Effect of Air Supply Temperature on the Performance ... - SAGE Journals

Original Paper

Environment

Indoor Built Environ 2005 14;2:103–115

Accepted: November 26, 2004

Effect of Air Supply Temperature on the Performance of Displacement Ventilation (Part I) – Thermal Comfort Zhang Lin

T.T. Chow

C.F. Tsang

L.S. Chan

K.F. Fong

Division of Building Science and Technology, City University of Hong Kong

Key Words Displacement ventilation · Thermal comfort · Air velocity · Percentage of dissatisfied people (PD) · Predicted percentage of dissatisfied people (PPD) · Air supply temperature

Abstract The effect of the air supply temperature on the performance of a displacement ventilation (DV) system has been investigated. The study looks into the case of a typical office building in Hong Kong, under local thermal environment and airflow conditions. These are characterised by the high cooling load density, which commonly occurs in Hong Kong. The lower supply temperatures were found to result in higher draft effects, yet increasing temperatures leads to an increase in the predicted percentage of dissatisfied people (PPD). The DV system was found to provide acceptable thermal comfort at certain air supply temperatures. The levels of PPD could be marginal should the parameters of the system not be optimised.

© 2005 Sage Publications DOI: 10.1177/1420326X05052563 Accessible online at: www.sagepublications.com

Introduction In recent years there has been a growing client aspiration regarding the quality of the working and living environments including factors such as better thermal comfort and indoor air quality (IAQ). This has placed demands on achieving better performance with existing conventional heating, ventilating and air-conditioning (HVAC) systems. Many employers and government organisations have also become more active in promoting better indoor environments in response to public demands. In Hong Kong the situation is further complicated by the recent large influx of heat generating devices characterised by uneven distributed loads in the occupied zone. Additionally, since the last decade there has been an explosion in the amount of cables and wires in many of today’s buildings [1], which also contributes to the everincreasing cooling load. All these factors indicate that there is a requirement for a HVAC system that is able to provide better thermal comfort and satisfy these new demands. Over the past decades a new type of ventilation system that is an alternative to conventional mixing ventilation systems has gained increasing popularity. The

Dr Zhang Lin Division of Building Science and Technology City University of Hong Kong, Hong Kong S.A.R. Tel. 852 2788 9805. Fax 852 2788 9716. E-Mail [email protected].

DV system was first proposed and implemented in the Scandinavian countries about 25 years ago [2]. The system was first envisaged for use in heavy industrial environments where a high level of cleanliness was required, in addition to cost effectiveness. However its application soon spread to other areas including offices and commercial buildings. The use of the floor supply DV system has gained popularity in other countries as well, including the United States, Germany and Japan. An investigation was conducted in the United States into the feasible application of the DV system [3,4,5]. A set of guidelines for designing DV systems was developed as a result [6]. In the case of Hong Kong, direct application of these guidelines is not feasible because of the relatively high cooling load density. So for there have been only a few recorded cases of the use of DV systems in Hong Kong. The problem of drafts affecting thermal comfort has been associated with the DV system for a long time. The supply of fresh air directly into the occupied zone and consequent thermal stratification increases the likelihood of drafts. Shute and Bauman stated that an air supply temperature of at least 17°C was required to avoid draft effects [7,8]. Matsunawa et al. indicated that occupants should be located at least 1 to 1.5 m away from the supply grilles [9]. One of the benefits of the floor supply system is that areas of stagnant air can be removed from an office with partitioned furniture [10]. McCarry found that supply diffuser locations at the ceiling level may lead to poor air circulation [11]. Several researchers have looked into the problem of air supply temperatures. Gan used computational fluid dynamics (CFD) to predict thermal comfort parameters within an office environment [12]. He reported that the cold thermal sensation and draft risk decreases with an increase in supply air temperatures. His investigations showed that thermal discomfort can be avoided by optimising the air supply velocity and temperature. The effects of low air supply temperatures and high relative humidity was studied by Trent, who showed that this can have direct effects on IAQ and can cause equipment damage to exposed hardware components [13]. The influence of the air supply temperatures on the thermal comfort of an office was also investigated by Chen et al. [14]. A CFD model was developed to study the influence of the diffuser supply parameters such as airflow velocity, air supply temperature and dimensions of the diffuser. They found that the supply flow rate and temperature are very significant factors in determining the level of thermal comfort within an office.

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Indoor Built Environ 2005;14;2:103–115

A series of experiments was carried out by Lian to determine the effect of under-floor supply air-conditioning systems on thermal comfort [15]. They found that the most important factor for thermal comfort was the distance between the occupant and the outlet. The air supply temperature and velocity were found to have moderate influence while the type of outlet had negligible influence. The present work focuses on the situation in Hong Kong where there is a high cooling load density due to the local subtropical climate and high density of electrical appliances. The air supply temperature is obviously a very important factor affecting the DV system performance. This suggests that a further study based on local design conditions is justified to assess its effect on system design and performance.

The Effect of the Air Supply Temperature One important objective of an air distribution system is to create a pleasant thermal environment with the proper combination of comfort variables [16]. The comfort variables are metabolic rate, clothing, air velocity, air temperature, air temperature stratification, radiant temperature, radiant temperature asymmetry, relative humidity and turbulence intensity in the occupied zone [17,18]. The aim of this paper is to examine the air supply temperature and assess its effect on thermal comfort. In the ASHRAE comfort standard 55-1992, air velocity is an important factor in determining the comfort level. However, the standard also restricts the extent to which air velocity can be used to achieve comfort, by limiting it to a specified maximum of 0.8 m·s1 in summer [19,20]. Most research into the velocity vs. temperature tradeoff has assumed that the temperatures of the ambient air and the air blowing over the skin are the same. Huber surveyed a building in which floor-supply ventilation was a source of draft, causing discomfort [21]. The transition from desirable to annoying air movement needs to be studied further, particularly for cases where the temperature of the localised air differs from the ambient surroundings. At the design stage, the HVAC engineer needs to predict the system performance. The thermal comfort level is one of the most important aspects of that performance. There are a number of thermal comfort models currently available. The most common and probably the best-understood parameters are the predicted

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mean vote (PMV) for thermal comfort and the associated percentage of persons dissatisfied [19]. However, the influence of turbulence intensity is not included in the PMV nor in most other previous models. Fanger et al. [15] pointed out that the turbulence of the airflow has a significant impact on the sensation of draft, and so he developed a mathematical model of draft risk including turbulence intensity. The model is able to predict the percentage of people dissatisfied due to draft as a function of mean air velocity, turbulence intensity and air temperature. The air temperature distribution, the percentage of dissatisfied people (PD) due to draft, and the PPD with their thermal comfort are now widely used as criteria to evaluate thermal comfort. All these performance parameters are determined by the thermal and flow boundary conditions, such as the size and geometry of the space, rates and temperatures of airflows and heat sources. Evaluation Criteria This investigation applied a CFD program validated both for displacement ventilation and mixing ventilation through a comparison of computed and experimental data sets for an individual office, a cubicle office and a quarter of a classroom [22], which was based on a commercial CFD code [23]. This program was used to calculate the flow and thermal distribution for a large number of selected boundary conditions. There are many turbulence models available. The “standard” k- model is probably most widely used in building engineering calculations due to its relative simplicity [24]. However, the model sometimes provides poor results for indoor airflow. Many modifications have been applied to the standard model. Yet the modified models still do not have a general applicability for indoor airflow. Chen calculated the various indoor flows with 8 different turbulence models [25,26]. His study concluded that the Re-Normalization Group (RNG) k- model is the best among the eddy-viscosity models tested [27]. Since the RNG k- model is valid for high Reynolds number turbulent flow, wall functions are needed for the near wall region where the flow Reynolds number is low. Launder and Spalding’s wall functions were therefore applied [24] except that for temperature boundary. This is because the wall function would predict grid dependent heat flux and cause an unacceptable error. The heat transfer coefficient used is the one based on experimental data [28]. This is undesirable in a numerical prediction, because the coefficient is generally unknown. A new 1 equation model for the near wall flow was developed [29]. The heat transfer can be correctly calculated with the new model.

Air Supply Temperature and Thermal Comfort

Table 1. Fineness of grid Individual office

Cubicle office

Quarter of classroom

84 70 51 43 35 25 29 35 25

84 53 44 43 27 25 29 35 25

75 56 44 39 28 25 29 35 25

The governing equations are solved numerically. The whole computational domain, the space of the room, needs to be divided into a number of finite volumes by a grid system. The flow variables, such as velocity, temperature and concentration, are solved at the centre of each finite volume. The more grids used, the more accurate the results will be. However, a fine grid will cost more computing time and capacity. Table 1 shows the 3 different grid sizes. The difference between the results with 2 groups of finer grids is very small. Therefore, the finest grids were safely used for the comparison with the experimental data. However, normally CFD programs do not calculate PD and PPD. The following models were therefore implemented in the CFD program to calculate these parameters. Firstly the model developed by Fanger et al. and stipulated in ISO7730:1994(E) was applied [24,25]: PD (34 T)(u 0.05)0.62(3.14 0.37u Tu) [%]

(1)

(for u 0.05 m/s, use u 0.05 m/s; for PD 100%, use PD 100%) Where Tu 100(2k)0.5/u [%]

(2)

The formulae for calculating PPD can also be found in ISO7730:1994(E) [24]: PPD10095exp(0.03353PMV4 0.2179PMV2)[%] (3) The predicted mean vote, PMV, in the equation is determined by PMV [0.303 exp(0.036M) 0.028] L

(4)

with L M W {3.96 10 fcl [(Tcl 273)4 (Tr 273)4] fclhc (Tcl T) 3.05 103[5733 6.99(M W) Pa] 0.42 (M W 58.15) 1.7 10 5M(5867 Pa) 0.0014M(34 T)} (5) 8


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fcl, Tcl and hc in the above equations are determined by: fcl 1.05 0.645Icl

for Icl ≥ 0.078

(6)

fcl 1.00 1.290Icl

for Icl < 0.078

(7)

Tcl 35.7 0.028(M W) Icl{3.96 108fcl[(Tcl 273)4 (Tr 273)4] fclhc (Tcl T)} (8) The convective heat transfer coefficient, hc, is determined from: hc 2.38(Tcl~T)0.25

for 2.38(Tcl~T)0.25 ≥ 12.1u0.5 (9)

˙

˙

and hc 12.1u0.5

for 2.38(Tcl~T)0.25 < 12.1u0.5

(10)

˙

In addition, a CFD program was used to determine the airflow pattern and the distributions of air velocity and temperature. The thermal comfort provided by DV of 3 different supply air temperatures were compared in terms of

(i) (ii) (iii) (iv)

airflow pattern, temperature distribution, PD, and PPD.

The IAQ for DV is another important aspect for evaluating the performance of the 3 cases. Research work was also conducted with a focus on IAQ and the results are reported in a companion paper. Cases Studied In order to save energy, most of the air handling units in Hong Kong are not equipped with a re-heater. Therefore the indoor relative humidity requirement has to be relaxed. To date chilled ceiling panels have never been used in commercial projects in Hong Kong because of the high relative humidity at times. The present investigation considered DV systems without the incorporation of chilled ceiling panels for the case of a typical office building. The study focused on a cooling operation because of the subtropical climate of Hong Kong, where space cooling is in need in large commercial buildings all year round. Three cases of different supply air tempera-

Exhaust Supply

Fax machine

Occupant

8.4 m X Computer

Z 2.7 m

Y

8.0 m

Origin (0,0,0)

Exhaust Photocopier

Supply

Fig. 1. Typical office room.

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ture were investigated. The temperatures were 19°C, 20°C or 22°C respectively. Figure 1 illustrates the typical space configuration for all the cases. The thermal and flow conditions were determined based on the following typical design conditions: • • • • •

0.2 m·s1 in the unoccupied zone. Higher velocities however are observed in the occupied zone, near the

Headroom 2.7 m Qt/A 150 W/m2, 0.1 Qoe/Qt 0.4, 0.3 Ql/Qt 0.75, ˙ 0 Qex/Qt 0.9.

(a) x 4.5 m

˙

Comparison of Performance For the 3 different case studies, the airflow patterns and the distributions of air temperature, PD and PPD were determined. The results of the 3 case studies, 1 to 3 (for air supply temperatures 19°C, 20°C and 22°C, respectively), were compared using the above parameters. All the parameters relevant to cooling load were specified the same. The details of all the major parameters are summarised in Table 2. The ventilation system adopted supply diffusers installed at the floor level and exhaust outlets at the ceiling. Airflow Pattern The office space was taken as installed with 1 linear diffuser in each of the 2 small rooms and 2 rectangular diffusers in the large open office (Figure 1). The rectangular diffusers are located in the centre of the larger open office. In each smaller room the supply is located along the wall behind the occupants. The diffusers are so sized that the face velocity is 0.3 m·s1 for the small offices and 0.26 m·s1 for the open office. The supply air initially spreads upon the floor, and then rises as it is heated by various heat sources. The air rises near the external wall and the window as it is heated (Figure 2(b)). This effect is not observed on the other three internal partitions, as they are considered to be thermally symmetrical, and the adjacent room is considered to have the same temperature. The heat sources are the photocopiers, computers, fax machines and the human occupants. The air velocity is low of the order of less than

(a) x 4.5 m

(c) z 1.1 m Fig. 2. Airflow profile for case 1.

Table 2. Major parameters used in the simulations Number of people

West wall Window (W) (W)

PC (W)

10

900

70 10 1280

3168


Lamp (W)

Photocopier (W)

Fax machine (W)

Air circulation Fresh air intake (m3·s1) (L·s1 per person)

800

200

0.809

10


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photocopier and fax machine, and the external wall. These two areas are the locations of the highest heat sources. Velocities tend to be in the range of 0.3 to 0.5 m·s1. At the external wall location the airflow rises rapidly into the ceiling area. The greatest velocities are at the window location with velocities averaging between 0.4 m·s1 to 0.5 m·s1. A typical airflow pattern for 3 sections for case 1 is shown in Figure 2. The direction and magnitude of the flow field for the other 2 cases are very similar. This indicates that the air supply temperature variation of 1°C to 3°C

(a) Case 1

(c) Case 3

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does not significantly affect the velocity flow field when the difference in temperatures of the supply air and the room air is modest. The airflow pattern is to a large extent determined by the heat sources, with the highest velocities at the exterior wall – the largest heat source. Temperature Distribution Figure 3 shows a section of the office at the occupant breathing level, (h 1.1 m) above the floor level. The temperature is almost uniform in this horizontal plane, except at locations near heat sources. The temperature

(b) Case 2

Fig. 3. Temperature distribution at z 1.1 m (°C).

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(a) Case 1

(a) Case 1

(b) Case 2

(b) Case 2

(c) Case 3

(c) Case 3

Fig. 4. Temperature at x 4.5 m (°C).

Fig. 5. Temperature at y 5.6 m (°C).

increases slowly with height, forming vertical temperature gradients (Figures 4 and 5). This temperature gradient is dependent on the location and distribution of heat sources. The gradient in the lower part is larger than that in the upper part of the office, where most of the heat sources are located. Comparison between the case studies indicates that there is slight difference between the 3 results. The increase in air supply temperatures increases the mean temperatures in the room as shown in Figures 4 and 5. In the occupied zone, the temperature stratification of the air can be seen. The zones of higher temperatures are larger for the case with higher supply temperature. In the unoccupied zone with re-circulation, higher temperatures were also observed for the cases with higher supply temperature. The temperature gradients were almost the same for all 3 test cases in the occupied zone. In the main office area the gradient is 6°C between the ankle and head levels of the occupants. In the individual office the same

parameter is 3°C. The values in the main office area exceed the 3°C as recommended in ASHRAE 1992 [12] and ISO1984 [13]. The fact that the temperature gradients are similar, suggests that the room temperature gradients are determined by the heat sources. This was pointed out by Neilsen (1996), who showed through a series of experimental results that there is a strong correlation between vertical air temperature gradient and the surface temperature of heat sources.



PPD and PD The PPD shows great variations in values between the 3 study cases (Figures 6 to 8). For case 1, in the main office area the PPD values were between 8–9%. In case 2 values were between 10–13%. For case 3, values were between 30–35%. ISO7730 recommends a comfort limit of 10% or less for PPD. This suggests that higher temperatures are unsatisfactory for thermal comfort. In the occupied zone, the PD or draft rating (DR) was

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(a) Case 1

(c) Case 3

generally less than 10% for the typical office building (Figures 9 to 12). This is within the ISO comfort standard of 15% as suggested by ISO7730. The draft rating however was higher for lower supply air temperature, cases 1 and 2, although the difference was only slight. A low air supply temperature typically below 20°C indicates that many occupants would feel a draft, especially at the ankle level. Higher levels of thermal discomfort exist due to draft for the occupants seated close to the diffuser locations. This is clearly seen in Figure 10 which

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(b) Case 2

Fig. 6. PPD index at z 1.1 m (%).

shows the PD values at the ankle level. The side of the occupant closest to the diffusers was exposed to higher levels of PD due to the draft from the supply. This was also seen in the smaller individual offices. The rear of the occupant facing the supply diffuser and window was exposed to high levels of PD. In the unoccupied zone, PD values in general were also typically higher, especially at the window location and at the exhaust grilles.

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(a) Case 1

(a) Case 1

(b) Case 2

(b) Case 2

(c) Case 3

(c) Case 3

Fig. 7. PPD index at x 4.5 m (%).

Conclusions A model has been developed to determine the effects of varying the air supply temperature of a DV system. The model was based on a CFD model of a typical Hong Kong office. The database generated by this validated model has provided data that can be used to analyse the influence of air supply temperatures on the overall thermal comfort of an office building. This study showed the importance of the air supply temperature in determining the thermal comfort of office buildings. It can be concluded that the lowest supply tem-


Fig. 8. PPD index at y 5.6 m (%).

peratures (19°C and below) resulted in higher levels of draft rating dissatisfaction. Increasing temperatures produce increasing levels of the PPD. It can be seen that under Hong Kong operating conditions, air supply temperatures for the DV system should not exceed 19°C.

Acknowledgement The work described in this paper was supported by Competitive Earmarked Grant for Research No. CityU 1153/03E of Hong Kong.


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(a) Case 1

(b) Case 2

(c) Case 3

Fig. 9. PD index at z 1.1 m (%).


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(a) Case 1

(c) Case 3


(b) Case 2

Fig. 10. PD index at z 0.1 m (%).


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(a) Case 1

(a) Case 1

(b) Case 2

(b) Case 2

(c) Case 3

(c) Case 3

Fig. 11. PD index at x 4.5 m (%).

Nomenclature A ce cs c fcl hc Icl K M Pa Qt

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floor surface area (m2) contaminant concentration at the exhaust air (ppm) contaminant concentration at the supply air (ppm) contaminant concentration in the room air (ppm) clothing factor convective heat transfer coefficient between the cloth and air (W·m2 K) clothing insulation (°C·m2·W1) the turbulent kinetic energy metabolism (W·m2) partial water vapour pressure (Pa) total cooling load in the room (kW)


Fig. 12. PD index at y 5.6 m (%).

Qoe heat generated by occupant, desk lamps, and equipment (kW) Ql heat generated by lighting (kW) Qex heat from exterior walls and windows and transmitted solar radiation (kW) T local air temperature (°C) Tcl clothing temperature (°C) Tr mean radiant temperature (°C) Tu turbulent intensity u air velocity (m·s1) W external work (W·m2)

ventilation efficiency density (kg·m3) mean age of air (s).

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