8 CASE STUDIES
8.1 Dimensioning of diffusers and design of different mixing ventilation systems 81
ventilation systems
The following case study demonstrates the application of some of the concepts de-scribed in Chapters 3 and 6.
First the ventilation in a room with a sin-gle wall-mounted diffuser is considered.
The diffuser is located 20 cm below the ceiling and the air flow from the diffuser has an upward direction of 45o (Figure 8.1). The flow after the deflection below the ceiling is a semi-circular wall jet, i.e. it has the characteristic flow of a three di-mensional wall jet in the symmetry plane and lower velocity outside this plane. The velocity in the symmetry plane can be described by eq. 3.3 and a Ka factor of 3.0.
The diffuser area ao is 0.0087 m2.
Figure 8.1. Discharge of air from wall-mounted diffuser location below ceiling.
The room has a length, L, of 4.6 m in the direction of the symmetry plane of the flow.
A heat load of 381 W is to be removed from the room. The flow rate can be found when it is considered that the diffuser is able to handle a maximum temperature difference of 11 K.
The heat load from the room Q (W) is given by the expression:
o o pq T c
Q (8.1)
where ρ is 1.2 kg/m3 and cp is 1 005 J/K kg. The flow rate, qo, is calculated by the expression:
h / m 103
~ s / m 0287 .
0 3 3
o p
o c T
q Q
(8.2)
In order to clarify whether this is a well-designed air distribution system in the case of isothermal flow, the throw of the flow from the diffuser can be found from the graph in Figure 8.2 (often given in the catalogue of the product supplier ). The throw length, Th, is compared to the room length L.
Figure 8.2. Throw length, lTh, flow rate, qo and pressure difference, Δpo, for the diffuser in Figure 8.1.
The throw length can also be found from equation (3.3) by replacing ux with 0.2 m/s
and x by the throw lTh (xo = 0.0) or directly well-designed air distribution system under isothermal flow conditions. When dealing with non-isothermal flow (e.g. ΔTo = 11 K) the penetration length of a wall jet be-low the ceiling can be found by equation
Supply temperature difference To 11K Penetration depth, xs, is calculated as
m
Ksa is a constant which depends on param-eters outside the wall jet, such as room dimensions, location of thermal load, etc.
The penetration length should be as long as possible because entrainment reduces the velocity and increases the temperature of the wall jet before it flows into the listed in Figure 8.3.
C
Figure 8.3 .The results of full-scale
experiments with the diffuser in Figure 8.1. The flow is 0.0287 m3/s and the temperature difference between return and supply is 11 K.
The results shown in Figure 8.3 are ob-tained in physical experiments made under conditions similar to the calculations in this chapter. The flow was 0.0287 m3/s and the temperature difference between return and supply was 11 K. The penetra-tion length was short, 1.4 m. The meas-urements reveal no draught in the mean plane of the room. The velocity of 22 cm/s at the floor (.Figure 8.3) is probably caused by the downward flow at the side walls generated by the semi-radial flow at the ceiling.
The importance of selecting the right dif-fuser is demonstrated by different solu-tions discussed in the following examples.
Influence of the Ka factor
First, a nozzle (with Ka = 7.0) is selected as a new diffuser to be used in the room (L=4.6 m). The throw of the supplied flow must be unchanged in order to obtain a
draught free occupied zone. The thermal load in the room is also unchanged.
This means that lTh5Kauo ao should be unchanged and equal to room length. Thus for the selected new diffuser and for the diffuser used in the previous example (“old diffuser”) it is valid that
Kauo
old
Kauo
new (8.7)The consequence of using a nozzle with a high Ka value is an extreme low supply temperature To. This type of solution has been discussed in the United States to minimize the channel dimensions in the building and also for minimizing the pow-er consumption of the fans.
The penetration length is given by:
m
When the room is ventilated by one diffus-er only, it is important to use a diffusdiffus-er with the lowest Ka factor to get the largest xs and the smallest temperature difference.
Influence of supply area ao
The supply area ao is increased by a factor of 4.
The throw lTh 5Kauo ao should be unchanged (and equal to the room length), giving penetration length xs.
Number of diffusers
Three nozzles, each with a supply area ao
of 0.0087 m2, are used in the room. They are mounted side by side with a horizontal distance so that the flows from one nozzle are uninfluenced by the other nozzles (Figure 8.4).
Figure 8.4. Three wall-mounted nozzles in a ventilated room.
One nozzle (ao = 0.0087) has the correct throw for uo = 1.4 m/s. If three jets just
cover the width of the room the same throw should be used for three nozzles.
Therefore are uninfluenced by the other nozzles).
Maximum thermal load
The connection between the removed thermal load and the supply velocity is addressed in the following. The diffuser introduced first in this case study is in-stalled in the room (Ka = 3.0). The pene-tration length of 1.46 m is considered to be the shortest acceptable and should be kept unchanged, which means that uo2/ΔTo
The ratio between the head loads is
381/161 = 2.37 (8.17) and the ratio between the inlet velocities in third power is
(3.3/2.48)3 = 2.36 (8.18) Conclusion:
The maximum thermal load that can be handled by the air distribution system is a function of the supply velocity in the third power. See also equation (6.3).
Case 8.1 illustrates how to use some of the equations in this guide book and it pro-vides some general conclusions valid in case of mixing ventilation.
8.2 Airflow interaction in spaces Room air distribution is result of a com-plex interaction of the ventilation flow with thermal flows generated by heat sources such as occupants, office ma-chines, solar radiation, etc. Different ex-amples of airflow interaction and its pos-sible impact on occupants’ comfort are demonstrated in Figure 8.5. Depending on the strength and the location of the heat sources, the flow interaction can change completely the air distribution in rooms. In the upper example shown in Figure 8.5 the thermal flow generated by a computer and an occupant counteracts the downward flow of cold and clean air supplied from two ceiling installed slot diffusers and prevents from draught discomfort. How-ever a strong thermal flow which can counteract airflow from supply diffusers may increase the risk of draught as shown in the middle example in Figure 8.5. The combined strong thermal flow from occu-pant, computer and heated window blocks the ventilation flow supplied from the left
side of a slot diffuser resulting in a strong-er flow discharged from the right side of the diffuser which in absence of other obstacles may move tangentially in the room. Eventually the cool air at relatively high velocity will reach occupant’s feet and may cause draught discomfort.
The interaction of the thermal flow with the ventilation flow will have different effect in case of presence of obstacles or another thermal flow for example generated by a second occupant as shown in the bottom example. In this case the discharged flow from the left side of the slot diffuser will merge with the flow supplied from the right side of the diffuser. The combined flow with enough strength will be able the over-come the weaker thermal flow generated by the second occupant and will increase the cooling of his/her upper body part which may lead to complaints due to draught.
Thus a ventilation flow designed to provide comfort for occupants may cause discom-fort due to draught, eye irritation, etc. In rooms in practice the airflow interaction may change during the day and thus the room airflow distribution may change (oc-cupants coming in and out of room, chang-es in solar radiation, etc.). Changchang-es in furni-ture layout in time also will affect the air-flow interaction and thus the air distribution in rooms. Thus occupants may be exposed to changeable indoor environment.
It is difficult to calculate the airflow inter-action with simple flow element models as those discussed in Chapter 3. Computa-tional Fluid Dynamics (CFD) modeling (Chapter 5) can be used only in simple cases of airflow interaction, e.g. assisting flows and transverse flows (see Fig-ure 2.4, Chapter 2). However the mostly used today CFD modeling based on
Reyn-olds-Averaged Navier-Stokes (RANS) equations fails to predict one of the most common cases of airflow interaction in rooms between thermal plume generated by heat sources and opposing downward ventilation airflow. As already discussed the airflow interaction in rooms is im-portant for occupants’ comfort and there-fore it has to be considered. Although rela-tionships for its calculation are not availa-ble it can be in many cases predicted based on experience or it can be studied by measurements and visualization under laboratory conditions and in practice. In the following several examples of airflow interaction are presented and discussed with focuse on its impact on occupants’
comfort.
Figure 8.5. Airflow interaction affects air distribution in rooms and thus occupants’
comfort.
Example 1: Thermal plume from standing human body and downward ventilation flow
An important element of the design proce-dure for an air distribution system is to ensure an acceptable low air velocity (typ-ically 0.2 m/s) where the flow passes the imaginary surfaces defining the occupied zone. The design value at the upper hori-zontal surface of the occupied zone can have higher values, depending on the type of air distribution system used. The ther-mal plume above a person has an upward velocity of approx. 0.25 m/s, and this flow often prevents draught at head height.
Figure 8.6 shows the thermal boundary layer around a standing thermal manikin with simplified body shape with size and heat production as an “average” person.
The free convection flow around the man-ikin is visualized by smoke. The interac-tion between the thermal plume generated by the manikin and a downward flow with different velocity is demonstrated with several photos. The boundary layer is pre-served with a downward velocity up to 0.25 m/s (Nielsen 2009).
Example 2: Ceiling mounted air supply nozzles
This example demonstrates the interaction between thermal plume generated by a seated human body with downward flow from ceiling mounted nozzles used for ventilation of separate workstations in a room. First, the interaction was studied by physical measurements employing dressed thermal manikin with body size and shape accurately resembling an average Scandi-navian woman. The body of the manikin was divided into 23 body segments each with surface temperature controlled to be as the skin temperature of an average per-son in state of thermal comfort under the
Figure 8.6. A thermal manikin with simplified body shape located in a downward air flow.
The boundary layer around the manikin at head height is preserved up to a downward velocity of 0.25 m/s.
studied thermal environment. The manikin was positioned bellow the flow from a nozzle with discharge diameter 0.095 m positioned at distance 2.5 m above the floor and 1.3 m above manikin’s head as shown in Figure 8.7. The interaction of the thermal plume generated by the mani-kin and the jet was studied by comprehen-sive measurements of the velocity field above the manikin as well as the heat loss from the manikin’s body segments. Multi-channel low velocity thermal anemometer with omnidirectional (spherical) velocity
sensor was used during the measurements.
Thus the speed in the flow was identified.
The airflow interaction was studied at different jet supply flow rate (i.e. different supply velocity) under isothermal and non-isothermal conditions, i.e. supplied jet air temperature equal and lower than the room air temperature. The results are reported in detail by Yang et al. (2008, 2009).
Figure 8.7. Seated thermal manikin exposed to downward airflow from a ceiling mounted nozzle. The airflow of interaction of the supplied jet and the thermal plum generated by the thermal manikin is studied by velocity measurements in the defined points.
In Figure 8.8 velocity distribution across the jet at distance 1.2 m (x/D=11.58) from the nozzle measured at supply flow rate of 4 L/s (initial jet velocity of 0.57 m/s) and isothermal conditions of 23.5°C is shown.
Three cases are compared: free jet without thermal manikin, jet with unheated mani-kin and jet with heated manimani-kin. The ve-locity profiles across the jet without the
presence of the manikin have typical Gaussian distribution with maximum ve-locity at the axes of the jet. The veve-locity distribution is substantially disturbed when the manikin is present and acting as an obstacle (unheated manikin). The velocity distribution is further disturbed by the thermal plume generated in the case of heated manikin.
Figure 8.8. Profiles of air speed across downward isothermal air jet from a nozzle (diameter of 0.095 m) with initial velocity of 0.57 m/s measured at distance 1.2 m in the case of free jet, with presence of unheated manikin (obstacle) and with presence of heated manikin generating thermal plume (Yang et al.
2009).
Human response to the conditions dis-cussed above is shown in Figure 8.9. The average thermal sensation at the face of thirty human subjects exposed to the downward flow from the nozzle at target velocity (measured at distance 0.2 m above the head) of 0.13, 0.36, 0.55 and 0.76 m/s (corresponding to supply flow rate of 4, 8, 12 and 16 L/s) is shown in the figure. It is clear from the figure that at low flow rate the strength of the jet is not enough to penetrate the thermal plume above people. Therefore, the subjects
par-ticipating in the experiment reported neu-tral thermal sensation. The strength of the jet increases with the increase of the sup-plied flow rate and it starts to cool the face resulting in slightly cool average thermal sensation reported by the subjects at 16 L/s (Yang et al. 2010). In this case the venti-lating flow is opposing to the thermal plum generated by human body. The pene-tration of the thermal flow leading to in-crease of the body cooling will be easier when the two flows are not acting against each other but are displaced.
Figure 8.9. Average facial thermal sensation reported by 32 subjects exposed to airflow from above at isothermal conditions of 23.5°C.
Thermal sensation scale: - 2 – cool, -1 –slightly cool, 0 – neutral, 1-slightly warm, 2 – warm.
Example 3: Passive beam installed over workplace
The passive battery type element itself is essentially a series of water tubes with fins.
Cooled water circulation through the coil generates airflow across the beam driven by natural buoyancy. Under undisturbed con-ditions, a flow of cold air is directly falling down into the occupied zone.
Experimental study of the velocity field generated by passive beams has been con-ducted (Fredriksen et al. 2001). The study indicated that the air velocity can create draught at specific cooling capacity of the beam higher than 200 W per linear metre.
As a rule of thumb (Virta et al. 2004) equates a specific cooling capacity to be
lower than 150 W/m when the beam is in an occupied zone, and up to 250 W/m when installed in a non-occupied zone.
However, the underneath heat gains have a significant effect on the velocity profile of a passive beam. The underneath heat load used in a study (Saarinen et al. 2007) comprised a PC and a human body (cylin-der), simulating an office situation. The heated floor covered a much larger area than the load below the beam. The basic arrangement is shown in Figure 8.10.
Figure 8.10. Heat load underneath a passive beam. Basic arrangement of the experimental setup is shown (Saarinen et al. 2010). All measures are in millimeters. Vertical middle line (red) marks the position of the beam centre point.
Two additional arrangements were also used. In one of them the heat load was shifted 0.5 m aside from the beam centre line, and in the other it was removed alto-gether. The human body simulator was constructed according to the standard DIN
4715-1 (1995), containing a description of a cooling load simulator. The dummy was heated electrically with a power of 90 W.
Depending on the cooling power 38 – 58%
of the heat loss from the dummy was through convection. Moreover, the dummy used in the experiments originally had three large holes on its flank, near the top sur-face. They had to be blocked, since the warm air blew horizontally out of them, when the dummy was under the chilled beam. The flow field was measured on four horizontal planes at 1.2 m, 1.5 m, 1.8 m and 2.1 m above the floor as well as on a verti-cal plane below the beam centre line. The density of the grid of measuring point on each horizontal plane was 0.1 m × 0.1 m.
The plumes from the heat loads were not powerful enough to fight the downward flow from the beam. Instead the heated air was captured by the downward flow from the passive beam and no significant up-ward buoyancy force was generated. This was true even with low cooling power.
Only when the load was shifted, or when an extra load was added, did an upward plume build up (Figure 8.11).
Example 4: Different heat load distri-bution
The impact of the heat load distribution on the air distribution in a room was studied in a test room (Kosonen et al. 2007a). Two chilled beams and six heated cylinders posi-tioned as shown in Figure 8.12 were used.
The primary airflow rate was 2 L/s/m2, room temperature was 24±0.2°C, surface temperature of windows was 41°C. The heat load was constant 84 W/m² and was distributed as follows: window - 350 W, six dummies - 720 W, heated floor (darker area in Figure 8.12) -300 W, lightning (incorpo-rated in the chilled beams) – 140 W. Two experiments were performed - with all heat sources located at one side of the room (Figure 8.12 top) and with the heat sources split on the both sides of the room (in this case the heated cylinders were located op-posite to the window (Figure 8.12 bottom).
Figure 8.11. Velocity profiles at different heights at three load setups at equal cooling power per length of the beam (about 170 W/m).
load: human +
PC + display extra load 200 W near
the PC load aside of the
beam centre line
vz [m/s]
Figure 8.12. Layout of chilled beams and heat sources used to study the impact of the heat load distribution on the room air distribution.
The performed smoke visualization shown in Figure 8.13 clearly identifies the differ-ent air distribution in the room due to the flow interaction. In the case of heat load concentrated on the right side of the room the generated strong thermal plume dis-charges the air supplied from the chilled beams (traced with smoke) toward the left side of the room. In the case of split heat load distribution the interaction of the thermal plumes generated on the two sides of the room with the ventilation flow sup-plied from the beams results in different room air distribution.
Comprehensive measurements performed with multichannel low velocity anemome-ter confirmed the importance of the air-flow interaction on the room air distribu-tion. The areas with constant velocity drawn based on the measurements identify higher velocities on the left side of the
room in the case of concentrated heat load location (Figure 8.14 top) while the high-est velocity in the middle of the room oc-curs when the heat load is spread (Figure 8.14 bottom). In practice both situations may occur due to occupants’ activities and changes in the heat load from the solar
room in the case of concentrated heat load location (Figure 8.14 top) while the high-est velocity in the middle of the room oc-curs when the heat load is spread (Figure 8.14 bottom). In practice both situations may occur due to occupants’ activities and changes in the heat load from the solar