) Throw length of non-isothermal jet

Conclusion

Footnotes

Acknowledgements

Author Contributions

Declaration of conflicting interests

Funding

ORCID iDs

References