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Dynamics of Semi Unidirectional Air Flows.

INTRODUCTION

Unidirectional air flow (often referred as "laminar" airflow) maintains the direction of the supply air without significant recirculation or mixing (often referred as "turbulent" airflows). Such unidirectional airflows are desirable in spaces such as in cleanrooms where a high level of cleanliness is desirable. In such cleanrooms, unidirectional flows are generally maintained by supplying air at higher velocities (90 fpm, 0.46 m/s or higher) through High-Efficiency Particulate Air (HEPA) filters covering the entire ceiling and extracting it through a raised perforated floor to create a "single pass" flow path. However, obstructions (e.g., people, furniture, and equipment) can disrupt the directionality of unidirectional airflows causing localized zones of non-unidirectional recirculating airflow patterns. Such spaces generally remain almost under isothermal conditions due to relatively high flow rates of the supply air. It should be noted that unidirectional flow can be obtained either when the air is supplied through the entire ceiling or when the restraining panels, such as in the case mini-environment cleanrooms, are placed surrounding the array of supply diffusers.

In other environments, such as in the hospital operating room, where only a certain region of the room (a "sterile" zone) needs to be maintained at high level of cleanliness, the air is supplied through an array of HEPA filters covering only part of the ceiling and extracted through low level return grills. Since the air is supplied at relatively low velocities (25 to 35 fpm, 0.13 to 0.18 m/s) and only through a part of the ceiling, the unidirectional flow can be maintained mostly in the core of the air jet; whereas, in the other regions of the room recirculating airflow patterns are formed, including at the periphery of the supply air jet (Khankari 2017a, Khankari 2017b). Such airflow patterns are referred as "semi unidirectional" in this study.

High sensible heat loads in such spaces can create non-isothermal conditions with large temperature difference between the supply air jet and its surrounding. Such thermal gradients can cause acceleration of the cold air jet and entrainment of the surrounding air mostly from the contaminated zone into the clean zone. Cook and Int-Hout (2009) explained that the unidirectional downward flow from laminar diffusers in a non-isothermal environment (such as in the operating room) accelerates as it approaches the surgical table. Their analysis indicates that the rate of acceleration increases with increasing thermal gradients, discharge velocity, and the area of supply diffuser array. Recent Computational Fluid Dynamics (CFD) analysis of a legacy HVAC configuration for a hospital operating room with a ceiling array of laminar supply diffusers and low exhaust grills on the opposite walls indicated that increasing the supply airflow rate can significantly reduce the thermal gradients which in turn reduces relative acceleration of the centerline velocity of the air jet. However, the highest acceleration in the centerline velocity occurs between 36 to 38 percent of the vertical distance from the ceiling diffuser irrespective of the supply airflow rate (Khankari 2017a). The subsequent analysis indicated that the HVAC configuration of the operating room can affect such acceleration. When the low wall return grills were replaced by series of ceiling grills the maximum centerline acceleration was reduced by almost 30 percent. This was also attributed to the reduction in the thermal gradients across the supply air jet (Khankari 2017b).

The current CFD study analyzes the behavior of semi-unidirectional airflows with a simple setup involving a single panel diffuser, a table, heat source, and two symmetric exhaust grills on opposite walls. The main goal of this study is to systematically evaluate the impact of supply air velocity on overall airflow patterns, temperature and velocity distribution in the space under isothermal and non-isothermal conditions with two different locations of sensible heat sources. Additionally, the impact of discharge velocity on the relative acceleration of the centerline velocity of the supply air jet is also compared.

VIRTUAL SET-UP OF A TEST CHAMBER

A three-dimensional, steady state CFD model of a virtual test chamber is developed for this study (Figure 1). The chamber has 168 sq. ft. (15.6 [m.sup.2]) floor area (14 x 12 feet, 4.3 x 3.6 m) with 10 feet (3 m) ceiling height. Figure 1 shows three different test configurations. A single panel laminar supply diffuser (2 x 4 feet, 0.61 x 1.2 m) is placed at the center of the ceiling at 10 feet (3 m) elevation. A 3 foot (0.9 m) high table (5 x 3 feet, 1.5 x 0.9 m) is placed at the center of the chamber under the supply diffuser. This table creates an obstruction to the airflow. The air is exhausted out through two long grills (14 x 0.5 feet, 4.3 x 0.15 m) which are placed symmetrically along the long sides of the chamber at 6 inches (15 cm) above the floor. For the two separate non-isothermal analyses a fictitious heat source representing a constant sensible heat load of 3412 Btu/h (1000 W) was placed in the room. In one case the source was placed at the center of the table under the supply diffuser (Figure 1b) and in the other (Figure 1c) two heat sources of 1706 Btu/h (500 W) each were placed on either side of the supply diffuser which represent ambient lighting or stratification near the ceiling. As noted the total sensible heat load in all non-isothermal cases kept constant.

A total of four supply airflow rates ranging from 200, 400, 600, and 800 cfm (94.4, 188.8, 283.2, and 377.6 lps) corresponding to the discharge velocities of 25, 50, 75, and 100 fpm (0.13, 0.25, 0.38, and 0.51 m/s), respectively were analyzed for both isothermal and non-isothermal cases. In the case of non-isothermal analyses the cooling capacities of the supply air correspond 1, 2, 3, and 4 times of the sensible heat load, respectively. The supply air temperature in all these cases was maintained at 65 F (18.3 C). A standard k-epsilon (k-e) turbulence model was employed to compute the turbulent viscosity of the air. Numerical computations were performed using Ansys Fluent CFD software. Computational results are presented in terms of normalized velocities which are obtained by dividing the local velocity values by the respective discharge velocities in each case.

RESULTS AND DISCUSSION

Isothermal Analysis

Figure 2 and Figure 3 show the airflow patterns and velocity distribution at the central plane. Figure 10a shows variation of normalized velocity - a ratio of centerline velocity at a specific distance along the vertical centerline to the respective discharge velocity. This variation is plotted against the non-dimensional vertical distance--a ratio of the height at a specific vertical location to height of the supply diffuser from the table. Thus, at a non-dimensional height of 0.0 at the supply diffuser, the non-dimensional centerline velocity is 1.0.

These analyses indicate that under isothermal conditions the supply air flow rate or discharge velocity has little impact on the overall airflow patterns and velocity distribution. The overall airflow patterns exhibit an inverted "Y" shape. The supply air stream maintains almost unidirectional flow for about 60 percent of the distance from the ceiling. Thereafter the downward velocity begins to decrease gradually and eventually reaches to a zero value near the surface of the table (Figure 10a). The main stream of the supply air, as it approaches the table, splits into two sideward jets which move towards the exhaust grills. The velocity distribution indicates that the sideward jets gain slight momentum as they separate from the table. Such airflow patterns create air recirculation zones above and below the sideward jets, although the recirculation zones adjacent to the table are formed at much lower velocity (Figure 2). Figure 10a also indicates a slight increase the centerline velocity between 15 to 50 percent of the vertical distance which indicates a slight entrainment of the surrounding air into the fast moving downward jet.

These analyses indicate that perhaps the room layout including the height of the space, height and locations of obstacles and the HVAC configuration including the size, type, and locations of exhaust grills as well as the number and locations of supply diffuser can impact the airflow patterns. However, it seems for a certain layout and HVAC configuration the supply airflow rate would have a little impact on the overall airflow patterns and the directionality of the supply jet.

Non-Isothermal Conditions--Heat Source on the Table

Figure 4, Figure 5, and Figure 6 show the airflow patterns, velocity, and temperature distribution, respectively at the central plane. These analyses indicate supply airflow rate or discharge velocity can significantly impact the performance of the unidirectional jet under non-isothermal conditions. These analyses show significantly two different flow patterns: one for the supply airflow rates up to 600 cfm (283.2 lps) and the other for the supply airflow rate for 800 cfm (377.6 lps). When the heat source is placed on the table the upward moving hot buoyant plume from the heat source pushes the downward moving cold air from the diffuser. At the same time the upward moving plume also entrains the air surrounding heat source. As this plume gets colder it cannot sustain the upward motion and starts falling sideward. In the case of low cooling capacity supply air jet (200 cfm, 94.4 lps) the weak cold supply air jet cannot sustain unidirectional motion and starts spreading sideward right after the discharge. It is evident from the temperature distribution (Figure 6a) that the low temperature zone is present only in the vicinity of the supply diffuser. With increasing supply airflow rate (or the cooling capacity), the downward moving jet can sustain unidirectional motion. As shown in Figure 6 with increasing the supply airflow rate the zone of high temperature starts shrinking and the average temperature in the chamber starts decreasing.

When the cooling capacity of the supply air reaches four times the cooling load (800 cfm, 377.6 lps), the jet becomes unidirectional and overcomes the upward buoyant force from the heat source. The cold supply air gets hot as it passes through the heat source and moves sideward and upward along the side walls of the chamber. As a result large recirculation zones are formed on both sides of the table which results in entraining the surrounding warm air into the cold supply jet (Figure 6d). Such recirculation results in thermal stratification in the chamber forming a zone of relatively cold air near the floor.

Figure 10b shows variation of normalized centerline velocity with normalized vertical distance. It shows up to the supply flow rate 600 cfm (283.2 lps) the discharge velocity decreases until about 70 percent of the vertical distance where it intersects with upward moving plume from the source. It should be noted such point of intersection moves downward with increasing discharge velocity (Figure 10b). In the case of supply airflow rate of 800 cfm (377.6 lps) such trend reverses entirely. The centerline velocity increases as the jet moves downward and reaches a peak value of 1.38 at about 67 percent distance from the diffuser. Thereafter the centerline velocity shows sharp decline until it reaches the top of the table. Such a trend indicates entrainment of the surrounding air into the supply air stream which is evident from the shrinking stream of cold air as it approaches the table (Figure 6d).

Non-Isothermal Conditions--Heat Source near the Ceiling

Figure 7, Figure 8, and Figure 9 show the airflow patterns, velocity, and temperature distribution, respectively at the central plane. These analyses also indicate supply airflow rate or discharge velocity has significant impact on the performance of the unidirectional jet under non-isothermal conditions. When the heat sources are located near the ceiling it sets natural convection circulating flows between the cold supply jet and surrounding hot air. Heat sources entrain the surrounding air which in turn gets entrained into the cold supply jet causing acceleration of centerline velocity (Figure 10c). These analyses also show two significantly different flow patterns: one for the supply airflow rates of 200 cfm (94.4 lps) and the other for the supply airflow rates above 200 cfm (94.4 lps).

The low supply airflow rate of 200 cfm (94.4 lps) maintains hot air near the ceiling while low temperature air spreads in the chamber (Figure 7a, 8a, and 9a). Due low discharge velocity the cold supply jet cannot overcome the entrainment towards the heat sources, and thus, cannot maintain strong unidirectional downward flow. It starts spreading sideward after the discharge, and thus, prevents the hot air spreading into the room. As a result the chamber remains at relatively lower temperature than the air near the ceiling. Figure 10c shows lowest acceleration in the centerline velocity due to uniform low air temperature surrounding the supply jet. Unlike the other cases, in this case the centerline velocity increases during the initial 20 percent travel, and then, it starts gradually decreasing as it approaches the table.

The other cases with higher supply flow rates show different patterns. With increasing discharge velocity the supply air jet becomes more unidirectional and the cold air reaches the top surface of the table. It should be noted only in the case of 800 cfm (377.6 lps) supply flow rate the cold stream reaches the table (Figure 9d). In this case the cooling capacity of the supply air is four times the sensible heat load. This is consistent with the analysis of the previous scenario where the heat source was placed on the table. With increasing air flow rate and increasing cooling capacity of the supply air the thermal gradients in the room start decreasing which results in reducing the acceleration of the centerline velocity (Figure 10c). In the case of 400 cfm (188.8 lps) supply airflow rate, the peak centerline velocity reaches about 2.2 times of the initial discharge velocity at about 60 percent distance from the ceiling. Whereas in the cases of 600 and 800 cfm (283.2 and 377.6 lps) the peak centerline velocity reaches about 1.84 and 1.49, respectively. These peak velocities occur at about 67 and 62 percent distance. In general the location of peak velocity shifts downward with increasing flow rates. As mentioned before low supply airflow rate of 200 cfm (94.4 lps) does not follow this trend. Otherwise it would have shown the highest increase in the centerline velocity. When the heat sources are located near the ceiling, low supply airflow rates create uniform temperature surrounding the supply air stream (Figure 9a).

Archimedes Number (Ar)

Archimedes Number (Ar), a non-dimensional parameter, is a ratio of the upward buoyancy force and the inertial force of the downward air jet. The supply airflow rates of 200, 400, 600, and 800 cfm (94.4, 188.8, 283.2, and 377.6 lps) correspond to the discharge velocities of 25, 50, 75, and 100 fpm (0.13, 0.25, 0.38, and 0.51 m/s), respectively. The corresponding Archimedes number for these supply airflow rates are 14.6, 1.8, 0.5, and 0.2 respectively. Increasing the discharge velocity (increasing the mass flow rate of the supply air) reduces theoretical temperature difference between supply and return temperature (delta T), which in turn results in lowering the Archimedes Number. Therefore, at higher airflow rates the lower values of Ar indicate flow dominated by inertial force of the downward air jet. As evident from the above non-isothermal analyses at low Ar number (200 cfm, 94.4 lps) the flow field and temperature distribution were dominated by the plumes and entrainment due to the heat sources. Only when the Ar number was less than 1 such as in the cases of 600 (283.2 lps) and 800 cfm (377.6 lps), the downward supply air jets could overcome the buoyant forces. Please note in these cases cooling capacities of the supply air were 3 and 4 times that of the respective cooling loads. It is also evident from these analyses that predominantly unidirectional flows can be obtained when the Ar number is smaller than 0.2

SUMMARY AND CONCLUSIONS

Unidirectional flows can be obtained either by supplying the air through HEPA filters covering the entire ceiling or by placing restraining panels around the array of supply (laminar) diffusers. In most other cases the flows are "semi unidirectional"--unidirectional only in the core of the supply air stream and recirculating in other locations. This study analyzes behavior of such semi unidirectional flows under isothermal and non-isothermal conditions. A three-dimensional, steady state Computational Fluid Dynamics (CFD) model of a virtual test chamber is developed to study the effect of supply airflow rates/discharge velocity on the airflow patterns, velocity, and temperature distribution. A total of four supply airflow rates ranging from 200 cfm (94.4 lps) to 800 cfm (377.6 lps) with corresponding discharge velocities ranging from 25 to 100 fpm (0.13 to 0.51 m/s) are analyzed. Non-isothermal analyses were performed for two different locations of sensible heat sources--one with the source on the top of a 3 foot (0.9 m) high table and the other with heat sources near the ceiling surrounding the supply diffuser panel.

For the conditions analyzed in this study indicates that under isothermal conditions the discharge velocity (flow rate) has little impact on the flow behavior of the supply air jet. The velocity of the isothermal downward air jet starts decreasing (and becomes non-unidirectional) after traveling about 60 percent of the distance from the discharge. However, in the case of non-isothermal conditions this analysis showed that the discharge velocity of the air jet can significantly affect the performance. When heat source was placed on the table, the upward buoyant plume from the heat source disrupted the downward motion of the cold supply jet. Only when the cooling capacity of the supply air was increased to four times the cooling load (discharge velocity 100 fpm, 0.51 m/s) the supply air jet could overcome the upward buoyant force and reach the top surface of the table. However such flow showed signs of entrainment of the surrounding hot air into the downward moving cold jet. When the heat sources were placed near the ceiling, the hot air surrounding the heat sources entrained into the downward moving jet causing acceleration of the centerline velocity, which showed to decrease with increasing airflow rate. Interestingly such acceleration was lowest for the lowest supply airflow rate 200 cfm, 94.4 lps (discharge velocity 25 fpm, 0.13 m/s).

This study further indicates that for non-isothermal conditions Archimedes Number can be a good initial indicator for estimating the directionality of the supply air jet. HVAC designs with Archimedes number smaller than one can overcome the upward buoyant forces and reduce the entrainment in the space. However, this number should be smaller than 0.2 for the cold air jet to travel longer distance without substantially losing its unidirectional behavior.

REFERENCES

Cook, G and Dan Int-Hout. 2009. Air motion control in the hospital operating room. ASHRAE Journal 51(3), 30-36.

Khankari, Kishor K. 2017a. Analysis of Airflow Distribution and Contaminant Flow Path in the Hospital Operating Room. ASHRAE Conference Paper: (LV-17-C008). ASHRAE Winter Conference, January, 2017. Las Vegas, NV.

Khankari, Kishor K. 2017b. Analysis of HVAC Configurations for a Hospital Operating Room. ASHRAE Conference Paper: (LB-17-C004). ASHRAE Annual Conference, June, 2017. Long Beach, CA.

Kishor Khankari, PhD

Fellow ASHRAE
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Author:Khankari, Kishor
Publication:ASHRAE Conference Papers
Article Type:Report
Date:Jan 1, 2018
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