Displacement Ventilation

Risto Kosonen (ed.) Arsen Melikov Elisabeth Mundt Panu Mustakallio Peter V. Nielsen rehva Federation of European Heating, Ventilation and Air Conditioning Associations

GUIDEBOOK NO 23

Risto Kosonen (ed.) Arsen Melikov Elisabeth Mundt Panu Mustakallio Peter V. Nielsen

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ISBN 978-2-930521-17-6

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1 DISPLACEMENT VENTILATION IN A NUTSHELL...... 1

2 TERMINOLOGY, SYMBOLS AND UNITS ...... 4 2.1 Terms and definitions ...... 4 2.2 Symbols ...... 5

3 ROOM AIR DISTRIBUTION ...... 8 3.1 Need for Ventilation ...... 8 3.2 Ventilation and room air distribution principles ...... 8 3.3 Displacement ventilation and thermal comfort ...... 11 3.4 Displacement ventilation and air quality ...... 12

4 PERFORMANCE OF DISPLACEMENT VENTILATION ...... 15 4.1 Displacement Ventilation Method ...... 15 4.2 Air flow pattern ...... 15 4.3 Temperature distribution ...... 16 4.4 Convection flows – the engines of displacement ventilation ...... 20 4.5 Contamination distribution ...... 28 4.6 Ventilation effectiveness ...... 29

5 CALCULATION OF SUPPLY AIRFLOW RATE ...... 34 5.1 Temperature based design methods ...... 34 5.2 Calculation of vertical room air temperature distribution ...... 35 5.3 Vertical position of the heat source ...... 39 5.4 Calculation examples when using temperature based design models ...... 39

6 AIR DIFFUSERS FOR DISPLACEMENT VENTILATION ...... 42 6.1 Commonly used diffusers ...... 42 6.2 Radial air flow or plane air flow from low-velocity diffusers ...... 44 6.3 Air flow from low –velocity diffusers ...... 44 6.4 Air distribution from a low-velocity diffuser giving a radial flow in the occupied zone ...... 45 6.5 Air distribution from wall-mounted diffusers giving plane flow in the occupied zone ...... 53 6.6 Air distribution from floor-mounted diffusers ...... 54

iii

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 7 DESIGN OF DISPLACEMENT VENTILATION ...... 56 7.1 Design criteria ...... 56 7.2 Design of air distribution ...... 56 7.3 Integration with separate heating and cooling systems ...... 60 7.4 Control of indoor conditions ...... 64

8 CASE STUDIES ...... 67 8.1 Air distribution with four typical air supply methods in a classroom ...... 67 8.2 Comparison of calculated and measured vertical temperature gradients for displacement air distribution ...... 70 8.3 Field measurements for a multipurpose arena ...... 72

9 RESEARCH FINDINGS ...... 74 9.1 A CFD Benchmark test for manikins in displacement flow ...... 74 9.2 Full-scale tests and CFD- simulations of indoor climate conditions ...... 74 9.3 Test on the performance of displacement ventilation– proper simulation of occupants ...... 77 9.4 Airborne cross infection risk in a room with displacement ventilation ..... 80 9.5 Displacement ventilation design based on occupants’ response ...... 83 9.6 Convective boundary layer around human body ...... 87

10 REFERENCES ...... 91

Displacement ventilation is primarily a This guide discusses methods of total vol- means of obtaining good air quality in oc- ume ventilation by mixing ventilation and cupied spaces that have a cooling demand. displacement ventilation and the guide It has proved to be a good solution for book gives insight of the performance of spaces where large supply air flows are re- the displacement ventilation. It also takes quired. into account different items, which are cor- related, to well-known key words: free con- Some advantages of displacement ventila- vection flow; stratification of height and tion: concentration distribution; temperature distribution and velocity distribution in the oc- • Less cooling needed for a given tempera- cupied zone and occupant comfort. ture in the occupied space; • Longer periods with free cooling; The guide book discusses two principal • Potential to have better air quality in the methods which can be used when the sup- occupied spaces; ply air flow rate of displacement ventilation • The system performance is stable with all system is calculated: 1) temperature based cooling load conditions. design, where the design criterion is the air temperature in the occupied zone of the Displacement ventilation has been origi- room and 2) air quality based design where nally developed in Scandinavian countries the design criterion is the air quality in the over 30 years ago and now it is also a well- occupied zone. Some practical examples of known technology in different countries the air flow rate calculations are presented. and climates. Historically, displacement ventilation was first used for industrial ap- The air flow diffusers are the critical factor: plications but nowadays it is also widely most draught problems reported in rooms used in commercial premises. with displacement ventilation are due to high velocity in the zone adjacent to the dif- However, displacement ventilation has not fuser. This guide explains the principle for been used in spaces where it could give the selection of diffuser. added values. For that there are two main reasons: firstly, there is still lack of This guide also shows practical case studies knowledge of the suitable applications of in some typical applications and the latest displacement ventilation and secondly, research findings to create good micro cli- consulters do not know how to design the mate close to persons is discussed. system. These and some other aspects are discussed REHVA published 2002 the first version of in this book. Authors believe you will find displacement ventilation guide. The aim of this guide useful and interesting when you this revised Guidebook is to give the state- design or develop new ventilation solu- of-the art knowledge of the technology. The tions. idea of this guidebook is to simplify and improve the practical design procedure. The authors

REHVA, now 54 years old, is an organisation of European professionals in the field of building services (heating, ventilating and air–conditioning). REHVA represents more than 100,000 experts from 27 European countries. REHVA’s mission is to promote energy efficient and healthy technologies for mechanical services of buildings, and to disseminate knowledge among professionals and practitioners in Europe and beyond. REHVA Guidebooks are the most important tools to diffuse knowledge on latest developments, and advanced technologies providing practical guidance to practitioners. REHVA has published 22 guidebooks to date, this one on Displacement Ventilation is the 23rd. – Anita Derjanecz, REHVA Managing Director

Member countries of REHVA

Working Group

This book was developed with a working group consisting of the following experts: • Risto Kosonen (Aalto University, Finland) • Arsen Melikov (DTU Technical University of Denmark) • Elisabeth Mundt (KTH Royal Institute of Technology, Sweden) • Panu Mustakallio (Halton Oy, Finland) • Peter V. Nielsen (Aalborg University, Denmark)

Reviewers

This book was reviewed with a working group consisting of the following experts: • Hazim B. Awbi (Reading University, United Kingtom) • Klaus Fitzner (Technical University of Berlin, Germany) • Jarek Kurnitski (Tallinn Technical University, Estonia) • Alfred Moser (Science Services, Switzerland) • Marco Perino (Technical University of Torino, Italy) • Jorma Railio (SULVI, Finland)

Acknowledgements

The authors wish to thank REHVA's Technology & Research Committee and Publishing & Marketing Committee as well as REHVA's staff members for all their valuable contributions to the guidebook.

The authors would also like to thank Tim Dwyer for proof-reading the manuscript and Jarkko Narvanne for the graphical design and typesetting of the guidebook.

The idea Less suited for Displacement ventilation, as presented in Displacement ventilation may be less pref- this book, is considered to be the technique erable than mixing ventilation in the fol- of supplying clean, cool air at floor level, lowing cases: letting warm air and contaminants rise to • Where surplus heat is the main problem, the ceiling and extracting the contaminated and relatively low specific outdoor air- air at ceiling level (Figure 1.1). flow rate is needed; • Where there are space constraints for supply diffusers and duct work; • When the requirement is to cool in low height rooms (in offices, consider mixing and cooling panels or chilled beams); • Where there are significant disturbances to air flow near the floor (for example, furniture); Figure 1.1. The concept of displacement • Where the contaminants are cooler/ ventilation. denser than the ambient air.

Best suited for Strong points Displacement ventilation is primarily a Some advantages of displacement ventila- means of obtaining good air quality in oc- tion are: cupied spaces that have a cooling demand. • Less cooling needed for a given tempera- It has proved to be a good solution for: ture in the occupied space; • Gyms; • Longer periods with free cooling; • Meeting rooms; • Potential to have better air quality in the • Classrooms; occupied spaces; • Tall rooms: Convention centres, Lobbies, • The system performance is stable with all Sport arenas, Auditoriums, Theatres, Mu- cooling load conditions. seums, Airports, Shopping centres, etc. Weak points Displacement ventilation is usually prefer- Some weak points are: able in the following cases: • Possibility of cold draughts along the • Where the contaminants are warmer floor - use the suitable air supply units, and/or lighter than the surrounding air; and take care of the zone in front of the • Where the supply air is cooler than the diffusers; ambient air; • Sensitive to furniture arrangement in • In tall rooms, for example, where the rooms; room heights are more than 3 metres; • Displacement principle may be disturbed • When there are heat loads in the upper by walking occupants; part of room; • Wall mounted diffusers often require sig- • Where large supply air flows are required nificant wall space and reduce occupied in rooms. zone.

The air supply diffuser – a crucial factor Most draught problems reported in rooms with displacement ventilation are due to high velocity in the zone adjacent to the diffuser (Figure 1.2). It is important to choose a diffuser that is suited for the application and only utilise diffusers from manufactur- ers that supply robust documentation together with the products.

Figure 1.3. Diffusers integrated in benches with air supplied through perforated side plates (courtesy of Halton).

Figure 1.2. Diffuser has a limited near (adjacent) zone with high velocities where the risk of draught is high (see Chapter 6).

Collaboration with the architect is required The diffusers require a certain amount of wall area, or space in, or on, the floor. Close cooperation with the architect is required to find a suitable location for the air diffusers. The supply units can also be designed to fit different architectural requirements: units Figure 1.4. Free-standing diffusers as archi- could be invisible (Figure 1.3) or be an ex- tectural elements (courtesy of Halton). posed architectural element (Figure 1.4). The occupied zone – the coolest part Air flow rates of the room To reach the same air quality in the occu- In displacement ventilation, the air temper- pied zone, displacement ventilation typi- ature increases from floor to ceiling (Fig- cally requires a lower air flow rate than ure 1.5). This means that the occupied zone mixing ventilation. When the main task is is the coolest part of the room. Vertical tem- to remove excess heat, both mixing and dis- perature profiles measured with different placement systems are likely to require individual types of heat load (occupants, similar air flow rates.

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 1. DISPLACEMENT VENTILATION IN A NUTSHELL warm floor, warm window and warm ceil- With displacement, the supply air tempera- ing) are shown in Figure 1.5 (Kosonen et ture is typically about 3 K to 5 K cooler al. 2016). With the load dominated by oc- than the room air temperature at a height of cupants or by a warm floor, a two layer 1,1 m. In areas where people are moving structure is generated with heat and pollu- for example, in shopping centres, the sup- tion accumulated in the upper part of the ply air could be 6 K to 8 K lower than the room. The data indicate an obvious mixing room air temperature. Depending on the layer. Across the mixing layer, the room air particular design, the temperature differ- temperature could be assumed to be con- ence between the supply and exhaust air is stant or exhibiting just a slight increase. typically between 6 K and 15 K. The warm window produces a near linear temperature profile with no clear two-layer Compared with mixing ventilation, dis- structure. With the heated ceiling, the con- placement ventilation supplies air at a vection heat remains mainly in the upper higher temperature and this implies longer portion of the room. periods of the year where free cooling can be applied, and so less energy consumption for cooling the supply air.

H(m) H(m) Do not heat occupied rooms with 5 5 displacement ventilation

4 4 If an occupied room is to be warmed by the Warm ventilating air, displacement ventilation, as 3 3 window described in this book, is not suitable. If Occupants 2 2 warm air is supplied at floor level, in a room cooler than the supply air, it will rise due to 1 1 buoyancy, and be extracted when it reaches 0 0 the ceiling (Figure 1.6). Thus, the supply 0 0,4 0,8 1,2 0 0,4 0,8 1,2 air will short circuit into the outlet and little Temperature ratio Temperature ratio of the clean and heated warm air will reach H(m) H(m) the occupied space. Displacement air distri- 5 5 bution can be used for heating up spaces prior to them being occupied. 4 4

3 3 Warm floor Warm 2 2 ceiling

1 1

0 0 0 0,4 0,8 1,2 0 0,4 0,8 1,2 Temperature ratio Temperature ratio

Figure 1.5. Vertical temperature profiles in room with displacement ventilation with differ- Figure 1.6. Supply of warm ventilation air ent heat loads (temperature ratio= 휃−휃푠 ). means short-circuiting. 휃푒−휃푠

2 Terminology, symbols and units

2.1 Terms and definitions Buoyancy: The vertical force exerted on a volume of air that has a density different Adjacent zone: The zone in front of a dis- from the ambient air. placement air distribution diffuser where draught discomfort may occur. Displacement ventilation [displacement air distribution]: Room ventilation cre- Air change rate: The ratio of the volumet- ated by room air displacement, by intro- ric airflow rate supplied to a space related ducing air at low level in a space at a lower to the volume of that space. It is usually air temperature than the room air. measured in air changes per hour, and normally relates to the outdoor air change rate. Draught risk rating: Percentage of occupants predicted to be dissatisfied due to Air exhaust opening: Air terminal device draught. used to extract air from a space. Draught: Unwanted local cooling of the Air flow rate: Mass or volume flow of air human body caused by air movement. passing a given plane divided by the time. Face velocity: Average air discharge ve- Air flow: Continuous movement of air. locity from the diffuser (supplied airflow rate divided by face area). Air jet throw: The distance an air stream travels on leaving a diffuser before its ve- Indoor air quality: Attributes of the res- locity is reduced to a specific value. pirable atmosphere (climate) inside a building including gaseous composition, Air pollution: Any material in the atmos- humidity, temperature and contaminants. phere that affects people and their environment (pollutants include materials such as Isovel: Boundary contours of equal local liquids, solids, aerosols, gases and odours). mean air velocity.

Air stratification: The layering of air Local air velocity: Velocity in a specific within a space, due to density differences. point in an air stream at a specific time.

Air supply diffuser: Air terminal device Local mean air velocity: Magnitude of the used to supply ventilation air to a space. time-averaged vector of velocity at a point of an air stream. The velocity vector and Air temperature: Dry-bulb temperature of its components on an orthogonal coordi- the air. nate system in any point of a turbulent stream is subject to fluctuations with re- Air velocity: Rate of motion of air in a spect to time. The time averaged vector of given direction measured as distance per velocity is a vector for which each compo- unit time. nent is averaged with respect to time.

Mechanical ventilation: Ventilation with Turbulent flow: Flow that is character- the aid of powered air movement compo- ized by irregular eddies associated with nents. momentum transfer between fluid layers.

Mean velocity: Instantaneous velocity Ventilation flow rate: Volume flow rate at averaged for a period of time. which ventilation air is supplied or removed.

Mixing ventilation [mixing air distribu- Ventilation: Designed supply and re- tion]: Air diffusion where the mixing of moval of air to and from a treated space. supply air and room air is intended. 2.2 Symbols Occupied zone: Volume of a space between the floor and 1,8 m above the floor Latin letters and more than 1,0 m from outside A Area or floor area [m²] walls/windows, 0,5 m from inner walls Ar Archimedes number [-] and excluding the adjacent zone generated Af Floor area [m²] by displacement diffuser. Awl Lower wall area [m²] Awu Upper wall area [m²] Plume: The air flow rising from a hot ao Air diffuser supply area [m²] body (or descending from a cold body). af Air diffuser face area [m²] arf Archimedes coefficient [-] Reference air temperature in a room B Width (of an air diffuser) [m] with displacement ventilation: Average D Diameter (of a source) [m] of at least five measurements of the mean Fmo Mixing factor of convection flow (in time) air temperature at a height of [-] 1,1 m from the floor within the occupied H Height of diffuser or room [m] zone outside the area directly influenced Im Entrainment factor of convection by the flow from displacement air supply flow [-] diffuser. Ka Air diffuser constant, jet discharge [-] Speed: Magnitude of mean velocity. KDr Air diffuser constant, low velocity discharge, radial flow [-] Temperature: Measurement of warmth KDp Air diffuser constant, low velocity or coldness with respect to an arbitrary discharge, plane flow [-] zero or absolute zero. A physical prop- Kob Air diffusion constant for flow bet- erty related to the average kinetic energy ween obstacles of the atoms or molecules of a substance Lp Sound pressure level [dBA] (according to Collins English Diction- N Number of convection sources [-] ary). T Absolute temperature [K] (= + 273) Turbulence intensity: Ratio of the stand- Tu Turbulence intensity [ %] ard deviation of the air velocity fluctua- W Depth of an air diffuser [m] tions around the local mean velocity to the bm Flow adjustment factor for air sup- local mean air velocity. ply diffuser [-]

bn Half-width of the adjacent zone qv,z Vertical air volume flow [m³/s] [m] qv,l Vertical air volume flow in a c Contaminant concentration plume above a line source [mg/m³, ppm, etc.] [m³/(s·m)] ce Contaminant concentration at the hTx Height of mixing layer [m] extract air [mg/m³, ppm, etc.] hmx Height of lower wall [m] cexp Contaminant concentration in s Vertical temperature gradient = breathing air [mg/m³, ppm, etc.] z [K/m] cmean Mean contaminant concentration v Velocity [m/s] in the room [mg/m³, ppm, etc.] vs Face velocity = qo /As [m/s] coz Mean contaminant concentration vx Horizontal velocity (x-direction) in the occupied zone [mg/m³, ppm, [m/s] etc.] vx,max Maximum velocity in the vertical cs Contaminant concentration in the velocity profile at the floor supply air [mg/m³, ppm, etc.] vy Horizontal velocity (y-direction) cp Specific heat at the constant pres- [m/s] sure of the air = 1004 J/kg K vz Vertical velocity (z-direction) (1 J = 1 Ws) [m/s] d Diameter [m] vs Face velocity = qo /As [m/s] do Diameter of “vena contracta”, i.e. v Mean velocity [m/s] the most contracted cross section x Length co-ordinate [m] of a plume [m] y Width co-ordinate [m] e Entrainment coefficient in the dis- zexp Height to the breathing zone [m] charge flow from the diffuser [-] z Height co-ordinate [m] g Acceleration of gravity [m/s²] Stratification height [m] h Height [m] zmax Maximum height for a plume in ln Length of the adjacent zone [m] stratified surroundings [m] l0.2 Length of the adjacent zone (to the zo Height between virtual (point) 0,2 m/s isovel) [m] source and the source [m] n Number (of people) zp Height between virtual (point) RH Relative humidity [ %] source and the chosen reference pd Dynamic pressure = ½  v² [Pa] height [m] ps Static pressure [Pa] zt Equilibrium height for a plume in ptot Total pressure = pd + ps [Pa] stratified surroundings and height ptot Total pressure drop across a dif- of mixing layer[m] fuser [Pa] z*, z** Non-dimensional height for a qB Ventilation rate for emissions from plume in stratified surroundings building, m³/(s·m²) vf Face of diffuser [m/s] qs,l Supply air volume flow per m w Velocity of convection flow [m/s] width of the diffuser/room [m³/(s·m)] Greek letters qp Ventilation rate for occupancy per  Heat flux [W, W/m] person, m³/s, pers  Total heat flux =  +  tot cf r qs Supply air volume flow [m³/s] [W, W/m] qv Air volume flow [m³/s]  cf Convective heat flux [W, W/m]

r Radiative heat flux [W, W/m] θ Temperature difference [K]  Heat transfer coefficient s Under-temperature in the supply [W/(m² K)] air = oz − s [K] or e − s [K] cf Convective heat transfer coeffi- θz Difference between maximum air cient [W/(m² K)] temperature in a plume and ambi- r Radiative heat transfer coefficient ent air temperature [K] [W/(m² K)] θ Temperature [°C] o Angular spread of the radial flow a Air temperature [°C] from an air diffuser [rad] θob Air temperature in the obstacle  Thermal expansion coefficient of opening [°C] -1 air = 1/ (θ + 273 °C) ~ 1/300 K θe Exhaust air temperature [°C]  Thickness of stratified flow near θf Floor temperature [°C] the floor; δ is the height where ve- θ Air temperature near (0,05 m) the af locity vx= 0,5 · vx,max [m] floor [°C] a Air change efficiency, a measure θmx Air temperature at mixing layer [°C] of how quickly the air in the room θoc Air temperature at 0,65 m height [°C] is replaced [-] θoz Mean air temperature in the occu-  c Mean ventilation effectiveness. pied zone [°C] Also called contaminant removal θs Supply air temperature [°C] effectiveness. It is a measure of θr Upper room air temperature [°C] how quickly an airborne contami- θsu Surface temperature [°C] nant is removed from the room [-] θwl Surface temperature of lower wall c  oz Ventilation effectiveness of the oc- [°C] cupied zone. Also called air quality θwu Surface temperature of upper wall index of the occupied zone [-] [°C] c  P Local ventilation index. Also  Dimensionless temperature of the called air quality index at a given air near the floor [-] point P [-]  Air density. For normal room tem- c  exp Personal exposure index. Also peratures  = 1,20 kg/m³ (θ = 21 °C) called air quality index of the in-  Ratio of the cooled ceiling output haled air [-] to the total cooling output  Temperature effectiveness [-]

3 Room air distribution

This chapter discusses methods of total vol- ture near the ceiling. This may result in ume ventilation by mixing ventilation and lowering cooling peak power and cooling displacement ventilation. Other methods energy consumption when only the envi- for achieving thermal comfort and good air ronment of the occupied area is actively quality in spaces which are now under de- controlled. velopment (including localized radiant and convective system, stratum ventilation, It is possible to achieve good indoor condi- etc.) and advanced air distribution (such as tions in an energy efficient manner by the personalised ventilation) are not considered use of well-designed displacement ventila- in this chapter. tion.

3.1 Need for Ventilation 3.2 Ventilation and room air distribution principles In buildings, the supply of clean outdoor air is needed for breathing and the removal of a) Target levels internal heat loads, gases and particulates. The aim of air conditioning and distribution Heat is generated by occupants, equipment into rooms is to maintain, in the most eco- (PC’s, lighting, etc.) and solar radiation. nomical way, (accounting for energy usage Vapour, gases and particulates are gener- and cost efficiency) the desired thermal en- ated by occupants, building materials, of- vironment and air quality in the occupied fice equipment, etc. and also introduced by zone so as to meet target levels during dif- infiltration from outdoors. ferent operating conditions. Depending on the design criteria, the designer may choose The air supplied to spaces is either filtered different room air distribution methods in outdoor air or filtered outdoor air mixed order to achieve the specified targets. with re-circulated filtered room air. Fur- thermore, it may be needed to humidify or b) Methods of room air distribution dehumidify the supplied air. Displacement The room air distribution method is critical ventilation aims to provide occupants with for the air conditioning of spaces. Often air clean air for breathing more effectively distribution in rooms is assisted by radiant than fully mixed air distribution. With dis- or convective heating and cooling methods. placement ventilation, it is possible to uti- However, it must be noted that in some lize buoyancy flows that transfer contami- cases a strategy of room air distribution can nant from the occupied zone towards the also be fulfilled without any mechanical in- upper room zone and so improves the qual- stallations using only buoyancy forces. The ity of air inhaled by occupants. Simultane- classification of ideal room air distribution ously, with better air quality, displacement methods is summarized in Figure 3.1 ventilation creates a vertical temperature (Hagström et al. 2000). Note that piston gradient in the room, with a high tempera- flow requires large amounts of air.

Fitzner (1996) points out that piston flow T = absolute temperature of the supply air [K] from the floor and upwards exists for Ar- v = mean air velocity upwards chimedes numbers less than 360. = air volume flow/floor area [m/s]

hg  For Ar > 360, buoyancy forces will domi- Ar  360 (3.1) nate and create a thermally stratified flow. vT 2 where: The Archimedes number can be also ex- 2 g = acceleration of gravity = 9,81 m/s² pressed as Ar ~ ∆θs/vs in a given geometry, 2 H = height of the room [m] or as ∆θs/qs , called (Arratio), because the  = e – s = temperature difference supply area a0 is constant within a given ge- between exhaust and supply air [K] ometry in the following explanations.

Strategy DISPLACEMENT ZONING MIXING PISTON STRATIFICATION

Description Unidirectional Utilise density Air flow from Uniform flow through the differences clean zones to conditions in room contaminated all parts of zones the room

Room dimension Room dimension Room dimension Room dimension Air quality; e e e e temp,θ humidity, RF contaminants,c s s s s s = supply θ, RF, c θ, RF, c θ, RF, c θ, RF, c e = exhaust

Main Flow pattern Flow pattern Flow pattern Flow pattern characteristics controlled by low controlled by controlled partly controlled by momentum supply buoyancy by buoyancy and high momentum air, strong enough partly by supply air supply air to overcome momentum disturbances Temperature effectiveness and θe- θ s c ce - c s εθ ε  ventilation θoz- θ s coz - c s effectiveness ∞ 1

Figure 3.1. The ideal performance of the total volume room air distribution principles.

Several boundary conditions including room forces (Nielsen 2011). The curve indicates geometry, type and location of supply and the position of the critical Archimedes exhaust air openings, sources and sinks of number where the air movement changes heat including their strength and location, between the two different flow types. enclosure surface temperatures, etc. will influence air distribution in spaces. It can be Δθs very complicated to describe all the details Δθ /q ² large and of the boundary conditions because they are s s convective flow is particular to each room, but a few primary dominating and common conditions and parameters, Δθs /qs² small and which are considered most important, will inlet momentum is be taken into account in the following dis- dominating cussion. These primary variables are: qs • Cooling mode or heating mode; Figure 3.2. Principle determination of airflow 2 • Archimedes ratio ∆θs/qs , or flow rate of in a room with a given ao /A ratio based on the air supplied to the room, qs and tempera- critical Archimedes ratio. Convective flow is ture difference between exhaust and sup- dominant on the left side of the graph while in- ply air, ∆θs; let momentum flow is dominant on the right. • The ratio between the total area of the supply openings and the wall/ceiling/ In cooling mode, the air distribution pattern floor area, a0/A; in a room can be addressed in a three-dimen- • Location, high or low, of the air supply sional graph defined by the flow rate of air opening(s); supplied to the room, qs, the difference be- • Heat load per floor area. tween exhaust and supply air temperature, ∆θs, and the ratio between the total area of The ratio between the total area, ao of the the supply openings and the wall area, ao/A air supply openings and the surface area, A as shown in Figure 3.3 (Nielsen 2011). of wall/ceiling/floor on/in which the supply openings are located, ao/A, is an important The whole “family” of air distribution pat- parameter for the air distribution in the terns can, in the case of cooling, be described room. The ratio, ao/A, is considered to be in two three-dimensional charts, “family small for values smaller than 10-3, medium trees”, one for a high location of the supply for values between 10-3 to 0.3, and large for opening and one for a low location of the air values larger than 0.3. The values smaller supply opening. The charts are shown in than 10-3 are typical for diffusers designed Figure 3.4 and Figure 3.5 (Nielsen 2011). for mixing ventilation, and the value 6·10-3 is typical for displacement ventilation dif- Δθo fusers.

Figure 3.2 shows a design graph (qs − ∆θs graph) for a constant value of ao/A. The area ao /A on the right side of the curve defines mo- qs mentum driven flow while on the left side Figure 3.3. Three-dimensional system which de- defines a flow driven by the buoyancy fines the room air distribution for the cooling case.

A high location of the air supply openings Displacement air distribution can, to some makes it difficult to work with stratifica- extent, be obtained with high level supply tion, and to obtain high air change effiency openings if the openings are large with a when cooling, see Figure 3.4. Most of the low momentum flow and if the heat flow is characterized by strong mixing, ei- sources in the room are located outside the ther due to the high momentum of the sup- downward flow from the openings. The ef- plied airflow or due to the interaction of the fect is seen in the case of vertical ventila- supplied cold flow moving downwards tion, Figure 3.4 (Nielsen et al. 2007). with that of the upward thermal plumes generated by heat sources. Full mixing is typical for this air distribution pattern. In the case of downward flow from a full diffuse ceiling (ao /A = 1,0), with a very high flow rate, it is possible to established piston flow and supply clean air to the working zone. This system is often called a laminar flow system.

Figure 3.5. Different room air distribution systems for cooling with low level supply openings.

3.3 Displacement ventilation and thermal comfort

Thermal comfort is that condition of mind which expresses satisfaction with the thermal environment. Peoples’ thermal sensa- Figure 3.4. Different room air distribution sys- tion is related to the thermal balance of their tems for cooling with high location of supply body as a whole. This balance is influenced openings. by physical activity and clothing, as well as several environmental parameters: air tem- Figure 3.5 shows the location of displace- perature, mean radiant temperature, air ve- ment ventilation in the “family tree” of locity and air humidity. The ranges of envi- room air distribution – which is the subject ronmental parameters for whole body ther- of this guide book. It can be achieved with mal comfort are described in handbooks low level supply openings which makes it and standards (ISO Standard 7730 2005, possible to work with a high ventilation ef- EN15251 2007, ASHRAE Standard 55 fectiveness because of the stratification ef- 2013). Nevertheless, a subject in thermal fect. Displacement ventilation works with comfort can be negatively affected by ad- large supply openings to obtain a low mo- verse local environmental conditions. mentum flow into the room.

In practice occupants may experience lo- Equation (3.2) can be used when air tem- cal thermal discomfort at one or more parts perature, mean velocity and turbulence in- of the body. Local thermal discomfort due tensity at the location of the occupants are to draught, vertical temperature differ- known (obtained by airflow predictions or ence, radiant temperature asymmetry and measurements). a cold/ warm floor may occur alone or in combination. Measurements of these parameters at four heights 0,1, 0,6, 1,1 and 1,7 m above the floor In rooms with displacement air distribution are recommended in the standards (ISO 7726 non-uniformity in the vertical temperature 1998, ASHRAE 55 2013). However, in field may cause local discomfort due to rooms with displacement air distribution the “warm head” and “cool feet” when the dif- highest velocity typically occurs below ference in air temperature between the head 0.1 m. Therefore, measurement at a height of level (1,1 m above the floor) and the ankle 0.05 m above the floor is also recommended level (0,1 m above the floor) is large. The in order to identify the highest velocity. The standards recommend vertical temperature turbulence intensity is typically approxi- difference between 1,1 m and 0,1 m above mately 40 % in case of mixing air distribution the floor to be in the range of 2 K to 4 K and approximately 20 % in case of displace- depending on the category of the aimed in- ment air distribution. Field surveys reveal that door thermal environment (ISO 7730 2005, combined discomfort due to draught and ver- EN 15251 2007, ASHRAE 55 2013). tical temperature difference is not typically a serious problem in rooms with displacement In rooms with displacement air distribution ventilation (Melikov et al. 2005). the relatively low temperature and high velocity near the floor may case draught at the 3.4 Displacement ventilation and feet. Draught is defined as unwanted local air quality cooling of the body due to air movement. The risk of draught increases when airflow The primary aim of ventilation of occupied temperature decreases and mean velocity spaces is to provide people with clean air for and turbulence intensity increase. The per- breathing. In this respect displacement air centage of occupants dissatisfied due to distribution may perform better than mixing draught, DR (%), can be predicted by the air distribution because the free convection following equation (ISO Standard 7730 layer around the human body is less dis- 2005, EN 15251 2007): turbed and its ability to transport clean air from the lower room level to the breathing 0,2 DR (34 a )(  0,v 05 (0,37) v Tu  3,14) zone can be better utilised. This has been (3.2) documented in numerous studies based on CFD predictions and physical measure- In this equation θa [C] is the air tempera- ments performed under laboratory condi- ture, v [m/s] is the mean velocity, and Tu tions. Only contaminant sources with heat [ %] is the turbulence intensity of the flow. production can be treated effectively by dis- The equation is valid when v is higher than placement ventilation (Wildeboer and Mül- 0,05 m/s; for v smaller than 0,05 m/s, v = ler 2006, Cermak and Melikov 2006). How- 0,05 m/s should be used; for DR > 100 %, ever, as will be discussed in the following DR = 100 % should be used.

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 3. ROOM AIR DISTRIBUTION chapters, the ability of displacement air dis- displacement ventilation than in rooms with tribution to provide room occupants with mixing ventilation (Dalewski et al. 2014). clean air for breathing depends on several This has been reported also in field surveys other factors, such as type and location of in rooms with displacement ventilation heat sources, movement of occupants, etc. (Melikov et al. 2005).

Cleanliness of the inhaled air is important The temperature and flow rate of the supply for occupants’ health and perceived air qual- air are important parameters for the design ity. In addition to air cleanliness, tempera- of displacement ventilation. Maintaining ture and relative humidity of the inhaled air relatively high room air temperatures may are also important. Sick building syndrome lead to energy saving. It is reported that at symptoms decrease and perceived air qual- a target air temperature in the occupied ity improves when cleanliness of the inhaled zone (1,1 m height) of 26 °C or 29 °C the air increase and its temperature decrease increase of the flow rate and temperature of (Fang et. al 2004, Melikov and Kaczmar- the air supplied by displacement ventilation czyk 2012). Elevated facial air movement (i.e. small difference between temperature improves perceived air quality and reduces of room and supply air) leads to an im- the negative impact of elevated pollution, provement of the perceived air quality temperature and relative humidity of the in- while a decrease of the flow rate and tem- haled air (Melikov and Kaczmarczyk 2012). perature of the supplied air (large differ- These findings are important for the perfor- ence between temperature of room and sup- mance of displacement air distribution. ply air) leads to improvement of occupants’ thermal comfort (Dalewski et al. 2014). Limited research on human response to the This is discussed further in Chapter 9. environment generated by displacement ventilation is reported in the literature (Car- Perceived air quality (PAQ) is used in the rer et al. 2012). At a comfortable room air standards to define the minimum amount of temperature of 23 °C (1,1 m above floor outdoor air needed for ventilation. Indoor level) and typical indoor pollution sources, air pollution is generated by occupants (bio Sick Building Syndrome (SBS) symptoms effluents) and emissions from building ma- (eye irritation intensity and eye dryness), terials. The total amount of outdoor air re- perceived air quality and thermal comfort quired for ventilation is defined as: were reported by people to be at the same level in rooms ventilated by displacement s p  qAqnq B (3.3) air distribution and those employing mixing air distribution (Dalewski et al. 2014). where qs = total ventilation rate for the breathing Compared to mixing air distribution the zone, m³/s; positive impact of inhaling clean air on per- n = design value for the number of the ceived air quality in the case of displace- people in the room; ment ventilation may be diminished by the qp = ventilation rate for occupancy per high temperature of the inhaled air mainly person, m³/s per person; originated from the free convection layer A = room floor area, m²; around the body. Due to the low velocity air qB = ventilation rate for emissions from may be perceived less fresh in rooms with building, m³/(s m²).

Because displacement air distribution is If the displacement ventilation is not suffi- considered to be more efficient than mix- cient to remove heat generated in the ing air distribution in providing clean air room, it may be combined for example, to the breathing zone of occupants, the to- with radiant cooling (see Chapter 7). tal amount of outdoor air may be reduced.

4 Performance of displacement ventilation

4.1 Displacement Ventilation ventilation flow rate determine the height of Method the boundary between the two zones. The sum of the warm convection flow rates to the The air flow pattern in a ventilated room is upper zone minus the downward directed principally divided into two types, mixing flow rates from cold surfaces to the lower (dilution) ventilation and displacement zone is equal to the ventilation air flow rate ventilation. In mixing ventilation, the ven- supplied to the room. An increased ventila- tilation air is supplied in such a way that the tion flow rate at fixed convection flow rates room air is mixed and the contaminant con- thus moves the boundary upwards and a de- centration is the same in the whole room. In creased flow rate moves the boundary down- displacement ventilation, which is the sub- wards. ject of this book, a stratified flow is created using the buoyancy forces in the room. The air quality in the occupied zone is then gen- erally better than with mixing ventilation. The ventilation system supplying the air to the room is not considered in this book, but only the air flow within the room.

Displacement ventilation has for many years Figure 4.1.Schematic illustration of the air flow been used in industrial premises with high that might be found in a room ventilated by dis- thermal loads. Since the mid-80’s it has also placement ventilation. been used more extensively in non-industrial premises, especially in the Scandinavian 4.2 Air flow pattern countries. Displacement ventilation presents the opportunity to improve both the temper- In a displacement ventilated room, the air ature effectiveness (Chapter 4.3.3) and the flow pattern is governed by the convection ventilation effectiveness (Chapter 4.6). The flows from heat sources and sinks present in principle is based on air density differences the room. This means that a distinctive fea- where the room air separates into two layers, ture of displacement ventilation is the for- an upper polluted zone and a lower clean mation of horizontal air layers. The warmest zone (Figure 4.1). As already discussed in air layers are at the top and the coolest air Chapter 1 this is achieved by supplying cool layers at the bottom. The air moves easily air with a low velocity in the lower zone and within a horizontal layer but the transporta- extracting the air in the upper zone. Free tion between the layers needs a stronger convection from heat sources creates verti- force (Figure 4.2). This means that the ex- cal air movement in the room. When the heat tract should be positioned at the layer in sources in the room are also the contamina- which the pollution concentrations are high- tion sources, the convection flows transport est or where the highest temperatures occur. the warm polluted air up to the upper zone. In most cases this means that the extract The convection flow rates relative to the should be in the upper part of the room.

The vertical air movement is caused by pollution concentration will increase. A convection flows from warm sources or person walking close to the air supply dif- cold sinks. Warm objects such as people, fuser will cause more disturbance than a computers, lamps etc. create rising convec- walking person more distant from the dif- tion flows. Depending on the power and ge- fuser. However, the displacement airflow ometry of the heat source the convection pattern will recover in a relatively short flows will rise all the way to the ceiling or time (Halvonova and Melikov 2010). settle at a lower height (Figure 4.3).

Figure 4.4. Short-circuit of air flow in a room Figure 4.2. Horizontal air movement in con- when the supply air temperature is warmer than nection with the extract. the room air temperature.

4.3 Temperature distribution

As already discussed in Chapter 1, the risk of draught at the feet and discomfort due to vertical temperature difference exist in rooms with displacement ventilation be- Figure 4.3. Vertical air movement caused by cause the cold supply air (sis released at convection. low level directly to the occupied zone (Figure 4.5) and warm exhausted air (e) is The supply air temperature must be lower removed at the ceiling level. The room air than the room air temperature, which is nor- temperature (at different heights will mally the case when there is a cooling load in not, however, vary by much in the horizon- the room. If the supply air temperature is tal direction, except close to the diffuser. warmer there will be a short-circuit (Fig- ure 4.4). However, the vertical air flow has a certain amount of entrainment, which causes some circulation in the rest of the room, this 2,5 is sometimes used for heating an empty room 2,0 prior to the time of occupation by means of a 1,5 displacement ventilation system. 1,0 The airflow pattern in rooms with displace- 0,5 Height above floor [m] floor above Height ment ventilation is sensitive to other flows. 0,0 Walking occupants will cause mixing of the 0 0,2 0,4 0,6 0,8 1 1,2 clean and cool air with the polluted and  - s Temperature ratio warm air at the higher level. This will dis-  - se turb the displacement principle. The tem- Figure 4.5. Temperature stratification in a dis- perature of the inhaled air will decrease and placement ventilated room.

4.3.1 Temperature at the floor In Figure 4.6 the dimensionless temperature of the air near the floor is shown as a The temperature of the supply air in the function of the ventilation flow rate per m² floor area rises due to induction and con- floor area. The points shown in the figure vection, as radiation from the other warmer are from measurements with distributed surfaces in the room in turn heat the floor. heat sources presented in eleven different A dimensionless temperature of the air near references (Mundt, 1996). the floor is often presented as ) s θ 1,0 

af s − α = 5 W/m²K

e cf   (4.1) θ 0,8  se

) / ( αcf= 3 W/m²K s where: θ 0,6 −

f = the air temperature near the floor af θ

( 0,4 s= the supply air temperature

e = the exhaust air temperature κ = 0,2

0 The total temperature difference together 0 1 2 3 4 5 6 7 8 with the air volume flow rate gives the amount of heat removed from the space: Ventilation flow rate per m² floor area, q / A [x10−3 m3/s m²] s Figure 4.6. Dimensionless temperature of the Φ cq   sepstot  (4.2) air near the floor as a function of the ventilation flow rate per m² floor area with different heat where: transfer coefficients due to convection. tot = the heat removed from the space [W] qs= supply air volume flow [m³/s] = the air density = 1,2 kg/m³ cp= the specific heat of the air = 4.3.2 Vertical temperature distribution 1004 J/kg K The vertical temperature distribution de- The following equation can be used to esti- pends on the vertical location of the heat mate the dimensionless temperature of the sources. When the heat sources are in the air near the floor (Mundt 1990): lower part of the room the temperature gradient is larger in the lower part and the tem- 1 perature more constant in the upper part. On

  the other hand, when the heat sources are lo-   cq  11  ps    1 (4.3) cated mostly in the upper zone, the tempera- A     r cf  ture gradient is smaller in the lower part and where increases in the upper part (Figure 4.7). A = the floor area [m²] r= the heat transfer coefficient due to The type and location of the source has a radiation [≈ 5 W/m²K] significant effect on the relative temperature difference (Figure 4.7). Point sources cf = the heat transfer coefficient at the floor due to convection [≈ 4 W/m²K] and horizontal sources (warm floor) create a clear mixing layer.

Over the mixing layer, the room air temper- When people are the primary heat source  ature could be assumed to be constant or to will have a value of 0,58, and evenly dis- only slightly increase. Vertical heated tributed heat sources will give a value of sources (for example, a warm window) pro- 0,65. It is obvious that this can vary in the duce a nearly linear profile with no clear same magnitude as that associated with dif- mixing layer (Figure 1.5). In rooms where ferent flow rates. the heat sources are located at a high level, displacement ventilation is efficient for Heat sources in the Heat sources in the keeping the occupied spaces cool (Fig- upper part of the lower part of the ure 4.8). The air temperatures near the room room floor, f, and the vertical temperature gradient are not only a function of flow rate and load, they are also a function of the type 2,5 of heat source in the room. 2,0 1,5 According to Nielsen (1996) and Brohus and Ryberg (1999) the relative air tempera- 1,0 ture near the floor,  (see equation 4.1) var- 0,5

ies between 0,3 and 0,65 for different types [m] floor above Height 0,0 of heat sources (Figure 4.9). 0 0,2 0,4 0,6 0,8 1 1,2

 -  A concentrated heat load such as a small Temperature ratio s  -  furnace in an industrial environment can e s give a  value of 0,3. A ceiling light will Figure 4.7. Relative change in the vertical tem- give a vertical temperature gradient with a perature in a displacement ventilated room with floor temperature of  = 0,5, which is gen- the heat sources at different levels. erated by radiation from the light source.

Height above floor

Temperature Figure 4.8. Roof heated by sun - an example where displacement ventilation is efficient.

The different temperature gradients are most practical purposes, temperature pro- shown in Figure 4.9 where it is assumed files are assumed as shown in Figure 4.10. that the vertical temperature distribution is a linear function of the height. If many dif- Extract air ferent heat sources are present in the room temperature, θe it is advised to use the “50 % rule” (Chapter 50% 50% 4.3.4). In real situations, vertical temperature stratification is often non-linear.

Distributed heat sources

Sedentary persons Height above floor Ceiling light Point heat source Temperature Supply air Air temperature

Height above floor above Height temperature, θ at floor, θ s af 0 0,3 0,5 0,65 1 Figure 4.10. The "50 %-rule" for vertical tem- 0,58 perature distribution.  -  Temperature ratio s  - se Figure 4.9. Vertical temperature distribution The “50 %-rule” for the vertical tempera- for different types of heat loads with assumption ture distribution indicates that the air tem- of linear vertical temperature distribution. perature at the floor is half-way between the supply air temperature and the extract air 4.3.3 Temperature effectiveness temperature. This is a general experience that may be used as a first approximation As the exhaust temperature is higher than for most normal rooms and normal air dif- the air temperature in the occupied zone, a fusers. temperature effectiveness can be defined: Example:  se   (4.4)  If the heat balance and air flow rate in the oz s room yields a temperature increase of

 −  10 K, then the temperature at where e s the floor level will become approxi-  = the mean air temperature in the occu- oz mately 5 K higher than the supply air pied zone temperature.

4.3.4 Simplified assumptions for the temperature distribution In rooms with higher ceilings than normal, it is often found that the temperature in- As shown in Figure 4.5 and Figure 4.7, the crease from supply air temperature to that temperature increases with height, and the of the air at the floor is less than 50 % of the temperature profile depends on the location total temperature increase. In these cases, a of the heat sources and the flow rate. For “33 % rule” may be appropriate.

4.4 Convection flows – the engines of displacement ventilation

Natural convection flows are the engines of displacement ventilation. A natural convection flow is the air current that rises above warm objects like people or computers, rises along a warm wall, or descends from cold objects like windows or outer walls, due to buoyancy (Figures 4.11 - 4.13). To Figure 4.11. Convection flows - the engine of understand displacement ventilation, one displacement ventilation. has to understand the nature of the natural convection flows, and to know the magni- Flow q tude of these flows. v

The convection flow rising above a hot object, including the human body, is called a     thermal plume, or simply a plume. Empiri- su su cal, analytical and computational fluid dynamics are commonly used methods to evaluate air temperatures, velocities and air Flow flow rates in thermal plumes above differ- qv ent heat sources and convection flows at vertical surfaces. Hot wall Cold wall  >    su su < All plumes encountered in practical venti- Figure 4.12. Convection flows at vertical sur- lation are turbulent flows, and follow the faces. similarity laws for fully turbulent flows.

The amount of air in the convection flow Flow increases with height due to entrainment of qv the surrounding air. The amount of air transported in a natural convection flow depends on the temperature and the geometry of the source and the temperature of the surrounding air. As the driving force in convection flows is the buoyancy force caused by the density difference (i.e. the tempera- z ture difference) a temperature gradient in the room influences the plume rise height.

With development of low power consuming office equipment, lighting, high quality windows, etc., which generate weak buoy- Figure 4.13. Thermal plume above a horizontal ancy flows, the importance of the natural source.

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 4. PERFORMANCE OF DISPLACEMENT VENTILATION convection flow around the human body, point and line heat sources with given heat especially in rooms with displacement air loads were derived based on the momentum distribution, will increase. Apart from the and energy conservation equations and as- impact on the room air distribution (due to suming Gaussian velocity and excessive the generated thermal plume) the free con- temperature distribution in thermal plume vection flow transports pollution generated cross-sections (Mundt, 1996). by the human body and in its surroundings to the breathing zone and therefore is important for occupant exposure and inhaled air quality.

In the comfortable temperature range, the maximum velocity in the convective boundary layer (CBL) around the human body may be as high as 0,25 – 0,30 m/s. It decreases when the difference between the body surface temperature and the surround- Point source Line source ing air temperature decreases (Licina et al. 2014, 2015, 2015a, 2015b, 2016). The ve- Figure 4.14. Plumes from a point source and locity and temperature distribution in the from a line source. CBL, as well as the thickness of the boundary layer is influenced by numerous factors These equations correspond with those pro- including body posture, clothing style and duced experimentally by other researchers thermal resistance, presence of obstacles in (Mierzwinski, 1981, Popiolek, 1981) and the vicinity of the body (such as a desk that are listed in Table 4.1. The equations in greatly reduces the strength of the natural Table 4.1 were derived assuming that the convection flow). (Licina et al. 2014). size of the heat source was very small and Breathing also influences the natural con- did not account for the actual source dimen- vection flow (Özcan et al. 2003, 2005). The sions. natural convection flow and its importance for human thermal comfort and inhaled air The coefficients in the equations differ quality is discussed in Chapter 9. slightly in different references depending on the entrainment coefficients used.  cf is 4.4.1 Point and line sources the convective heat flux in W or W/m from the heat source and z is the height above the Thermal plumes above point and line level of the heat source. The convective sources (Figure 4.14) have been studied for heat flux cf can be estimated from the en- many years. Among the earliest publica- ergy consumption of the heat source tot by tions are those from Zeldovich (1937) and Schmidt (1941). Turner (1973) gives a  cf k  tot (4.5) comprehensive record of most of the phe- nomena encountered in connection with The value of the coefficient k is 0,7–0,9 for buoyancy effects in fluids. Analytical equa- pipes and ducts, 0,4–0,6 for smaller compo- tions to calculate velocities, temperatures nents and 0,3–0,5 for larger machines and and air flow rates in thermal plumes over components (Nielsen, 1993).

4.4.2 Convection flow along vertical 4.4.3 Extended sources and horizontal surfaces In reality heat sources are seldom a point, a Convection flow along vertical surfaces is line or a plane vertical surface. The most also of significance. When the vertical ex- common approach to account for the real tension of the surface is small, the convec- source dimensions is to use a virtual source tion flow is mainly laminar and at larger di- from which the air flow rates are calculated mensions the flow is turbulent. The basic (Mundt 1992 and Skistad 1994) (Fig- equations for a surface with a constant tem- ure 4.15). The virtual origin is located perature are given in Table 4.2 (Jaluria along the plume axis at a distance z0 on the 1980, Etheridge and Sandberg1996). other side of the real source surface.

 is the temperature difference between the surface and the surrounding air and z is the height from the bottom of the surface. Flow The flow changes from laminar to turbulent qv at Gr·Pr=7·108, which for air and moderate temperature differences means around z = 1 m and for air at higher temperatures around z = 0,5 m. z Convection flows from horizontal surfaces are very difficult to determine in the same basic way as for point, line or vertical z0 sources. The reason is that the flows behave Virtual in a very unstable way and leave the flat source surface from different positions at different times, partly depending on the total air a) Point source b) Extended source movement in the room. These surfaces are mostly treated as plumes from extended Figure 4.15. Illustration of the position of the sources see Chapter 4.4.3. virtual source.

Table 4.1. Characteristics of thermal plumes above point and line sources. Parameter Point source Line source 1/3 – 1/3 1/3 Centreline velocity, vz [m/s] vz = 0,128  cf z vz = 0,067  cf 2/3 – 5/3 2/3 –1 Centreline excessive temperature, z [K] z = 0,329  cf z z = 0,094  cf z 1/3 5/3 1/3 Air flow rate, qv,z [m³/s for point source, m³/sm for line source] qv,z = 0,005  cf z qv,z = 0,013  cf z

Table 4.2. Characteristics of convection flows along vertical surfaces. Parameter Laminar region Turbulent region

Maximum velocity, vz [m/s] vz 10   z, vz 10   z, Thickness of boundary layer  [m] 0,05  0 25  z0,, 25 011   z, ,, 7010 0 25 0,, 75 ,, 2140 Air flow rate, qv,z [m³/sm width] z,v 0,q 00287  z z,v ,q 002750   z

The adjustment of the point source model For a flat heat source Morton et al. (1956) to realistic sources using the virtual source suggest the position of the virtual source to method gives a reasonable estimate of the be located at z0 = 1,7 – 2,1·D below the real air flow rate in thermal plumes. source. Mundt (1996) calculates the thickness of the boundary layer (see Table 4.2) The weak part of this method is how to es- at the top of a vertical extended heat source timate the location of the virtual located and adds this to the source radii and then point source. The method of a "maximum calculates the position of the virtual source case" and a "minimum case" provides a tool as z0 = 2,1·(D+2·) before using the point for such estimation (Figure 4.16) (Skistad source equation. According to Bach et al. 1994). According to the "maximum case", (1993), the volume flow from the vertical the real source is replaced by the point surfaces should be added to the volume source such that the border of the plume flow calculated by the equations for point above the point source passes through the or line sources. top edge of the real source (for example, a cylinder). The "minimum case" is when the Example: diameter of vena contracta of the plume is about 80 % of the upper surface diameter Calculate the convection flow rate 0,5 m and is located approximately 1/3 diameter above a cylinder with height 1 m and diameter above the source. The spreading angle of 0,4 m. The convective heat flux is 50 W. the plume is set to 25º. For the low-temperature sources, Skistad (1994) recommends For the maximum case (Figure 4.16) the "maximum case", whereas the "mini- o tan2( 12 2)5 255  ,D,,/Dz 90 m mum case" best fits the measurements for and larger, high temperature sources. The  ,,,hzz 415090 m “maximum case” gives z0 = 2,3·D and the 0 “minimum case” z0 = 1,8·D with z0 defined and from Table 4.1 in Figure 4.16. / 3531 z,v 0 005 cf  z,q which gives 3531 z,v 0050 50  ,,,q m³/s032041 In the minimum case (Figure 4.16) d d h 0 h 0 D/3 o tan2(80 12  08041)5 ,D,,/D,z 72 m z z and z z 0 0 0 03 72 013  150 ,,,,hDzz 09 m H H which gives 3531 D D z,v 0050 50 109  ,,,q m³/s0210 (The position of the virtual source is in this Maximum Minimum case 1,804  131 47 D,D below the up- case case Figure 4.16. Convection flow above a vertical per edge of the source.) cylinder.

4.4.4 Plume interaction When several heat sources are positioned close to each other the plumes merge into a When a heat source is located close to a wall, single plume (Figure 4.17b). The total the plume may be attached to the wall, Fig- flow from N identical sources is then given ure 4.17a. In this case the entrainment will by (Nielsen, 1993) be reduced compared to the entrainment in a 31 free plume. The air flow rate from a heat N,z,v  qNq z,v (4.8) source can then be calculated as half of the flow from a source with a heat emission of where 2cf (Nielsen, 1993). See equation (4.6). qv, z = the flow in the plume from one of the sources

3531 0,005 2( Φcf )  z 351/3 When the heat sources are more separated q  0 0032 Φ z, v,z 2 cf the total flow is equal to the sum of the (4.6) flows from each heat source.

4.4.5 Plumes and temperature gradi-

ents

When there is temperature stratification in a room, like in a room ventilated by displacement ventilation, the plumes are influenced by the temperature stratification. The driving force for the plume is the temperature difference between the plume and the surroundings and when this difference dimin- ishes the plumes will disintegrate and spread horizontally in the room (Figure 4.18). The individual plumes rise to particular levels as plume 1 and 2 in Figure 4.18. The total ef- a) Plume attached b) Interaction between fect on the temperature gradient is the com- to a wall two plumes bination of the various heat loads. Figure 4.17. Thermal plumes.

Plume Plume 3 1 If the heat source is located in a corner the air flow rate is equal to 40 % of the air flow  from a heat source with a heat emission of Plume 1 4 cf (Nielsen, 1993). Plume  2 Plume 2  0020  Φ z,q 3531 Room z,v (4.7) Figure 4.18. Schematic illustration of the air flow pattern in a room ventilated by displacement.

Batchelor (1954) noticed the influence of a Point source temperature gradient surrounding the plume and Morton et al (1956) gave a solu- The position of the virtual source is calcu- tion for calculating the maximum plume lated. A dimensionless height z* above the rise from a point source in surroundings virtual source is calculated with a temperature gradient. The volume flow rates in plumes in a room with temper- * /  / 4183 2 86 sz,z  Φcf (4.9) ature stratification is slightly decreased compared to the volume flow rates calculated with the equations presented for a where: non-stratified media, Mundt (1992). Jin s = vertical temperature gradient (1993) studied the maximum plume rise (/z) in the room [K/m] height for plumes above welding arcs.  cf = convective heat from the source [W] In the presence of a temperature gradient, the convective plume reaches the equilib- As can be seen from Figure 4.19, only z* rium height (zt) where the temperature difference between the plume and the ambient values less than 2,1 are relevant to further calculations. The volume flow rate at the air disappears, see Figure 4.19. Also, there * is another level in the plume, where the air height z is then given by velocity equals zero. This is referred to as 0,q 00238 Φ 43  85  Zs the maximum height of the plume (zmax ). v cf 1 with Point Line * *2 *3 source source 1  0620380003900040 z,z,z,,Z d z* z** (4.10) s = > 0 dz 2,8 2,95 where: q = the volume flow rate in m³/s. 2,1 2,0 v

zmax zt The maximum height zmax is given by Equa- tion (4.9) for z* = 2,8

41  83 max 0 98 Φcf s,z (4.11) Figure 4.19. Vertical plume in a room with temperature gradients and stratification. and the height zt by Equation (4.9) for z* = 2,1 The plume spreads horizontally between these two heights. The convective flow be- 41  83 t 0 74 Φcf s,z (4.12) low zt can be calculated from the following model (Mundt, 1996).

Line source objects found in non-industrial environments (Figure 4.20). The line drawn in the The position of the virtual source is calcu- upper figure is calculated by the equation lated. A dimensionless height z** above the for the air flow rate in Table 4.1. The con- virtual source is calculated vection flow above a sitting person is thus approximately 0,02 m³/s (Figure 4.21). In ** /  / 3121 order to keep the inhaled air at a lower pol- 5 78 sz,z  Φcf (4.13) lution concentration than the surrounding air at the same level, a lower air flow may where: however be used in calculations, see Chap- s = vertical temperature gradient ter 4.6. (/z in the room [K/m]  cf = convective heat from the 200 Equation, Table 4.1

source [W/m] m³/s] -3 ** 100

As can be seen from Figure 4.19, only z [x10

vz 80 values less than 2,0 are relevant to further q calculations. The volume flow rate at the 50 height z** is then given by Vertical temp gradient 30 s = 0,3 K/m ,q 004820 Φ 32  21  Zs l,v cf 2 20 s = 0,09 K/m with ** ** 2 ** 3 2 0 004 0 477  0 029  0 018z,z,z,,Z rate, Convection flow 10 (4.14) 1,0 2,0 3,0 4,0 5,0 Height above object, z [m] where 80 qv, 1 = the volume flow rate in m³/(s m) Personal m³/s] -3 50 computer The maximum height zmax is given by Equa- 75W

** [x10 tion (4.13) for z =2,95. vz 30 q Fluorescent / 31  / 21 lamp 36 W max 0 51 Φcf s,z (4.15) 10 and the height zt by Equation (4.13) for z**=2,0. Desk 5 lamp

Convection rate, flow 60 W 0 35 Φ 31 s,z  21 (4.16) 3 t cf 0,3 0,5 1,0 1,2 1,4

Height above object, z [m] 4.4.6 Plumes from real objects Figure 4.20. Convection volume flow at nor- From the theories above and practical ex- mal room temperatures above a sedentary per- periments, Nielsen (1993) has summarised son, upper figure and above some objects. the convection flows above some common (Mundt, 1992/Nielsen, 1993).

The convective flow above the human body, referred as thermal plume, is influenced by body posture, surrounding air temperature and its stratification, design of clothing (tight or loose) and its thermal insulation, furniture design for example, chair design and desk positioning relative to the seated human body, etc. (Homma and Ya- kiyama 1988, Zukowska 2011a, Zukowska et al. 2010a and b, 2011b, 2012a and b). A standing person generates a symmetrical thermal plume which is relatively easy to characterize. However, for a sitting person the boundary layer develops asymmetri- cally due to the impact of the thermal flow rising from the thighs and lower legs (Fig- ure 4.22). The characteristics of an asymmetrical thermal plume above a sitting human body (room occupant) can be accurately calculated (Zukowska et al. 2010b). The normal height of the ceiling in rooms is often insufficient to allow full development of the plume and the formation of symmetrical profiles of air temperature and velocity distribution.

2,5 m]

r [ 2,0 qvz = 0,020 m³/s oo

e fl 1,5

bov 1,0 a s = dθ/dz = 1,5 K/m 0,5

Height 0 Figure 4.22. Maps of temperature excess (K) above room temperature (above) and air velocity (m/s) (below) measured 0,7 m above the Figure 4.21. Convection flow in plume above a head of a sitting thermal manikin resembling sedentary person in a normal environment. room occupant (Zukowska et al. 2010b).

4.5 Contamination distribution However, if the source is too weak, the plume might disintegrate at a lower level The contamination distribution in a dis- and the contaminants will then be trapped placement-ventilated room depends on the at this level (Figure 4.24) and only slowly position of the contamination sources and if transported indirectly by the stronger con- the heat sources are also the contamination vection flows to the upper zone. sources. In the ideal case with warm concentrated contamination sources, all con- A typical situation is the stratification of taminants are transported directly into the human exhalation in a room with a vertical upper zone by the convection flows, see temperature gradient of 0,5 K/m as shown Figure 4.23. According to Krühne and Fitz- in Figure 4.25B (Bjørn and Nielsen 2002). ner (1995), Cermak et al. (2006), Cermak and Melikov (2006) if the contamination The contaminant concentration is also in- sources are cold and evenly distributed at fluenced by the downward directed convec- the floor, the contamination distribution tion flows that might occur at the outer will be like the temperature distribution walls in cold seasons, especially when the (Figure 4.10). walls are poorly insulated.

2,5 [m]

z 2,0

1,5

1,0

0,5 Height above floor, above Height 0 0 0,2 0,4 0,6 0,8 1,0

Contamination ratio, c /c room e Figure 4.23. Schematic illustration of the contamination distribution in a room ventilated by displacement ventilation and with warm contaminant sources.

2,5 [m] z 2,0 croom

1,5

1,0 plume 2  0,5 plume 1 room Height above floor, floor, above Height 0

Contamination, croom Temperature,  Figure 4.24. Schematic illustration of the contamination distribution in a room ventilated by displacement ventilation, when the contaminant source (the person) is not the warmest source.

The influence of a poorly insulated roof will, in the cold season, decrease the concentration gradient, due to the downfall of cold air, just like with the cold walls (Fig- ure 4.26). However, if the roof is heated by the sun this will help stabilise the displacement ventilation as it heats the air in the upper zone (Figure 4.8).

Figure 4.26. Poor building air tightness and insulation may reduce the benefit of displacement ventilation, and make it more like mixing ventilation.

4.6 Ventilation effectiveness

In order to assess and compare different air B distribution patterns different definitions of Figure 4.25. A) Exhalation in surroundings ventilation effectiveness have been intro- with a small vertical temperature gradient, duced. These are discussed in detail in the 0,1 K/m. The exhalation rises to the ceiling. B) Stratified exhalation from a manikin (person) in REHVA’s Guidebook No. 2 on Ventilation a room with a larger vertical temperature gra- Effectiveness (Mundt et al. 2004). In defin- dient, 0,5 K/m. ing ventilation effectiveness, a distinction must be made between two terms:

These downward flows will then transport • the contaminant removal effectiveness, the contaminants from the upper zone back  c, which is a measure of how quickly an to the lower zone (Yamanaka et al. 2007). airborne contaminant is removed from However as long as there is a positive con- the room (Brouns and Waters, 1991) and centration gradient in the room, the contam- • the air change efficiency, a, which is a inant concentration in the occupied zone measure of how quickly the air in the will always be lower than by mixing venti- room is replaced (Sutcliffe, 1990). lation.

In a displacement ventilated room, the air 4.6.1 Contaminant removal effective- change efficiency is mostly higher (a 60– ness 70 %) than in a room ventilated by mixing ventilation (a 50 %), (Mundt, 1994). A The contaminant removal effectiveness is good survey of the relationship between the defined by: different versions of ventilation effectiveness is given by Nielsen (1993). Flows in  cc  c  se rooms can be defined based on the air (4.17)  cc smean change efficiency (Figure 4.27). Perfect mixing ventilation is defined by an air where change efficiency  equal to 50 % and a a c = the contaminant concentration in contaminant removal effectiveness  c equal e the exhaust to 1. Short circuit flows lead to values c = the contaminant concentration in smaller than 50 % for the air change effi- s the supply ciency. The quality of displacement venti- c = the mean contaminant concentra- lation systems depends on the contaminant mean tion in the room source. Only contaminant sources with heat production can be treated effectively by or for the occupied zone: displacement ventilation (Wildeboer and Müller 2006). c  cc   se (4.18) oz  cc s >1

where coz = the mean contaminant concentra- c Short circuit flow, ε

emoval tion in the occupied zone contaminant source Displacement close to exhaust ventilation flow Piston opening 4.6.2 Personal exposure index. effectiveness

Contaminant r Thermal flow around a person and flow generated by the movement of a person may give 1 Air change efficiency εa an inhaled concentration that is different from 0 % 50 % 100 % the concentration at head height if the meas- Perfect urements are made without a person. mixing Displacement ventilation, This can be expressed by the following per- Short circuit flow contaminant sonal exposure index, Brohus and Nielsen source without (1996a): heat generation

c  cc se exp  (4.19) exp  cc s <1 Figure 4.27. Definition of different flow types where based on ventilation effectiveness measures. cexp = the inhaled concentration.

It is possible to work with a stratification z cc st  cc )( z < z height that is lower than the height of the exp p fp st exp (4.20) zexp breathing zone. The personal exposure index will often be larger than the local ven- where z is stratification height, z is tilation index because clean air is moved st exp height to the breathing zone, c the concen- from the lower part of the room up to the p tration at breathing height outside the breathing zone by the free-convection breathing zone and c the concentration in boundary layer around the person (Fig- f the lower zone (floor level). c is equal to ure 4.28 and 4.29). exp cp when zst > zexp (Figure 4.30).

z cexp cp zexp

zst

cf cR c

Figure 4.30. Inhalation of air in a room with stratified flow and stratification layer below the Figure 4.28. Thermal flow around a person breathing zone. may give cleaner breathing air. The transport of clean air in the personal boundary layer can only take place when people are not moving. Figure 4.31 shows that the effect disappears when a person is moving with a speed equal to, or greater than, 0,2 m/s in a room with a stratified layer (concentration cp at head height) (Bjørn and Nielsen 2002).

cexp / cp 1,4

1,2 Figure 4.29. Iso-concentration map showing 1,0 the dispersion pattern of a tracer gas emitted directly above a 4 W heat source in the lower 0,8 zone. The dummies are situated in the measur- 0,6 ing plane (Stymne et al, 1991). 0,4 0,2 Usually the stratification height will be 0,0 around 1 m in a room when the air distribu- 0,0 0,2 0,4 0,6 0,8 1,0 Speed [m/s] tion is designed for an appropriate temperature distribution. The concentration in the Figure 4.31.Concentration in inhalation cexp inhalation air cexp of a standing person can relative to concentration in front of the breath- be found from (Brohus and Nielsen 1996a). ing zone cp versus speed of movement.

Moving people will also influence the ver- Measurements of the personal exposure in- tical contaminant distribution as seen in dex made in situations with air movement Figure 4.32. The figure shows the relative in the occupied zone and contaminant concentration distribution for four seden- sources close to a person can give rise to a tary people and for two sedentary people very small exposure index (Brohus and and two people in motion. It is shown that Nielsen 1996b). two people in motion are able to smooth the vertical gradient slightly, but it is possible Displacement ventilation should not be to observe a stratification of CO2 (Nielsen used when the contamination sources are 1992a and Brohus and Nielsen 1994). Hal- mostly cold. vonova and Melikov (2010) also reported that in a room with displacement ventila- As pointed out above, the ventilation flow tion walking people will destroy the strati- rate should not be set equal to the convection fication and will decrease the quality of air flows above the people present in a room, inhaled by seated room occupants. The dis- because this will, in practice, lead to too high turbance decreases with the distance be- air flow rates. Figure 4.33 shows the im- tween the walking person(s) and the air provement in inhaled air quality relative to supply diffuser(s). the air quality in the ambient as a function of the ventilation flow rate per person.

Figure 4.32. Concentration distribution in a room with thermal manikins, sedentary people and people in motion. Figure 4.33. The ratio between the concentration in the breathing zone and in the ambient air Transport of air from the lower zone to the at the same height (Etheridge and Sandberg, breathing zone is an additional positive ef- 1996). fect in displacement ventilation, but in the case of movement the lack of air from the With a ventilation flow rate of lower zone means that the concentration in 0,020 m³/(s·person), the boundary is above the inhalation air corresponds to the fully the person. However, a ventilation flow rate mixed concentration which will be found in of 0,010 m³/(s·person) gives a concentra- rooms with mixing ventilation. tion of only 20 % of the concentration in the

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 4. PERFORMANCE OF DISPLACEMENT VENTILATION ambient at the same level. Figure 4.33 ap- to be little risk of resuspension of particles plies when the breathing zone is in the up- from the floor into the supply air flow. The per layer, i.e. when the sum of the convec- sizes studied were however only particles tion flow rate is larger than the ventilation larger than 0,5 μm and more research is flow rate. In these cases, the concentration needed for smaller particles. Recent studies in the upper layer is almost equal to the ex- (Rim and Novoselac 2009, Licina et al. haust concentration and the reduction in 2015a) reveal that the human convective concentration in the inhaled air can be cal- boundary layer transports particles of small culated from Figure 4.33. Measurements size and substantially influences the per- by Mundt (1994) showed the rapid, almost sonal exposure when the pollution origi- instantaneous, recreation of the thermal nates at the low level, for example, foot flow around a person when the person level. With stratified airflow patterns, such moves from one place to another in a room. as displacement ventilation the inhaled concentration of particles generated at floor Particle transportation in a displacement- level and the near proximity to an occupant ventilated room was studied by Mundt may be several times higher than the ambi- (2001). The results indicate that there seem ent concentrations.

5 Calculation of supply airflow rate

Two principal methods can be used when In the case of steady-state conditions when the supply air flow rate of displacement the supply air flow rate in the room is de- ventilation system is calculated: 1) temper- creased, the vertical temperature stratifica- ature based design, where the design crite- tion in the occupied zone and the ceiling rion is the air temperature in the occupied temperature will increase. This implies that zone of the room and 2) air quality based the thermal radiation from the upper zone to design where the design criterion is the air the lower zone will also increase and so will quality in the occupied zone. increase the air temperature at the floor level.

In Chapter 4.4, the thermal plume equations In a real building, the thermal mass of the are presented that are applied in the air building also influences the room air tem- quality based design. Based on the known perature. The resulting temperature will be heat loads and their locations, the supply air dependent on the thermostat location and flow rate is set by calculating the air flow the selected control strategy, i.e. variable rate induced by convection flows. airflow volume (VAV) strategy or constant airflow volume (CAV) strategy. In commercial buildings, the removal of the excess heat is likely to be the main concern. The displacement ventilation design meth- The cases where cooling is the main issue, ods applicable for manual calculations are the temperature based design is the most based on empirical coefficients and nomo- commonly applied method. For that reason, grams, in which the influence of thermal ra- this chapter focuses only on temperature diation exchange between the upper and based design. lower parts of the room is included (Skistad 1994, Halton Oy Design Guide 2000, etc.). 5.1 Temperature based design The advantage of these methods is their methods ease of use and also the accuracy of the estimation of temperature and contaminant In the design process, the challenging task distribution particularly in industrial type is to estimate vertical contaminant or tem- applications. perature gradients in the room space. While the contaminant stratification level is In modern practice, it is more common to mainly affected by the relation of supply air use simulation software where the contam- flow rate and convective air flow rate, ther- inant and temperature gradients are mod- mal stratification is also affected by thermal elled. In some models, the temperature dis- radiation exchange between different room tribution is modelled to be linear over the surfaces. The thermal radiation from upper room height, i.e. a constant vertical temper- level surfaces warms lower level surfaces ature gradient is assumed (Mundt 1996, and thus affects the air temperature at floor Arens 2000, Nielsen 1995, 2003). Temper- level and in the occupied zone. ature based design methodology where the

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 5. CALCULATION OF SUPPLY AIRFLOW RATE space is divided into zones: the lower occu- Some nodal models are currently available pied zone and the upper unoccupied zone is in energy simulation tools. Rees’ model can introduced by Livchak and Nall (2001). be used with ESP-r (Hensen and Hamelinck The heat loads are split into two zones ac- 1995). Mundt (1996) and da Graça (2003) cording to their type and location. The radi- models are implemented in DOE Ener- ation between upper and lower zones is also gyPlus (2015). Mundt (1996) and Erikson taken into account when the air tempera- et al. (2012) models are incorporated in tures of the zones are iteratively solved. IDA ICE (Sahlin 1996).

An approach to determine the required air 5.2 Calculation of vertical room flow rate and the supply temperature by use air temperature distribution of fractional coefficients applied for three selected heat loads is suggested (Chen et al. A heat-balance-based method is used when 1999, Chen and Gliksman 2003). The frac- excess heat is considered the main indoor tional coefficients set the ratio of the con- climate concern. In rooms with displace- vective heat load that is released in the ment ventilation the vertical temperature room space between head and foot level. distribution depends on the output, characteristics and location of the heat sources and Several nodal models have been introduced on the airflow rate. that allow different slopes for the temperature profile between nodes. Three-node The following chapters introduce three models (Li et al. 1992, da Graça 2003 and models that calculate the vertical tempera- Mateus and da Graça 2015) and multi-node ture distribution in rooms with displace- models have been proposed that apply jet ment ventilation. and thermal flow elements to track individual jets from heat or mass sources (Erikson 5.2.1 Linearized vertical temperature et al. 2012). Multi-zone models where air distribution calculated by flow rates between the nodes are predefined Mundt model by a CFD method has also been proposed (Rees 1998 and Griffith 2002). In this model, the radiative energy transfer from the ceiling to the floor is balanced by The influence of coupling displacement convective heat transfer from the floor sur- ventilation with chilled ceilings or floor face to the air at floor level. heating on the temperature distribution have been analysed (Novoselac et al. 2006, αr·Af·(θe – θf) = αcf·Af·(θf – θaf) (5.1) Rees and Haves 2001 and Rees and Haves 2013). The convective heat transfer from the floor is in turn equal to heat transferred to the Compared with CFD-simulation the nodal supply air near the floor, neglecting any in- models require less computation time, they duction of room air into the supply air flow are more suitable for engineering calcula- upon entering the room: tions and can be added to the whole building simulations. ρ · cp · qs ·(θaf – θs) = αcf·Af·(θf – θaf) (5.2)

Where qs= airflow rate (m³/s).

The dimensionless temperature κ, defined The vertical temperature profile is given by equation 4.1, can be calculated accord- from the floor temperature, Figure 5.1, and ing to equation 4.3 introduced in Chapter 4 the stratification height zt: (Mundt 1996).

e  af   z   for z > zt (5.7) Together with an energy balance (equation z af 5.3), it is possible to estimate vertical room t air temperature profile and to select the re- θ = θe for z > zt (5.8) quired supply air flow rate.

Using equations 5.3 – 5.8 and Figure 5.1, it (5.3) tot = qs · ρ · cp · (θe – θs) is possible to determinate the room air tem-

perature distribution. Using the energy balance equation (5.3) and the dimensionless temperature in equa- Figure 5.2 shows measurements of vertical tion 4.3, the supply air temperature may be temperature distribution in a room with determined: four thermal manikins as heat sources. The predicted temperature profiles are found θs = θaf − · tot /(ρ · cp · qs) (5.4) from Figure 5.1 and equation (5.6). The predictions seem to give an improved de- 5.2.2 Vertical temperature distribu- scription of the vertical temperature distri- tion calculated by Nielsen model bution in comparison with a linear distribution over the entire height of the room. In the Nielsen model (Nielsen 1995 and 2003), a linear temperature gradient be- κ tween floor and the height of mixing layer 0,8 (stratification height) is predicted. A B Over the mixing layer, the room air temper- 0,6 C ature is assumed to be constant. The mixing layer temperature (that is the same as the D 0,4 exhaust air temperature) is calculated with the energy balance (equation 5.3). The floor temperature is determined with specific 0,2 ArA- number of supply air:

0,0 ArA = β·g·H·(θr – θs)/(qs/Af)² (5.5) 402010 60 400200100 -3 ArA.10 The height of the mixing layer z is calcu- t A = Distributed heat source lated: B = Sedentary persons A = Ceiling light 2/5 -3/5 A = Point heat source zt = 0,62·Φcf ·(θe – θaf) (5.6) Figure 5.1. Dimensionless floor temperature where Φcf = total convection load. versus the Archimedes number Nielsen (1995).

1,0 FMO characterizes the fraction of the con- / H z vective heat loads that are mixed into the 0,8 occupied zone, and, therefore, not con- 0,6 veyed directly to the mixed layer (0 < FMO <1). The lower level mixing does 0,4 not occur in an ideal displacement system (F =0). A default value of F can be 0,2 MO MO used 0,4. 0,0 0,0 0,2 0,4 0,6 0,8 1,0 1,2 . 3 ArA=41 10 zt / H = 0,83 θ – θs Ar =28.104 z / H = 0,62 A t θe – θs Figure 5.2. Vertical dimensionless temperature distribution in a room with four thermal manikins, Nielsen (1995).

5.2.3 Vertical temperature distribution calculated by Mateus and da Graça model

The estimation accuracy of the vertical temperature distribution can be improved by modelling based on three-nodes (Mateus and da Graça 2015 and Kosonen et al. (2016). In this way, it is possible to obtain different slopes of the vertical temperature profile between the nodes. Three room air temperatures are predicted: at heights of 0,05 m, and 0,65 m above floor level and at the mixed layer. Figure 5.3. Mateus and da Graça (2015) model Mateus and da Graça model (Figure 5.3) in- scheme. cludes convective energy conservation equations for the three-nodes and radiative energy conservation equations for the room The radiation between the surfaces is mod- surfaces: floor, ceiling and two lateral wall elled by dividing the wall into lower (Awl) areas. and upper (Awu) surface areas. The height of the lower wall area (hTmx) is calculated with In the model, two additional parameters IM the height of ceiling and the height of the and FMO are included. IM describes entrain- mixing layer (equation 5.9): ment generated accumulated flow rate. De- fault value of IM is 0,6. The parameter of hTmx = hmx + (H – hmx)/3 (5.9)

The convective heat balance for the three- the height 0,65 m, θmx is room air tempera- nodes can be set by the following equa- ture at the mixed layer. tions: Without detailed information of heat sources, the breakdown between convec- ρ·cp·qs·(θaf – θs) – IM·ρ·cp·qs·(θoc – θaf) tive and radiative heat loads of 50 % / 50 % can be used. The height (hmx) of the mixing = αcf,f·Af(θf – θaf) (5.10) layer is calculated with the following thermal plume equation where the plumes are assumed to be like a point source and located at the floor level. ρ·cp·qs·(θoc–θaf)+IM·ρ·cp·qs·(θoc–θaf)–Φcf·FMO 3 1/5 hmx = 23,95·(qs /Φcf) (5.13) = αcf,wl·Awl·(θwl– θoc) (5.11)

With the seven linear equation (5.10–5.12 and 5.14–5.17), it is possible to solve for the seven unknown room air nodes and sur- ρ·cp·qs·(θmx – θoc) – Φcf·(1−FMO) = face temperatures.

αcf,c·Ac·(θc – θmx)+ αcf,wu·Awu·(θwu – θmx) (5.12) The mixed layer temperature is assumed to Where Φr,f, Φr,w and Φr,c = radiant heat load be equal to the exhaust temperature. The ra- in on floor, wall and ceiling. αcf,f, αcf,wl, αcf,wu diant heat exchange equations for the room and αcf,c are convective heat transfer coeffi- surfaces are introduced in the room surface cients on floor, lower and higher level walls energy conservation equations, considering and ceiling, αr is radiative heat transfer co- the equal impact of the radiative heat trans- efficients, θaf is room air temperature at the fer to all the surfaces: height 0,1 m, θoc is room air temperature at

αcf,c·(θc – θmx) + αr,c·(θc – (θfAf + θwlAwl + θwuAwu)/(At – Ac)) = Φr /At (5.14)

αcf,f·(θf – θaf) + αr,f·(θf – (θcAc + θwlAwl + θwuAwu)/(At – Af)) = Φr /At (5.15)

αcf,wl·(θwl – θoc) + αr,wl·(θwl – (θcAc + θfAf + θwuAwu)/(At – Awl)) = Φr /At (5.16)

αcf,wu·(θwu – θmx) + αr,wu·(θwu – (θcAc + θfAf + θwlAwl)/(At – Awu)) = Φr /At (5.17)

Where Ac, Af, Awl, Awu and At are areas of ceiling, floor, lower level wall, higher level wall and total area of all surfaces.

5.3 Vertical position of the heat Table 5.1. Input data of calculation examples. source Size of the room: Height H 5,12 m A vertical location of the heat source may also influence the temperature distribution, Width W 4,55 m Figure 5.4A. This effect is especially Length L 4,55 m pronounced when the source may be Heat Transfer Coefficients of the room surfaces: considered as a point heat source with limited radiation. Figure 5.4B shows Convective heat transfer coefficients: measurements of two vertical temperature Ceiling αc,c 1,5 W/(m²∙K) profiles for sources with two different Floor αc,f 4,0 W/(m²∙K) vertical locations (Nielsen 1995). Wall surface below the mixed layer αc,wl 1,5 W/(m²∙K) Wall surface above the mixed layer αc,wu 1,5 W/(m²∙K)

Radiative heat transfer coefficients:

Ceiling αr,c 5,8 W/(m²∙K)

Floor αr,f 5,8 W/(m²∙K) Wall surface below the mixed layer αr,wl 5,8 W/(m²∙K) Wall surface above the mixed layer αr,wu 5,8 W/(m²∙K)

Air properties: Density ρ 1,2 kg/m³

Specific heat capacity cp 1005 J/(kg∙K) Thermal expansion coefficient β 3,43 10-3 1/K Case studies:

Supply air temperature θs 18 °C Figure 5.4. A) Vertical dimensionless position of the heat source in the room. B) Dimension- Air temperature at the less temperature distribution in a room with a height 1,1 m θ1.1 23 °C point heat source located at two different vertical positions, (Nielsen 1995). The room and the combinations of heat loads used in the following examples are 5.4 Calculation examples when the same as described in Chapter 8.2, where using temperature based measured and calculated room air tempera- design models ture profiles are compared.

In this chapter, three models are used in cal- Using temperature based design models of culation examples of three simplified de- Mundt, Nielsen and Mateus and da Graça sign cases where the input data for the room for the fixed room air temperature (23 °C at geometry, heat transfer coefficients, air 1,1 m level), the required air flow rate properties and the supply and target tem- should be calculated by using an iterative peratures are presented in Table 5.1. method.

Calculation example 1: Single buoyant Calculation example 2: Single buoyant flow element of occupants (900 W) flow elements of window (520 W) Heat load of 12 people (simulated by Heat load of 520 W generated in the heated cylinders with height of 1,0 m, di- room by heated window with height of ameter 0,3 m) generating 75 W heat, i.e. 3,6 m and width of 3,6 m is considered. total 900 W, heat load in the room. The window is installed at height of 0,8 m above the floor. Vertical temperature profiles obtained with three models are shown in Figure The vertical temperature distribution in 5.5. The estimated vertical temperature the room obtained with three models is distribution calculated with the three shown in Figure 5.6. In Mundt model, models varies. The vertical air tempera- room air temperature is linearized be- ture distribution between floor and ceil- tween the floor and ceiling while Mateus ing is linearized with the Mundt model, and da Graça and Nielsen models esti- while Mateus and da Graça and Nielsen mated the mixing layer starting respec- models estimate the mixing layer starting tively at 1,6 m and 2,6 m. The required at 1,1 m and 3,0 m respectively. The re- air flow rates were 0,049 m³/s (Mundt), quired air flow rates were 0,073 m³/s 0.06 m³/s (Nielsen) and 0,072 m³/s (Ma- (Mundt), 0,096 m³/s (Nielsen) and teus and da Graça). 0,149 m³/s (Mateus and da Graça).

m 6 m 6 t, t, 5 5 Heigh 4 Heigh 4

3 3

2 2

1 1

0 0 20 22 24 26 28 30 20 22 24 26 28 30 Temperature, °C Temperature, °C

Mundt model q = 0,073 m³/s s Mundt model qs = 0,049 m³/s Nielsen model q = 0,096 m³/s s Nielsen model qs = 0,06 m³/s Mateus, da Graca q = 0,149 m³/s Mateus, da Graca q = 0,072 m³/s s s Figure 5.5. Temperature profiles of case 1. Figure 5.6. Temperature profiles of case 2.

Calculation example 3: Combination of heat loads (1762 W) The vertical air temperature distribution is linearized between floor and ceiling In this case, heat load in the room is re- with the Mundt model while Mateus and sult of: da Graça and Nielsen models estimated the beginning of the mixing layer respec- • Heat load (900 W) of 12 people simu- tively at 2,9 m and 2,2 m. The required lated by 12 heated cylinders (height air flow rates were 0.118 m³/s (Mundt), 1,0 m, width 0,3 m) each generating 0.208 m³/s (Nielsen) and 0.202 m³/s 75 W heat; (Mateus and da Graça). • Heat load (520 W) from window (height 3,6 m and width 3,6 m) installed 0,8 m above the floor; Conclusion of the calculations • Heat load (260 W) from the floor (floor heated area 2,4 m x 2 m); In this chapter, the room air temperature • Heat load (232 W) from fluorescent gradient and required supply air flow rates lighting units installed at ceiling. are calculated using three temperature based design models. Two of the examples The vertical temperature distribution cal- are with single buoyant flow elements (oc- culated with three models is shown in cupant and warm window) and in one of the Figure 5.7. examples a combination of typical heat loads in office is used.

6 m The examples demonstrate that the calcu- t, 5 lated supply air flow rates are quite differ-

Heigh ent with the three models. When the room 4 air temperature is assumed to be linear be- 3 tween the floor and ceiling level, the required air flow rate is lower than that calcu- 2 lated with the models predicting the height 1 of mixing layer. 0 20 22 24 26 28 30 32 In design work, it is recommended to use a Temperature, °C model that calculates the height of the mixing layer. At the moment, the best average

Mundt model qs = 0,118 m³/s accuracy of the simplified models is given by the Nielsen and Mateus and da Graça Nielsen model qs = 0,208 m³/s Mateus, da Graca q = 0,202 m³/s models. However, the accuracy is depend- s ing on the type of the flow element (see Figure 5.7. Temperature profiles of case 3. more in Chapter 8.2).

6 Air diffusers for displacement ventilation

6.1 Commonly used diffusers

There are several types of diffusers used for displacement ventilation. The most commonly used types are integrated in the walls. Other types are placed at the walls or in a corner, free-standing on the floor, or integrated in the floor. The layout of the room should be considered in connection with the selection of the type of air diffusers.

Figure 6.1 shows circular and semi-circular diffuser and Figure 6.2A shows a wall Semi-circular, wall mounted mounted diffuser.

The air is supplied in the lower part of the room and it is therefore convenient to locate the diffusers, near to corridors and other unoccupied areas to obtain an area in front of the diffusers where a higher velocity can be tolerated.

The diffusers in Figure 6.1 will create a radial flow at the floor close to the diffusers because of gravity effect on the cold air leaving the diffusers and partly because of Semi- circular, corner mounted the diffuser design. Plane wall mounted diffusers may also create radial flow at the floor close to the wall. Figure 6.2A and B shows how the cold low velocity air supply at the surface of a diffuser falls towards the floor and creates a radial flow at the floor.

Figure 6.3 shows a wall mounted diffuser with integral nozzles for the adjustment of the supplied flow pattern. The flow close to the diffuser can be directed parallel to the wall with only a small amount of forward directed flow. The diffuser will create a Circular, free standing short “adjacent zone” (discussed in Chapter 6.4.1). Figure 6.1. Semi-circular and circular diffusers.

much flow along the floor. Figure 6.4B shows a carpet diffuser which can cover the whole room. This type of supply generates a very low velocity and in this case the air movement in the room originates from the heat load. However, care should be taken to consider the contamination which could be A emitted from the carpet in this type of supply.

Figure 6.2. A) plane wall mounted diffuser with air velocity directed in the direction perpendicular to the diffuser surface. B) Smoke experiment showing how the flow creates a radial pat- A tern on the floor close to the diffuser although the air is supplied perpendicular to the diffuser surface.

Figure 6.3. Wall mounted low velocity diffuser Figure 6.4. A) Floor mounted diffuser with high with integrated nozzles for the adjustment of the velocity swirl supply. B) Supply of air through flow patterns at the surface. the carpet.

Custom made diffusers integrated into the Displacement diffusers can also be installed room design can also be seen in many situ- above the occupied zone or in the ceiling. ations. They supply cold downward flow locally for example by using a wall surface to sup- With a floor mounted diffuser, there is a port the flow into the occupied zone (Niel- draught area in the flow above the diffuser sen et al. 2010). (Figure 6.4A) but the diffuser does not create

6.2 Radial air flow or plane air flow from low-velocity diffusers

The diffusers described in Chapter 6.1 will create a local flow in close proximity to the diffusers. The flow in the occupied zone of the room is not necessarily controlled by this local flow, but will depend on the num- A ber and location of diffusers and on the room geometry. Two types of flow can be generated, namely (semi) radial flow or plane flow, see Figure 6.5. A radial flow along the floor will take place if the distance between the diffusers is as large as the distance to the opposite wall while a plane flow will take place if the diffusers are lo- v cated close to each other, or the room is x very long and narrow in the direction of flow as indicated in Figure 6.5B. A diffuser with a high sideway discharge, Fig- B ure 6.3, will also create plane flow even if the diffusers are not particularly close to Figure 6.5. A) A diffuser creating radial flow each other (see Figure 6.19). along the floor in the occupied zone due to the use of a single diffuser and due to room geometry. B) A number of diffusers creating plane flow in a room. 6.3 Air flow from low –velocity diffusers

Normally, the supply air is between 3 K and 5 K cooler than the room air. In these case, the supply air falls towards the floor when it leaves the diffuser, and spreads, like a blanket, across the floor.

When the supply airflow is isothermal, i.e. has the same temperature as the surrounding air in the lower part of the room, the flow will be distributed horizontally into the room according to the initial flow pat- Figure 6.6. Isothermal air supply. The flow has tern at the surface of the diffuser (Fig- a constant velocity core as in a large three di- ure 6.6). mensional wall jet.

As already discussed in Chapters 3 and 4 if zone, ln, is defined as the distance from the the supply temperature is higher than the diffuser to the point where the maximum surrounding temperature, the supplied air- velocity has decreased to 0,2 m/s when the flow will rise to the ceiling without spread- temperature difference between the room ing in the occupied zone (Figure 6.7). air (at 1,1 m height) and supply air is 3,0 K. Therefore, displacement ventilation can be used effectively only when the supply air is Draughts result from high mean velocity air cooler than the room air. flow with high turbulence intensity and low air temperature (Chapter 3.3). For a given velocity (for example, 0,2 m/s) the risk of draught will be low when the air temperature is high and conversely, the risk will be high when the temperature is low. There- fore, determination of the near zone based on only the velocity is misleading.

ISO Standard 7730 (2005) suggests three categories A, B and C of thermal environment with corresponding 10, 20 and 30 % Figure 6.7. Supply of warm air. dissatisfied occupants due to draught. In some cases, less stringent requirements The flow indicated in Figure 6.7 may be may be used if a room space is used, for ex- accepted in special cases when the air dis- ample, by moving people. The draught risk tribution system is used for pre-heating is calculated based on measured or pre- spaces in periods when unoccupied. This dicted local mean velocity, turbulence in- could be night heating in, for example, of- tensity and air temperature. Figure 6.8 fices or pre-heating of a concert hall. shows an example of an adjacent zone de- Asymmetrical heating or cooling loads may fined based on measurements of velocity, give some recirculation in the room which turbulence intensity and temperature near will mix the air in the case of warm air sup- the floor in a room with displacement ven- ply. tilation. 6.4 Air distribution from a low- velocity diffuser giving a radial flow in the occupied zone

6.4.1 The “Adjacent Zone”

The air from a single wall-mounted diffuser flows over the floor and generates radial flow. Close to the air supply diffuser there is a zone where the flow has relatively high velocity and low temperature. In this zone the risk of draught may increase. This zone Figure 6.8. Adjacent zone defined by velocity of is called “adjacent zone”. It is commonly 0,2 m/s and by 10, 20 and 30 % dissatisfied oc- accepted that the length of the adjacent cupants (Melikov and Langkilde 1990).

The adjacent zone based on the draught risk In practice, the number of diffusers and the (DR) of 10, 20 and 30 % is shown. The com- diffuser design, room dimensions and fur- parison shows that the adjacent zone defined nishing may give a semi radial flow with a by the velocity of 0,2 m/s does not comply higher velocity and longer adjacent zone with the requirements in the standard and than expected for the given diffuser type in there is risk more than 30 % of the occupants the radial flow case, see Chapter 6.4.5 will be dissatisfied due to draught. “Flow between obstacles” and (Melikov and Langkilde 1990, Nielsen et al. 2004). From Figure 6.8, it becomes clear that the length of the adjacent zone will depend on Avoiding draught is the major challenge in the category of the designed thermal envi- developing low velocity air diffusers. To ronment. The same category, i.e. the same reduce the size of the adjacent zone, the level of draught risk, can be achieved by number of diffusers in the room or diffuser different combinations of velocity, turbu- face area should be increased. This also lence intensity and air temperature. This is generates a more homogenous indoor envi- shown in Table 6.1. ronment in the occupied zone. Different diffuser types may be used. Typically, As discussed in Chapter 6.2, if the diffus- those that supply air in only one direction ers are located close to each other or the will generate a longer adjacent zone than room is very long and narrow in the direc- those distributing supply air from semi-cir- tion of flow a plane flow will be generated cular or circular diffusers. One way to re- (Figure 6.5B). In this case the flow veloc- duce the draught in the occupied zone is to ity does not decrease with the distance direct the supply air sideways to the wall from the diffuser (see Chapter 6.5) and it outside the occupied zone. Figure 6.9 is impossible to define an adjacent zone. shows two typical adjacent zones estab- Therefore, it is more correct to relate the lished with forward discharged flow and required comfort to the draught rating with sideway discharged flow. model discussed in Chapter 3.3.

Table 6.1. Adjacent zone of the same category can be achieved by different combinations of local air temperature and velocity of the supplied flow.

Draught rate requirement (1 Local airflow temperature[°C] Local airflow velocity[m/s] 19 0.11 Category A: 10 % 21 0.12 23 0.14 19 0.19 Category B: 20 % 21 0.21 23 0.25 19 0.27 Category C: 30 % 21 0.30 23 0.35 1) Draught rate calculated according to ISO-7730 indoor environment category by assuming typical turbulent intensity of 20 % in displacement ventilation flow.

6.4.2 Velocity distribution

Figure 6.5A shows how the cold stratified layer of supply air flows into the occupied zone as a radial air movement that covers the whole floor in the room.

Nordtest (2002) differentiates between the acceleration region near the diffuser and the velocity decay region outside the acceleration region (Figure 6.10). The flow near displacement diffusers has been studied and a model for the velocity distribution has been developed (Magnier-Bergeron et al. 2017).

A typical height of the discharge flow is about 20 cm in the velocity decay region. The maximum velocity in the flow is located at a height of approximately 2 to 5 cm above – see Figure 6.10. This is also reported by Melikov and Langkilde (1990).

Typical ~ 2 - 5 cm depth vx 20 cm v

Velocity Acceleration decay region region Figure 6.9. Adjacent zones for wall-mounted diffusers. a) Forward discharge, b) Sideways discharge. Diffuser data: Height: H = 0,9 m, Figure 6.10. Velocity distribution in front of a Width: B = 0,6 m. Supply air flow: qs = diffuser, when the supply air is colder than the 0,04 m³/s. Under-temperature: Δθs = 6 K. room air (Jacobsen and Nielsen 1992).

Measurements of this air movement show vx [m/s] that the entrainment in the horizontal flow 0,60 is very small and the height of the flow is constant along the whole velocity decay re- 0,40 gion. The height of the stratified layer is a function of the Archimedes number (Niel- 0,20 sen 2000). The height of the diffuser is important because the supplied cold flow is in- 0,10 fluenced by initial vertical acceleration due 0,08 to gravity. The Archimedes number is de- 0,06 fined as: 0,04 0,2 0,4 0,6 1,0 2,0 4,0 6,0 x [m] Hg oz  s )( Ar  2 (6.1) Figure 6.11. Maximum velocity close to the v f floor versus distance x from the diffuser. or as the Archimedes ratio The cold air has a high initial acceleration due to the buoyancy effect, and the highest  )( velocity is, in this case, obtained at a dis- oz s 2 (6.2) tance of 0,6 m from the diffuser. The meas- qs urements indicate that the velocity v is pro- x portional to 1/x for distances larger than where ~1 m from the diffuser.

β = volume expansion coefficient, 1/θ K-1; oz Therefore, it is possible to find the maxi- g = acceleration of gravity = 9,81 m/s²; mum velocity v at any distance from the H= height of diffuser; x diffuser when the adjacent zone l , is known θ – θ = under-temperature, i.e. differ- n oz s from measurements. The maximum veloc- ence between the temperature at a ity is given from height of 1,1 m inside the room and the supply temperature; l vf = face velocity (qs/as); n x  ,v [m/s]20 (6.3) qs = supply air flow rate. x

A radial stratified flow with constant height where x is the distance from the diffuser. will have a velocity distribution that is in- versely proportional to the distance from The velocity vx will for example, be equal the diffuser or from a point (virtual origin) to 0,075 m/s at a distance of 4 m if the very close to the diffuser. Measurements of length ln is 1,5 m. The maximum velocity the flow from diffusers confirm the theory will be located 2 to 5 cm above the floor of this development (Nielsen 1992b, 2000 when the temperature difference is large, and Skåret 2000). Figure 6.11 shows an ex- but it will have a higher location if the tem- ample of measurements of maximum ve- perature difference is small. It is assumed locity in the stratified flow along the floor in equation (6.3) that ln is within the veloc- from a wall-mounted diffuser. ity decay region (Figure 6.11).

It is also possible to find the velocity dis- dial/axial distribution of velocity, Fig- tribution in the occupied zone as a function ure 6.1 and 6.2, and the lower part is typical of the volume flow rate and the tempera- of a “flat” velocity distribution, Figure 6.3. ture difference. The maximum velocity is given by Some diffusers may also have initial high mixing to reduce the velocity level along the floor. vx = qs KDr⋅(1/x) [m/s] (6.4) The Archimedes number may be reduced to

zero in some situations (for example, in the where K is a function of the flow rate and Dr situation where a CAV-system reduces θ - the under-temperature of the supply flow oz θ to zero by the temperature control). Fig- (function of Archimedes number). The s ure 6.12 shows that K obtains a given equation is valid for the velocity decay re- Dr level and equation (6.4) describes the flow gion, x > 1−1,5 m. as an isothermal radial wall jet, which will be formed in the occupied zone. The length of the adjacent zone, ln, can be found from equation (6.5): Figure 6.12 indicates that KDr is a function of the square root of the Archimedes

 5 sn Kql Dr [m] (6.5) number, or √휃표푧 − 휃푠 / 푞푠. The maximum velocity vx will therefore be a linear function of the square root of the Archi- KDr has to be measured for each individual medes number. diffuser. Figure 6.12 shows the variation of KDr for different types of diffusers. The figure indicates that the first generation of diffusers – located in the upper part of the K (m-1) shaded area – had a radial distribution of Dr 16 the flow and a high level of the KDr value. Some diffusers even had a forward dis- 14 charge of the flow at a low Archimedes 12 number, which in this situation will give a 10 further increase in KDr, but the gravity ef- 8 fect turns the flow into a radial pattern at 6 higher Archimedes numbers. The newest generation of diffusers has a distribution 4 with high velocity parallel to the wall and a 2 lower velocity perpendicular to the wall 0 (sideways discharge). This will give the 14121086420 16 θ – θ 2 low KDr values shown in the lower part of oz s . -3 Ks 2 10 [ 6 ] Figure 6.12. qs m Figure 6.12. KDr-values for different types of The upper part of the shaded area in the fig- wall-mounted diffusers for displacement venti- ure is therefore typical of diffusers with ra- lation.

The KDr value is expressed by the following agreement with the values given in Figure equation (Nielsen 2000). 6.12. A further increase in the velocity level will be obtained by a design where αo is eb smaller than π, which also is typical of an  ,K 90 m -1 Dr [m ] (6.6) early diffuser design. o

A design with sideways discharge, see Fig- where ure 6.3, can for example be expressed by a e = factor that represents the initial in- b value of ~ 0,85, which gives a K value crease in the flow rate, which is due m Dr of ~ 6, which is typical of the new genera- to entrainment in the downward action of diffusers. celerating flow close to the opening, b = factor that adjusts the flow in the di- m New diffusers with adjustable nozzles be- rection of the x-axis to the flow pro- hind the front cover can be set to provide file generated by the diffuser, see strong mixing and generate such a high Nielsen (1994a, 2000), level of turbulence in front of the diffuser α = the angular width of the radial flow o that the high entrainment removes a large close to the diffuser part of the temperature difference in the δ = the thickness of the stratified flow de- flow. The flow will therefore have a re- fined as the height from the floor to stricted velocity in the direction away from the level where the velocity is 0,5 v . x the diffuser. The restricted temperature dif-

ference in the flow means that δ will be The variables are indicated in Figure 6.13. large and K will therefore be small. A Both e and b are functions of the Archime- Dr m small K means a low velocity v and, in des number. Dr x this case, that disturbance from other sources in the room could dominate the flow in the occupied zone. qs e qs  A Nordtest method (2002) gives a test pro- 0 b > 1 m cedure for diffusers and an expression for x0  the velocity distribution that can be estab- vx lished from measurements.

Figure 6.13. The diffuser has a forward dis- Example: charge, bm > 1, and αo is smaller than π. The distance xo can be ignored in many practical Calculation of the adjacent zone for a cases. wall-mounted diffuser. In practice, the crucial question is: “How long will the An all-round conventional diffuser has a adjacent zone be for a given supply air KDr value of ~ 7 corresponding to δ ~ 0,1 m, flow rate?” The following example αo = π and bm ~ 1. Many of the early diffus- shows the calculation of the adjacent ers had a radial flow distribution with a rel- zone for a wall-mounted diffuser with atively high level in the symmetry plane as, adjustable nozzles behind the front cover for example, a bm value of 1,5. This will (Figure 6.14). give a KDr -value of 11, which is in good

6.4.3 Air distribution from free standing and corner mounted diffusers B = 0,54 m A frequently used low velocity diffuser is H = the circular freestanding unit. The supply 0,45 m duct can be either from below (as shown) or from above.

l0,2

Figure 6.14. The wall-mounted diffuser of the example. H = 0,45 m, B = 0,54 m.

Figure 6.15. Circular, freestanding unit. The KDr-value for the diffuser with a given standard set up of the nozzles is given by: This diffuser can be regarded as two semi- circular diffusers standing beside each  other, i.e. it can supply twice as much air as oz s -1 Dr 118,K 5 2  7,78 [m ] (6.7) a semi-circular diffuser for a given length qs of the adjacent zone. The KDr factor can be expressed in the same way as for wall This KDr-function could be evaluated from mounted diffusers and the adjacent zone is laboratory measurements. From equation given by equation (6.5) and the velocity (6.7) and (6.5) the values in Table 6.2 can along the floor by equation (6.3). be calculated. A corner-mounted diffuser can be regarded as half a wall mounted diffuser or a semi-cir- Table 6.2. Length of adjacent zone for cular diffuser with regard to the length of the θ – θ = 3 K. oz s adjacent zone. However, some manufactur- -1 qs [m³/s] KDr [m ] ln [m] ers make the discharge flow direct along the 0,02 8,63 0,86 walls in order to avoid draught along the 0,03 8,14 1,22 floor in the occupied space in the room. 0,04 7,97 1,59 0,06 7,85 2,35

The obtained length of the adjacent zone may be too long for some applications. It is possible to adjust the nozzles in the diffuser to obtain a flow with more sideways discharge. This adjustment will reduce the KDr-values to a lower level, and decrease Figure 6.16. Corner-mounted diffuser (Half of the length of the adjacent zone. a semi-circular diffuser).

6.4.4 Documentation for diffusers giv- vx is maximum velocity in distance x from ing radial flow the opening and qob is the excess air supplied on the other side. vx is measured in the Documentation that is suitable for use by symmetry plane. computer calculation methods should have the diffuser constants given as a function of the under-temperature and the supply air volume flow:

KDr = f {(θoz – θs), qs} (6.8)

The length of the adjacent zone could then be calculated from formula (6.5) and the maximum velocity in the occupied zone can be calculated from (6.4).

6.4.5 Flow between obstacles

The flow in the vicinity of the floor may be θ influenced by furniture and by other obsta- θob oz v cles in the occupied zone. The maximum ve- x locity in the flow is located close to the floor (between 2 to 5 cm above the floor), and a x large part of the air movement will therefore take place in this region (Melikov and Lang- kilde 1990, Nielsen 1992a). Furniture with some air gap at floor level will only have a small influence on the air movement while obstacles placed directly on the floor will block the flow. An opening between these Figure 6.17. Office with short movable walls. types of obstacles will work as a new supply Flow through openings between obstacles and opening because the flow in the room is definitions of the variables. stratified. Figure 6.17 shows an example from a room with short movable walls. Figure 6.18 shows the measurements of Experiments have shown that the flow from Kob in equation (6.9). The structure of equa- an opening between obstacles can be de- tion (6.9) and the distribution of Kob -values scribed as a semi-radial flow like the air are equivalent to the structure of equation movement from a wall-mounted supply (6.4) and the structure of KDr -values. The opening, Nielsen (1992b, 2000). The veloc- temperature difference θoz − θob is the dif- ity decay can be described by the equation: ference between the temperature at a height 1,1 m in front of the opening and the lowest temperature in the opening between the ob- vx 1  Kob (6.9) stacles. qob x

The width of the opening is varied from of the flow rate and temperature difference 0,1 m to 1,5 m in the experiment in Figure (Archimedes number) and it is dependent 6.18 and measurements show that the width on the type of diffuser and on the installa- has limited significance. tion of the diffusers (distance between diffusers). The equation shows that the veloc- -1 Kob [m ] ity is independent of the axial distance x, but it must be assumed that the Archimedes 16 number has to have a certain level. The 14 same type of plane flow will also be gener- 12 ated in a narrow and deep room with a single diffuser located at the end wall. 10 8 It is clear that it is not possible to work with the adjacent zone theory for sizing except if 6 the velocity is high close to the diffuser. A 4 diffuser selection just requires that vx is smaller than a given value, which will be 2 the case in the whole occupied zone up to 0 86420 the location of the heat loads.

θ – θ 2 oz ob. -3 °C s Even a single diffuser can generate a plane q 2 10 [ 6 ] ob m flow in the occupied zone if the room is Figure 6.18. Kob versus flow rate and tempera- deep and the diffuser has a high sideways ture difference, Nielsen (1992b, 2000). discharge. Figure 6.19 shows, as an example, the flow from a wall mounted diffuser and the direction of the velocity in the oc- 6.5 Air distribution from wall- cupied zone. mounted diffusers giving plane flow in the occupied zone

The flow from a number of diffusers placed close to each other on the wall will merge into a plane, stratified flow. See Figure 6.5B. The velocity does not decrease with the distance from the diffusers in this area, and can be expressed by the following equation, see Nielsen (1994b):

 l,sx Kqv Dp (6.10)

The flow rate qs,l should, in this case, be Figure 6.19. Velocity distribution in a room given as flow rate per m width of the main with a low velocity diffuser with high sideways air movement. The KDp value is a function discharge.

Figure 6.20 shows typical KDp values for diffusers located close to each other (blue and black) and for a diffuser with high sideways discharge (green), Nielsen et al. (2004).

The Archimedes coefficient is given by

 oz s 3 Figure 6.21. Floor-mounted diffuser with arw  2 10 (6.11) q l,s swirling flow.

An advantage of floor-mounted diffusers is the large entrainment of room air into the primary air. This makes them well suited 12 for large temperature differences, and they are often used in rooms with high thermal 10 loads. The supply area is small and the supply velocity should have a sufficient mo- ] 8 -1 mentum in the vertical direction (with a

[m 6

Dp supply velocity of 2 ~ 4 m/s). When using K 4 floor-mounted diffusers, care should be taken to apply the right air volume flow. 2 Too much air might be discharged up into 0 the upper layer, so that it creates a mixing 0 10 20 30 air distribution. On the other hand, too little arw [Ks2/m4] air will create too low momentum and in-

sufficient mixing with the room air, which Figure 6.20. Examples of K values for plane Dp creates a cold air layer along the floor. flow versus an Archimedes coefficient arw.

Figure 6.22 shows the vertical throw, zm, from a floor mounted diffuser, Fitzner (1989). This throw should not be confused with the stratification height which is present 6.6 Air distribution from floor- in a room with displacement ventilation. mounted diffusers

A floor-mounted diffuser supplies air verti- cally from the floor, and they are often de- 21°C signed to generate a flow with a swirl. In zm this way, room air is entrained efficiently zst into the primary air, which implies that the velocity decreases very rapidly. Also, the Figure 6.22. Maximum height of flow from a temperatures are mixed very rapidly. See floor mounted diffuser and stratification height Figure 6.21. in a room with displacement ventilation.

The height zm can be found from, Koestel vz / vo (1954): 1,0 0,8 ,50 0,6   zm  vK sa  0,4  3,33 (6.12) 0,3 a   )( a  o  oz os  0,2

The velocity decay in the upward flow for 0,10 0,08 a floor-mounted diffuser with swirl can be 0,06 described by the same equation as for a 0,04 free, circular jet: 0,03 0,02 Free jet Jet with swirl v K ao z a  (6.13) 0,01 vo 2 z 1 2 4 6 10 20 40 z

a where o vz = the maximum velocity at the dis- Figure 6.23. Dimensionless velocity decay vz/ vo tance z above the floor; in a free jet and in a jet with swirl versus the vs = supply velocity for floor mounted height z above the floor. diffusers = qs /ao; ao= the supply area of the diffuser; Ka= wall jet diffuser constant. Wall-mounted and floor-mounted diffusers can handle a load in the room of ~ 50 W/m². The velocity decay for two different floor- Carpet diffusers, which are designed as an mounted diffusers is shown in Figure 6.23. upper part of a double floor covered by a It can be seen that the velocity decays much carpet, can handle loads larger than more rapidly for a jet with swirl than for a 100 W/m². Carpet diffusers have a surface jet without swirl. The value of Ka for the (floor) temperature θf equal to θs, which in- free jet (without swirl) is 6,8, and the Ka- fluences the vertical temperature gradient value for the jet with swirl is 0,42. See Niel- and make the gradient larger than in a dis- sen et al. (1988). This expresses how the placement ventilated room with one of the swirl will generate a high entrainment and other types of diffusers. Carpet diffusers do very fast velocity decay. The diffuser is of- not create any velocity in the room due to ten used in a group of four within an area of the extreme low supply velocity qs/Afloor. 0,6 m x 0,6 m. The velocity level will in The velocities in the room are generated by this case be higher than the velocity level the heat sources in the room and the design from a single diffuser but both arrange- procedure is mostly about finding the max- ments will have the same velocity level at a imum thermal load giving an acceptable ve- height of 0,8 m (Bauman 2003). locity level (Nielsen 2011).

7 Design of displacement ventilation

7.1 Design criteria 7.2.1 A design chart for the room air distribution systems Air quality based design is typically used in industrial applications where the contami- Figure 7.1 provides a chart that describes nant stratification plays an important role. the main considerations required when de- Chapter 4.4 provides the equations for air signing air distribution for rooms that re- quality based design. quire cooling (Nielsen 2007). The chart is based on the minimum and maximum al- Temperature based design is the most com- lowable flow rate qs supplied to the room, mon method in commercial buildings. and also on the maximum temperature dif- Chapter 5 provides methods of how to eval- ference between return air and supply air. uate vertical room air temperature distribu- Figure 7.1 indicates the required outdoor tion. air minimum flow rate into the room to obtain a given air quality. The minimum flow A calculation method should be used that is rate can, for example, be that given by suitably accurate for the particular heat standard EN 15251 (2007). It can also be loads. Dynamic energy simulation should seen that draughts will occur at a particular be used in order to accurately estimate the flow rate which is dependent on the temper- cooling load. Simplified steady-state calcu- ature difference θe−θs and on the type, and lations typically overestimate the actual de- location, of diffusers. mand and cannot take into account the effect of thermal mass on room air tempera- θe – θs ture. In demanding design cases a full-scale Draught or/and mock-up is recommended together with temperature CFD. difference Draught 7.2 Design of air distribution Low air quality Low Design of air distribution is crucial in displacement ventilation to guarantee draught q free conditions and good air quality across s the whole occupied zone. The principles for Figure 7.1. A design chart that indicates the re- the calculation of the required air flow rate strictions on the flow rate qs and on the tempera- are presented in Chapter 5. The required to- ture difference θe-θs between return and supply. tal air flow rate is the basis for the selection of supply units. In Chapter 6, the perfor- The temperature difference, θe−θs, between mance of different types of displacement the return and supply air is also limited as ventilation air diffusers is discussed. Spe- indicated in Figure 7.1. An excessive tem- cifically, the adjacent zone, close to the perature difference may either cause units, should be analysed to prevent draught in the occupied zone or create an draught. extreme temperature gradient in the room.

Figure 7.2 shows an example on a design and locations can be selected taking into ac- chart for an office with two persons and a count the following considerations: give displacement diffuser. The curve shows the combination of θe−θs and qs • dimension of the room which encloses an area that fulfils thermal • the location of heat loads and/or pollutant comfort requirements. sources • possible locations of units In Figure 7.2, a maximum velocity of • internal obstacles of the flow 0,2 m/s is accepted where the stratified • restriction of installation flow enters the occupied zone (adjacent • architectural aspects zone equal to 1 m) and it gives the restriction indicated by the blue line (equa- Typical standard diffuser types are: wall tions 6.4 and 6.5). The red vertical line mounted, corner mounted, freestanding and shows the minimum flow rate based on air floor mounted. The diffusers require a cer- quality. The blue horizontal dotted line tain wall and floor area. With regard to the shows the restriction on the temperature type and location of diffusers the following difference to the room which limits the ver- aspects should be taken into account: tical temperature gradient in the room to 2,5 K/m. Thus, the area enclosed by the • the supplied air should be uniformly dis- curve, the horizontal line and the vertical tributed in the room by a sufficient num- line (the white area) defines the permitted ber of units; ranges for variation of θe-θs and qs in a room • special attention needs to be given to the with this type of displacement ventilation. adjacent zone around the diffuser so that it is as small as possible.

Knowing the total airflow rate, it is possible to estimate the number of required supply units. Table 7.1 (Halton 2000) presents a preselection of the supply unit.

As a rule of thumb in a large open space layout, the maximum distance between supply units is 30 m (Figure 7.3). If the distance between the supply units is more than 30 m, an additional row of supply units be-

Figure 7.2. Design chart for the air distribution tween these units is needed. where lines show the different limitations for the air flow rate (m³/s) and the temperature difference, which ensure thermal comfort in the room Exhausts in the ceiling (white area) (Nielsen 2007) Low velocity units 7.2.2 Location and number of units on the floor level max 30 m When the total required supply air flow rate Figure 7.3. A rule of thumb of maximum dis- is calculated the diffuser type, the number, tance between the supply units.

Table 7.1. Air flow rates and covered floor areas with different nominal sizes of displacement units.

Nominal size [mm] Airflow rate per unit [m³/s[ Floor area per supply unit [m²] 100 Up to 0,030 10-15 125 0,020-0,030 10-20 160 0,030-0,080 10-30 200 0,070-0,150 10-50 250 0,100-0,200 15-60 315 0,200-0,400 20-70 400 0,250-0,500 30-100 500 0,400-0,800 40-150 630 0,600-1,300 50-170

If there are special requirements for interior 7.2.3 Selection of supply unit design, supply units can be recessed in the structure. If needed, the supply units can be In practice, supply units will often be se- covered with special decorative panels. In lected with product related software or prod- this case the free supply area should be de- uct specific design graphs. Figure 7.4 shows signed to guarantee the normal perfor- an example of product selection. With se- mance of the air supply units. The units can lected air flow rate and supply air tempera- be painted to meet particular decoration ture, adjacent zone and possible draught risk needs. being the main selection criteria. The pressure difference over the supply unit and the The selected supply units may serve as a sound pressure level (Lp) are also important visible architectural element of the interior considerations in the design process. design, for example, by using free-standing units installed on the floor within the space. 7.2.4 Ducting systems The use of free-standing units makes design of uniform air distribution for large spaces For ducting, there are three possible solu- relatively simple. tions: ducting through the ceiling and walls;

Figure 7.4. Example of product data (adjacent zone, pressure drop and sound level) in displacement unit selection.

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 7. DESIGN OF DISPLACEMENT VENTILATION ducting through the floor; and employing and 30 Pa. With relatively low pressure an under-floor plenum. The exhaust of the drop (<15 to 20 Pa), it is possible to meet room air is located above the occupied zone the high acoustic demands that are required - preferably close to ceiling level. in concert halls.

Ducting through ceilings and walls When ducting through the ceiling the typical location for supply units is close to the walls where the units are freely mounted on the floor or embedded in the wall structure. Ducts are installed either visibly on the wall or hidden inside the structure.

Air supply units can also be integrated in a column, which creates an ideal ventilation solution for the central part of large open area. The supply units can be designed to look like columns, which can suit the interior environment (Figure 7.5). The height of the diffuser is selected based on the design concept.

Ducting through floor When the units are connected to the ductwork through the floor, it is possible to select the suitable location of the supply unit quite flexibly. Figure 7.5. Ceiling ducted displacement units that are integrated with pillar structure. The supply unit can be a visible element as part of the interior design. The industrial design of the unit could be specifically tai- lored to meet the needs of the interior decoration. Figure 7.6 shows an example of visual displacement units that are ducted through the floor.

Supply plenum Air supplied through a under floor plenum system is an excellent solution for applications such as concert halls and theatres (Figure 7.7). With supply plenum system typically, the height of the raised floor is 0,30–0,45 m. To reduce the air leakage and noise generation of the supply units, the rel- Figure 7.6. Floor ducted displacement units ative pressure is maintained between 10 Pa that are architectural visual elements.

Figure 7.7. Supply plenum with under floor supply units in a concert hall.

In office applications, where standard the cold walls/windows (Figure 7.9). One raised floor elements are used, as a rule-of- concentrated convector may cause mixing thumb, the expected leakage of supply air is of the room air. 10–30 % depending on the quality of the structure. The maximum size of the underfloor plenum can be about 300 m². The maximum distance between air inlet and the point of discharge is 15–18 m.

Due to heat transfer between supply air and plenum structure, the supply air is warmed Radiator in the plenum. This makes difficult to provide a fast response, controlled, room air temperature.

7.3 Integration with separate heating and cooling systems Figure 7.8. Radiant panel – well suited for dis- 7.3.1 Integration with separate heat- placement ventilation. ing systems

Radiant panels Using radiant panels is a good method for room heating with displacement ventilation (Figure 7.8). The panels should preferably be located below the coldest elements in the Convector room, i.e. the windows and the outer walls. The larger part of the heat emission is the radiation. The minor part is convection, which will counteract the cold down- draught from windows and cold walls.

Convectors Figure 7.9. The convector works well with dis- Convectors go well with displacement ven- placement when located below the cold walls or tilation when the heat is distributed along windows.

Heating by ceiling panels The “50 % rule” can no longer be used for the Heating by ceiling panels is very suitable for rough calculation of the vertical air tempera- displacement ventilation (Figure 7.10). In ture gradient in a room when a radiant floor is normal conditions, without heating demand, used together with displacement ventilation. the ceiling is 3 K – 4 K warmer than the floor, yielding a heat transfer from ceiling to Heating panels floor of about 20 W/m². Thus, a slight increase in ceiling temperature will provide sufficient room heating. The convection part of the heat from the ceiling panels will counteract the heat loss through the ceiling.

Heated ceiling panels stabilise the thermal stratification, and thus benefit displacement Figure 7.10. Heating by ceiling panels. ventilation.

Floor heating In normal conditions, floor heating is well suited for displacement ventilation (Fig- ure 7.11). Part of the heat transmission from the floor is radiation towards the cold surfaces of the room. The convective heat transfer will heat the supply air that spreads Heated floor across the floor. Figure 7.11. Floor heating with displacement ventilation. If the floor is too warm, it will heat the air and make it rise so that it causes mixing, at A new method of calculation is proposed, least in the lower part of the room. How- using an “80 %-rule” as the limit, for a floor ever, practice has shown that with a floor heating capacity of about 60 W/m². In Fig- temperature below approximately 25 °C ure 7.12, the effect of the specific floor heat- and the supply air being some 2 K or cooler ing capacity on the vertical room air stratifi- than the room air, the supply air spreads cation is shown (Causone et al. 2010). along the floor (Causone et al. 2010).

Dimensionless temperature profile in occupied zone and floor heating capacity 1,0 Floor heating capacity: 60 W/m² 50% 0,8 Floor heating capacity: 50 W/m² 65% 0,6 y Floor heating capacity: 40 W/m² H 70% 0,4 Floor heating capacity: 35 W/m² 75% 0,2 Floor heating capacity: 0 W/m² 80% 0,0 0,0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9 1,0 1,1 (Ty – Ts) = (Te – Ts) Figure 7.12. Correlation between the dimensionless room air temperature profile in the occupied zone and the floor heating capacity (Causone et al. 2010).

7.3.2 Integration with separate cool- Chilled ceiling 2,5

ing systems [m]

z η = 0,6 2,0 The combination of chilled ceiling and dis- η = 0,5 η = 0,4 placement has been proved to meet the ther- 1,5 η = 0 mal comfort requirements of ISO7730 (2005) when designing for sedentary work- 1,0 ers in offices (Loveday et al. 2002, Hodder η = ratio of the cooled 0,5 ceiling cooling output et al. 1998, Alamdari et al. 1998). Also, to the total cooling passive chilled beams can be integrated in a 0,0 output (Tan 1998) similar way to displacement ventilation, but aboveHeight level, floor 0,8 1,0 1,2 1,4 that has not been studied extensively. Relative air temperature The combined system of a chilled ceiling, (relative to temp. at 0,1 m above the floor) displacement ventilation and desiccant de- Figure 7.13. Vertical air temperature distribu- humidification is proposed for space condition in a room with chilled ceiling. Tempera- tioning in hot and humid climates to im- tures relative to temperature at 0,1 metre above the floor (Tan et al. 1998) prove energy efficiency (Hao et al. 2007). Figure 7.14 shows the contamination ratio Choosing displacement ventilation as an air (contaminant in the occupied zone /ex- distribution method does not by itself result haust) as a function of the relative cooling in a stratification strategy, if the whole load from the chilled ceiling (Krühne room air conditioning system is not de- 1995). Figure 7.14 shows that when the signed for that purpose. One example of relative cooling of ceiling is increased the that is a system consisting of displacement contaminant in the occupied zone also in- ventilation and chilled ceilings. Low veloc- creases. ity air supply and cooled ceiling systems

behave like mixing systems when the e

cooled ceiling provides a substantial part of oz c /c the cooling (Tan et al. 1998 Rees and Haves 1,2 Heat surplus per unit floor area 2013 and Schiavon et al. 2015). 1,0 /A = 12 W/m² 0,8 /A = 20 W/m² In Figure 7.13 the vertical air temperature /A = 30 W/m² distribution as a function of that de- 0,6 /A = 50 W/m² scribes the ratio of the cooled ceiling output /A = 65 - 93 W/m² to the total cooling output, is presented 0,4 /A = 50 W/m² (Tan et al. 1998). The stratification effect 0,2 decreases gradually as the relative cooling load of the ceiling, , increases. When is 0,0 0,0 0,2 0,4 0,6 0,8 1,0 less than about 0.6 there still is some thermal stratification in the room. A similar Contamination ratio in occupied zone, Relative cooling load of cooled ceiling,  type of behaviour has also been found with Figure 7.14. Contamination ratio in occupied the contaminant stratification. space versus cooling effect from ceiling panels.

When chilled ceiling panels are used, the sensible heat removal terminal due to its di- supply air temperature should be kept con- rect absorption of solar radiation (Zhao et stant and the panels should be used to con- al. 2016). trol the room temperature (Figure 7.15). The best alternative is variable air volume Figure 7.15 presents the air temperature systems controlled by an air quality sensor, stratification in room when radiant floor and having the cooling panels control the cooling and displacement ventilation is operative air temperature in the room. used (Causone et al. 2010).

The radiant floor effects the temperature Chilled ceiling controls Air quality gradient and higher vertical air temperature the room temperature sensor differences are expected. The “50 %-rule” is not valid in the occupied area. It is also evident that increasing the air flow rate, and thus raising the floor temperature, the vertical air temperature differences decrease.

θs=Constant It should be noted that in many cases the vertical air temperature differences be- Temperature sensor tween head and ankles could be higher than Figure 7.15. Temperature control in a room the limits imposed by ISO standard due to with chilled ceiling. the action of the floor cooling. This phe- nomenon is probably less important in buildings with high solar loads. In these Radiant floor cooling works well with dis- buildings, lower vertical air temperature placement ventilation and it improves ther- differences occur, and thus the use of floor mal comfort in large spaces (Simmonds et cooling should not create any local thermal al. 2000). A radiant floor is an effective discomfort (Causone et al. 2010).

2,5

S3 2,0 S2 S1 1,5 50%

Height [m] Height 1,0

0,5

0,0 20 21 22 23 24 25 26 27 28 29 30

Air temperature [°C] Figure 7.16. Vertical air temperature profile at three locations (S1-S3) in the room of 16,8 m² with airflow rate 0,050 m³/s and supply air temperature of 20 °C. Radiant floor cooling capacity is 31 W/m² (Causone et al. 2010).

It is reported that radiant floor cooling does air temperature should not vary too much to not disturb convection flows of heat loads avoid creating draughts along the floor. and thus gives ideal additional cooling for displacement applications (Babiak et al. 2009). In that case, as applied in the Bang- kok airport terminal building, the surface cooling system provides a maximum cooling load of 70–80 W/m² when it works with a constant supply water temperature of 13 °C and a return temperature of 19 °C.

7.4 Control of indoor conditions

Control of displacement ventilation does not differ much from the control of mixing ventilation. The main difference is the location of the sensors for air quality and air temperature.

The location of the temperature and air quality sensors should be carefully considered. The location of the temperature/air quality sensors depends on the height of the room and on the ventilation system.

Constant Air Volume System In some applications, supply air temperature and air flow rate have been kept constant with acceptable results. In these cases, Figure 7.17. Idealised temperature distribu- the thermal mass of the building is utilized tion curves for varying heat loads and different and a slight increase in the room air temper- levels of the temperature sensors. ature is accepted during the occupied period. In those applications, the occupied pe- An example of a control curve for a venti- riod is quite short, for example a 1,5 to 3- lation system with cooling of the supply air hour concert. is shown in Figure 7.18. When the temperature in the occupied zone, oz, exceeds a In a constant air volume system, the supply certain maximum value, oz 2, the supply air air temperature should be controlled to give temperature is kept at its minimum value, a constant room air temperature at a certain say 18 °C (depending on the performance height in the occupied zone. In these cases, of the diffuser). When oz decreases below the result depends on the height of the tem- a certain minimum value, oz 1, the supply perature sensor, as illustrated in Fig- air temperature is kept at its maximum ure 7.17. It should be noted that the supply value, say 20 °C.

NB: Supply air temperature, s Supply temp s > room temp oz Not recommended!

22°C

s max (~20°C?)

s min (~18°C?) Temperature in the occupied zone,  18°C 20°C 22°C oz

oz 1 oz 2 (~21°C?) (~23°C?)

Figure 7.18. Supply temperature, s, controlled by air temperature in the occupied zone, oz.

Variable Air Volume System Temperature sensor Air quality sensor In the case of variable air flow rate, the air volume flow is controlled according to the air quality or according to the temperature in the occupied zone. Occupied To prevent cold draughts along the floor, zone the temperature sensor for displacement system should be located at 0,6 m above the floor for rooms with wall-mounted or freestanding diffusers and the air quality sensor Figure 7.19. Location of temperature and air at the height of breathing zone of a sitting quality sensors in a room with larger ceiling person (1,1 m) (Figure 7.15) (Fitzner height. 2001). Displacement ventilation is well suited for In tall rooms the air quality sensor should variable air volume systems. When the air be located at the top of the occupied zone, flow rate through a low velocity diffuser is because this is the zone most prone to infe- reduced, the adjacent zone also decreases rior air quality (Figure 7.19). (Figure 7.20).

Reduced air volume flow = Reduced adjacent zone Figure 7.20. Low velocity diffusers are well suited for variable air volume flow (VAV).

When the dominant contamination source The same logic applies to tall rooms as to are people, CO2 is a natural control param- rooms with normal ceiling heights. Addi- eter. The control logic can be illustrated by tionally, the difference in temperature be- the curve in Figure 7.21. The air quality tween the lowest and the highest level in the limits shown in Figure 7.21 are examples, occupied zone has to be taken into account. and must be chosen according to the target When the vertical temperature difference values set in each specific case. becomes too large, increasing the ventilation rate can reduce it. When there is no dominant or measurable contaminant source, the temperature may be the best controlling parameter (Fig- Ventilation rate, qs ure 7.22). Max

Ventilation rate, q s Min Max Room air temperature,  Min Max oz Min (~ 20°C) (~ 23°C?) CO2 - Figure 7.22. Control logic NB:for supply air flow concentration rate, controlled by the temperaturelocation of in the occu- Min Max pied zone. the sensor (~ 600 ppm) (~ 1000 ppm) Figure 7.21. Control logic for supply air flow rate, controlled by air quality (for example, CO2-concentration).

8 Case studies

8.1 Air distribution with four air distribution was identified at three dif- typical air supply methods in a ferent load conditions: summer conditions classroom with maximum occupancy (cooling load of 54 W/m²) and partly occupied (cooling The performance of four typical air distribu- load of 40 W/m²) and winter conditions tion methods in winter and summer condi- with partly occupied room (heating de- tions with different occupancy ratio was mand of 38 W/m²). The room was venti- studied by physical measurements and lated at 0,006 m³/s per person in all cases. smoke visualization in a mock-up classroom In the winter condition, a radiator was in- (Kosonen and Mustakallio 2010) and by troduced underneath the window to pre- CFD- simulations (Mustakallio and Koso- vent the risk of draught due to downward nen 2011). buoyancy flow from the cold surface of the window. The heat balance and breakdown The measured mock-up room (6,0 m x of the loads in the measurement cases are 4,4 m x 3,3 m (H)) was half of an actual presented in Table 8.1. Utilizing dynamic classroom (floor area 6 m x 10 m). The energy simulations, room air temperatures mock-up chamber was located inside an- in winter and summer were set to corre- other chamber. During the test, the sur- spond to average conditions in Scandina- rounding temperature was controlled vian classrooms. During the full-scale based on demand to reach the room air room measurements heat loads were offset temperature setpoint. The simulated win- by heat transfer through the structures to dow size was 4,4 m x 1,4 m (H). The room attain the room air temperature required.

Table 8.1. Heat balance and the breakdown of the loads in the mock-up classroom section.

Heat loads and heat losses of the simulated Summer Summer Winter classroom (half size of the actual classroom) Full Occupancy Half- Occupancy Half-Occupancy Room air temperature 26 °C 24 °C 21 °C Occupants - 58 W/person (total heat load) 15 (870 W) 7 (406 W) 7 (406 W) Lighting 15 W/m² 360 W 360 W 360 W Solar load or heat loss from window 197 W 296 W -448 W (surface temperature of window) (30 °C) (30 °C) (11 °C) Power of a radiator underneath window 0 W 0 W 250 W Total heat loads 1427 W 1062 W 1016 W Supply airflow rate 0,090 m³/s -972 W -756 W -324 W (supply temperature) (17 °C) (17 °C) (18 °C) Heat loss through structures -455 W -306 W -244 W Total heat losses -1427 W -1062W -1016 W

The performance of four typical air distri- Smoke and CFD visualizations of air distribution methods was studied: a corridor- bution in the fully occupied summer cases wall grille, a ceiling diffuser in the middle are shown in Figure 8.3. In summer condi- of the ceiling, a perforated duct diffuser in tions, thermal plumes did not have a signif- the middle of the ceiling, and two displace- icant effect of the performance of a wall- ment ventilation units in the floor corners grille: the momentum flux of a wall-grille (Figure 8.1). was strong. With a wall grille, the jet reaches the other side of room. Also, air is The supply units were selected based on the spread effectively over the whole occupied throw pattern analysis. The supply airflow zone with the low velocity units, whereas in rate was 0,090 m³/s (0,006 m³/s per person) summer time air supplied from the ceiling in all cases (half classroom). The supply air diffuser tends to be carried along thermal temperatures were 17 °C and 18 °C in sum- plumes from heat sources. A perforated mer and winter cases respectively. The room duct diffuser had a tendency to create un- air temperatures were 26 °C and 24 °C in stable flow conditions and varied loads can summer case with full and half occupancy unexpectedly change the throw pattern. respectively. In winter conditions, the room air temperature was set to be 21 °C. High velocities (over 0,3 m/s) were measured in the occupied zone in all conditions with a wall-grille. The highest velocities (above 0,2 m/s) were measured near the window (at a distance of 0,25 m). In all conditions velocities higher than 0,2 m/s were also measured near the floor, and at 0,1 m height, as far as 3,6 m from the window.

The displacement ventilation concept was not sensitive to load variation and air velocities were low (< 0,15 m/s) except at measurement points close to the corner-installed supply unit. With a ceiling diffuser, air velocities were, in all cases, between 0,19 and 0,23 m/s. With a perforated duct diffuser, Figure 8.1. Air distribution schemes: A) Wall relatively high velocity (0,15 – 0,2 m/s) grille, B) Displacement ventilation, C) Multi- was measured near the floor (at 0,1 m nozzle ceiling diffuser and D) Perforated duct height). In the two summer conditions the diffuser. velocity was above 0,2 m/s (up to 0,31 m/s with full occupancy) close to the floor for Air velocity and temperatures were meas- locations 3 m, and 4,8 m from the window, ured at 24 pole locations and at 7 heights i.e. the influence of heat load increased air (0,1, 0,5, 0,9, 1,3. 1,8, 2,4 and 3,1 m above velocities. This indicates more unstable the floor) at each location. a total of 168 performance with a perforated duct diffuser points. The dimensioned simulated class- when higher heat loads are introduced in rooms are shown in Figure 8.2. the classroom.

Figure 8.2. The geometry of the measured half (left) and simulated classroom (right).

Figure 8.3. Visualizations in a half classroom of air distribution using smoke and CFD in cooling mode with full occupancy. Supply air units: A) a wall-grille, B) displacement ventilation with low velocity units, C) a ceiling diffuser and D) a perforated duct diffuser.

The air distribution with the corridor wall- The supply air jet from the ceiling diffuser grille gave high velocities in all load con- tended to be carried along with thermal ditions. In winter conditions, air velocities plumes from the heat loads during summer. were particularly increased close to the In winter when there was no window plume window. effect and so the air distribution was more uniform. Ceiling diffusers can provide an ap- In principle, the throw length could be op- propriate solution in varied load conditions. timized for winter conditions and so de- liver more moderate velocities at work- With the perforated duct diffuser, the perfor- spaces close to the window (for example mance was quite unstable and sensitive to by selecting a larger wall-grille). How- the subsequent removal of high heat loads. ever, this increases the risk of draught in In such conditions, the supply air could un- summer conditions. expectedly drop causing an increased risk of draught in certain work spaces.

When applying mixing ventilation, the lated room (employing 100 mm polysty- types of heat loads have a significant effect rene) with 20,7 m² (4,6 m x 4,6 m) floor on air distribution. Therefore, when such an area and room heights of 5,2 m and 3,3 m air distribution strategy is designed, the sys- Figure 8.4). tem performance should be analysed under different conditions. In the design phase, it In Chapter 5, the principles of temperature is not possible to analyse the interaction of based design models were described. In this convection flows and jets without using chapter, three of the models are applied and CFD- simulation or laboratory mock-ups. compared with measurements. Those models are: Mundt (1996), Nielsen (1995 and Displacement ventilation was least sensi- 2003) and Mateus and da Graça (2015). tive for different heat load conditions of all studied concepts. Using a ceiling diffuser, air velocities were reasonable low in all cases. A wall grille gave high velocities in both summer and winter conditions. With a perforated duct diffuser, air distribution was quite unstable causing increased draught risk in some load conditions. The performance of the wall-grille and the perforated duct diffuser were particularly sensitive to the strength and location of heat loads.

8.2 Comparison of calculated and Figure 8.4. Test facility. measured vertical temperature gradients for displacement air distribution The measured and calculated vertical profiles of room air temperature are compared During the full-scale experiments some for a space with 12 occupants and one with typical convection flow elements as well as a warm window (Figure 8.5). Figure 8.6, combinations of flow elements were meas- shows a comparison of measured and cal- ured (Kosonen et al. 2016). The internal culated vertical temperature profiles for heat loads consisted of heated cylinders combinations of heat load from occupants, representing people, heated cube-shaped warm window and floor. boxes representing computers, fluorescent ceiling lighting units, heated foil panels on The vertical air temperature gradient with a the wall representing window solar load, warm window and ceiling gains are quite heated foil panels on the floor representing linear whereas temperature profiles are far direct solar load, heated foil panels in the from linear with the heat provided by the ceiling representing skylight solar load. computer, people and a warm floor. With those heat loads, the major part of the ver- The test setup consisted of displacement tical air temperature gradient occurs across diffusers and ceiling exhaust in a well-insu- the occupied zone.

a) The agreement between prediction and ex- 5 periment of the three-node model is quite good when the heat loads that are located in 4 the low level of the occupied zone, such as computers, people and the warm floor. Height (m) Height 3 When the heat loads are located at high level (warm ceiling) or a linear type (warm 2 window), the modelled results did not cor- relate well with the measurements. 1 Mateus and Da Graça’s model works fine 0 with occupants, computers and floor heat 18,5 19,5 20,5 21,5 22,5 23,5 loads. However, accuracy with the warm window and warm ceiling were not so good. Measured Temperature (°C) The linear two node models (Mundt) works Mundt well with a warm window and ceiling, but Nielsen the linear model cannot accurately describe Mateus and da Graça heat loads that exist in the occupied zone. Φ = 900 W, θ = 18,7°C, q = 0,15 m³/s 12occupants s s The three-node model (Nielsen) predicts

different slopes for the temperature profile

between the nodes and gives better accu- b) racy than the linear two node models. 5

4 3,5 3 Height (m) Height 3 2,5 Height (m) Height 2 2 1,5 1 1 0,5 0 0 19,5 20,0 20,5 21,0 21,5 20,0 22,0 24,0 26,0 28,0 Measured Temperature (°C) Measured Temperature (°C) Mundt Mundt Mateus and da Graça Mateus and da Graça Nielsen Nielsen Φ10occupants = 750 W, Φfloor= 260 W, Φ = 232 W, Φ = 520 W, Φwindow_3.5 = 520 W, θs = 19,1°C, qv = 0,15 m³/s light win_3.5m θs = 18,1 °C, qv = 0,1 m³/s

Figure 8.5. Measured and calculated room air Figure 8.6. Measured and calculated room air temperature profiles a) occupants and b) warm temperature profiles with the combinations of window in a 5,2 m high test room. heat loads in a 3,3 m high test room.

The accuracy of the three-node model is (301-304) and two air-recirculating units quite good when applied with heat loads that (305-306) that were operating during are located in the low level of the occupied events providing up to 70 m³/s airflow rate. zone, like computers, people and warm At the upper level high momentum down- floor. All in all, the three nodal models are ward jet flows enhance an effective mixing quite useful in engineering calculation of indoor air. where the supply air flow rate of displacement ventilation is defined and energy con- The smoke tests indicated an upward flow sumption of the whole building is calculated. along the lower-seating area and a downward flow along the upper-seating area 8.3 Field measurements for a from where the air was moving into the multipurpose arena middle of the space volume.

8.3.1 Air distribution concept

This case study presents the use of displacement ventilation in Malmö-arena in which the seating capacity was up to 13 000 individuals. The dimensions were 100 m (L) x 90 m (W) x 30 m (H) (Figure 8.7). The arena is designed to offer a wide range of entertainment events for people ranging from a hockey game to various entertainment events. The arena comprises an ice rink, seating area, en- Figure 8.7. Malmö arena. closed suites and surrounding service areas for example, restaurants and shops.

Physical measurements were taken in the seating sectors, the ice rink and in the centre of the space volume. In addition, computational fluid dynamics (CFD) simulations were undertaken to provide a generic view of air distribution (Lestinen et al. 2012).

Displacement ventilation was employed for the lower-seating area and zoned ventilation used for the upper-seating area. The displacement supply air was distributed from under the retractable stands (on movable stands) beside the ice rink. The supply air flowed through openings below the seats.

The exhaust air was taken at the ceiling level (Figure 8.8). The overall ventilation system contained four air-handling units Figure 8.8. Ventilation system in arena.

8.3.1 Measurements and CFD- simu- 8.3.2 Performance during an ice lations hockey game

The room air sensors were installed onto During the measurement period, there were measuring masts at heights of 1,5 m, 4 000 spectators. The experiments indi- 2,5 m and 3,5 m over the seating-row and cated that the arena indoor air temperature the variables were recorded over a 3 min increased about 2 K during the game when average. (Lestinen et al. 2016). the airflow rate increased up to near 70 m³/s when CO2 reached 900 ppm with a 15 °C – CFD-simulations were used to investigate 16 °C supply air temperature. the flow patterns in the arena enclosure. The grid sizes were between 0,1 m and The room air temperature during the game 0,6 m for the whole arena model. In the was 12 °C – 17 °C at the lower-seating area arena model, the total number of grid and 15 °C – 17 °C at the upper-seating area. nodes was 19,68x106 with 102,78x106 ele- Experiments show the relatively low tem- ments. perature stratification (less than 2 K) and the well-mixed conditions in the arena (Figure 8.9). The corresponding air speeds were below 0,35 m/s in the seating areas.

Figure 8.9. The simulated room air temperatures in three cross-sections (panel a). The measured and predicted room air temperature profiles in the centre of the seating area in a hockey game (panel b). The scheme of air movement in the arena (panel c).

9 Research findings

9.1 A CFD Benchmark test for manikins may also improve the design of a manikins in displacement flow CFD manikin and so lead to new standards.

It may be important to test and adjust Com- The CSP in the displacement ventilation putational Fluid Dynamics (CFD) software benchmark is standing, facing the diffuser by comparing results with a benchmark in a displacement ventilated room, as illus- test. The benchmark test can be used as a trated in Figure 9.1. The ceiling, floor, side test of a new program, but can also be used wall and end wall should be simulated as for the development of a virtual person to solid surfaces. Full details are given on the be situated in a ventilated room. This chap- web page, www.cfd-benchmarks.com or in ter shows a benchmark test that considers Nielsen et al. (2003). the three-dimensional flow around a person in displacement ventilation The thermal manikin benchmark tests have, as at 2016, been used in more than 30 dif- This benchmark test is defined on the web- ferent papers, theses and articles. page: www.cfd-benchmarks.com

For many years thermal manikins have been used in full-scale indoor environment experiments. CFD provides an alternative to full-scale measurements. Research centres around the world have therefore developed different configurations (subroutines) to represent a Computer Simulated Person (CSP).

The CSPs can be very different in respect to size, form (employing a rectangular grid or body-fitted grid), heat emission details, Figure 9.1. A person exposed to a flow field in etc. The variations may reflect the different a displacement ventilated room. software possibilities, but may also be determined by different standards from country to country. Some examples of predic- 9.2 Full-scale tests and CFD- tions made with CSPs are given by Mura- simulations of indoor climate kami et al. 1995), Brohus and Nielsen conditions (1996b) and Topp et al. (2002). Full-scale tests and CFD-simulations have The idea behind a benchmark test is to com- been commonly used to ensure that indoor pare individual concepts under the same climate conditions meet the displacement boundary conditions. The tests of different ventilation system design criteria.

Figure 9.2. Illustration of 5,13 m high test room with objects: A) 2 supply air diffusers used in tests, B) exhaust, C) 4 heated foil panels (window), D) 10 heated cylinders (occupants), E) lighting units and P1-P3) locations of vertical temperature gradient measurement. Ceiling height of 3,29 m test room with two heated foil panel windows marked.

The required cooling power is determined Vertical temperature profiles were measured so as to maintain the occupied zone of a at three locations (P1-P3 in Figure 9.2) at room at design conditions. CFD-simula- eight heights. The inner wall surface, ambi- tions are especially useful when designing ent air, supply air and exhaust air tempera- ventilation for large and complex rooms. It tures were measured. Supply and exhaust air is essential to validate the accuracy of flow rate measurements were undertaken. widely used CFD-modelling methods Supply and exhaust air flow rates were bal- against full-scale test measurements anced for each measurement. The ductwork (Deevy et al. 2008). This example presents in the test room was insulated and all sur- indoor climate predictions with CFD and a faces of the test room were covered with full-scale test for a test room with two room 0,1m polystyrene boards and plastic foil to heights. Different CFD models are used minimise the effect of surrounding condi- (Mustakallio et al. 2012). tions on the vertical temperature stratification. However, heat flux through the walls 9.2.1 Full scale test affected the measurements especially in the case of the tests with the higher ceiling. This The test setup consisted of two displace- was noted by measuring the total amount of ment diffusers, with perforated front faces, electrical power supplied for the internal and ceiling exhaust in a well-insulated heat loads, the supply air flow and tempera- room with 20,8 m² floor area. A room with ture, and comparing the calculated exhaust two ceiling heights 5,13 m and 3,29 m was air temperature and the measured exhaust air studied. The internal heat loads consisted of temperature. The obtained heat flux can be 10 heated cylinders representing people, defined for the room surfaces during the heated foil panels on one wall representing CFD-simulations. solar load from a window surface, and fluorescent lighting units. Full-scale tests were Part of the difference between calculated undertaken in steady state conditions. The and measured values was also caused by full-scale test setup is shown in the illustra- measurement inaccuracy, but this effect was tion in Figure 9.2 and the photograph in assumed to be small and neglected when de- Figure 9.3. fining the corresponding CFD model.

Figure 9.3. Photos from the full-scale test rooms. Rectangular heated boxes on the tables simulate computers, floor standing and heated vertical cylinders simulate occupants. The middle of the three floor standing displacement diffusers seen on the left and right photos was not used.

9.2.2 CFD-simulation The supply air inlet was specified as the whole area of the displacement diffuser CFD-simulations were carried out with with an additional momentum source in the Ansys CFX 14.0.A grid comprising of front of perforated front face to account for 1,9/1,6 million unstructured elements or the effect of the free area of the perforated 430/370 thousand nodes (room height of plate. Realistic CFD models of displace- 5,13 m/room height of 3,29 m) was used. ment diffusers are needed to evaluate ther- Inflation layers were used near surfaces. mal comfort in the near zone of the diffuser. Grid independency was tested by dou- bling grid size. A SST turbulence model 9.2.3 Results with automatic wall treatment was used for the simulated cases, but for compari- The vertical temperature distribution was son same cases were calculated with measured and compared to the CFD-simu- standard k-e and RNG k-e turbulence lation results. The measured temperature models. distribution in three locations is presented in Figure 9.4. The distribution was nearly Buoyancy was modelled with Boussinesq the same in all locations. The first and last approximation and compared with ideal readings in Figure 9.4 show the supply and gas model predictions. Cases were solved exhaust air temperature. with a high (2nd) order discretization scheme, except for the turbulence equation Comparison of CFD simulation results with which was solved with a 1st order scheme. different turbulence models (Figure 9.5) The effect of additionally having a high or- showed a significant difference to the der scheme for turbulence and also includ- measured temperature in the occupied ing buoyancy turbulence source terms zone. The difference was about 1 K to were compared to the initial case. Radia- 1,5 K at 1 m to 1,5 m height. The effect of tion was modelled with a discrete transfer changing the buoyancy model from Bous- model. CFD-simulations were solved as a sinesq to the ideal gas model was negligi- steady state case so as to reach good con- ble, as also was the effect of high order dis- vergence. cretization for turbulence.

Figure 9.4. Measured vertical temperature distribution in A) 5,13 m and B) 3,29 m high test rooms.

Figure 9.5. Comparison of CFD-results and measurement in A) 5,13 m room with k-e, RNG k-e and SST turbulence models, with Boussinesq (Bsqm) and with ideal gas models for buoyancy (Igm), and with higher order turbulence discretization scheme (TurbHdsch), and B) 3,29 m room with k-e, RNG k-e and SST models.

The predicted vertical temperature stratifi- 9.3 Test on the performance of cation of the CFD-simulations had similar displacement ventilation– form to the measured temperature stratifi- proper simulation of occupants cation. The SST turbulence model made the best prediction of the temperature dis- The design of displacement ventilation of- tribution. The temperature in the lower ten includes one or more of the following part of the room was still significantly elements: full scale laboratory tests, field lower in the CFD-simulation than in the measurements and CFD simulations. Be- full-scale test for both test room heights. cause displacement air distribution is a re- This could be partly corrected by using sult of the thermal flows in spaces, it is im- displacement diffuser CFD models that portant to properly simulate the heat improve the mix of supply air with the sources. Occupants are important heat room air in the occupied zone. sources in rooms. With development of low

powered office equipment and lighting, The values of the CO2 concentration ratio windows with controlled surface tempera- above 100 % indicate that the concentration tures, etc. occupants will produce major at some points of the room was higher than part of the heat load in rooms. Therefore, in the exhaust air (Figure 9.6). It is clearly the buoyancy flow, i.e. the thermal plume, seen that for the lower heights the concen- generated by the human body is important. tration ratio values obtained with the thermal manikins are larger than the values ob- Simulators of occupants with simple geom- tained with the cylinders. The complex ge- etry, such as a cylinder and a rectangular ometry of the manikin body leads to more box, generate significantly more concen- intense air mixing around, and pollution trated plumes compared to simulators with diffusion from, the upper zone to the lower a complex body shape, such as thermal zone. In the case of the cylinders as occu- manikins. As a result, the volume flux in pant simulators, the concentration profile in the thermal plumes above a cylinder and the lower zone is steeper. above a thermal manikin can differ by 40 % (Zukowska et al. 2012b). The differences In Figure 9.6, the concentration ratio equal may affect the air distribution in rooms with to 50 %, i.e. between the lower and upper displacement air distribution. However, it is zones, is used to show the difference in the easier and cheaper to simulate occupants vertical CO2 distribution in the room be- with simplified geometries. The question tween the cases with thermal manikins and rises “how far it is possible to simplify the with cylinders. The height obtained for the geometry of the human body” without af- case using manikins is approximately 1,3 m fecting the displacement air distribution and for the cylinders it is 1,7 m. This is be- and validity of the obtained results. cause the manikin generates a thermal plume with volume flux greater than the The graph on Figure 9.6 presents normal- cylinder, and therefore equals the supply ized vertical distribution of CO2 concentra- airflow rate at a lower height than in the tion based on measurements performed in a case with the cylinder. full-scale test room (4,7 m x 5,4 m x 2,6 m) with displacement ventilation with a supply Thermal flows from the manikins cause air temperature 21,6 °C and a total flow rate more mixing in the lower zone and there- 0,080 m³/s (Zukowska et al. 2008). Two oc- fore more pollution is diffused from the up- cupants seated at two identical workstations per zone to the lower zone. The different were simulated first by two thermal mani- shapes of the occupant simulators cause kins accurately resembling human body different contaminations distributions in shape and then by two heated cylinders (as the upper zone – higher CO2 concentration shown in the photograph in Figure 9.6). ratios for the cylinders (Figure 9.6). The manikins and the cylinders had the same surface area of 1,63 m² and the same Thus, simulation of occupants by objects heat generation of 73 W. CO2 supplied with simplified geometry, such as cylin- from the top of their “head” was used for ders, is insufficient for obtaining accurate simulating human body bio-effluents. CO2 results when studying airflow in rooms concentration was measured at 16 heights with displacement ventilation. Simulation in 9 locations and at 20 points in a horizon- of the complex shape of the human body is tal plane 0,2 m below the ceiling. highly recommended. However, use of

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 9. RESEARCH FINDINGS thermal manikins during physical measure- dummy can be used to obtain reliable re- ments is expensive. Implementing and ap- sults in case of displacement air distribu- plying complex body shapes during CFD tion. However, it should be made clear that predictions is rather time consuming. Fig- heated dummy is not appropriate to study ure 9.7 shows a heated dummy with sim- the airflow in the vicinity of human body. plified geometry but with the same body In this case a thermal manikin with a com- surface area and heat production as the ther- plex body shape is to be used. Small non- mal manikin also shown in the figure. uniformity in the velocity field (±0,005 m/s) and in the temperature field Comprehensive measurements reveal that (±0,02 K) of the surrounding environment the two human body simulators (thermal affects the development of the thermal manikin and heated dummy) generate ther- plume above a sitting person and causes mal plumes with similar characteristics skewness of the plume cross-section (Zu- (Zukowska et al. 2012b). Thus, the heated kowska et al. 2010a).

Figure 9.6. Vertical CO2 concentration distribution with manikins and with cylinders. C, Cs and Ce are respectively CO2 concentration at the measured point, at the supply and at the exhaust.

Figure 9.7. Heated dummy with simplified body shape and thermal manikin with complex body shape generate thermal plums with similar characteristics.

9.4 Airborne cross infection risk in words, if the air distribution system is de- a room with displacement signed to make an efficient ventilation of ventilation the room, there will still be a microenvironment around people close to each other, We often address the ventilation of a room which can’t be influenced efficiently by at a macro scale level. Macro scale is the tra- the general air distribution system, see ditional level of description of air distribu- Figure 9.8B. tion in rooms as, for example, in standards where it is expressed that contaminant removal effectiveness in a room with displacement flow should have the level of  c ~ 1,2.

A person in a stratified flow is assumed to be exposed to the same level of contamination from another person independent of the position in the room so long as the people are not too close to each other and are standing with their faces at the same height, Figure 9.8A. The distance should be larger than 1,2 m in case of breathing, Nielsen et al. (2008).

When the people are close to each other, Figure 9.8. A) Displacement flow, same con- < 1,2 m, the exposure can rise to a high centration everywhere along the breathing level independent of the general contami- height. B) Illustration of the microenvironment nant level of the occupied zone. In other with a local high exposure.

The microenvironment around two people The concentrations cexp and coz can be given is illustrated in Figure 9.9. The exhalation in a dimensionless form when the concen- from the source person can be divided in tration is divided by the return flow concen- two parts. One part of the exhalation flows tration. into the macroenvironment, and the general room air distribution system dilutes, A personal exposure index εexp can be de- stratifies and transports this part out of the fined as (Brohus and Nielsen 1996a, Mundt room and creates a concentration distribu- et al. 2004): εexp = ce /cexp . The exposure is tion around the target person, coz. The alternatively defined as cexp /ce. other part of the exhalation from the source person flows directly to the target 9.4.1 Exposure in stratified flow at person’s breathing zone, or to this person’s the macro scale level thermal boundary layer. The contaminant distribution in a room with displacement ventilation may, in some situations, be stratified into layers through the room. The air exhaled from a person may, for example, be concentrated in a layer. Figure 9.10 shows full scale experiments with stratification of exhalation in a hospital ward. The two illustrated situations are where the source patient is lying on his/her back and then on his/her side.

Figure 9.9. A source manikin (right) and a target manikin (left). The contaminant flow between the two manikins is indicated by smoke.

The target person is exposed to a level of cexp, and this exposure therefore consists of an indirect exposure from the macroenvironment, coz, and a direct exposure from the source person’s exhalation. The concentration c can be measured direct at the stand- oz Figure 9.10. Two patients in a hospital ward. ing target person’s chest. This concentra- A) The source patient is lying on their back tion is also the target person’s inhalation and the exhalation is flowing to the upper zone concentration if this person is not influ- resulting in high contaminant removal effec- enced by a direct exposure, because the in- tiveness. B) The source patient is lying on their halation normally originates from the ther- side and the exhalation is stratified in a layer mal boundary layer (Brohus and Nielsen in the breathing height of the target patient 1996a, Bjørn and Nielsen 2002). with a high concentration, even across the full width of the ward.

Figure 9.10 shows how the exhaled air higher value indicates that a small fraction from a person can either pass to the upper of direct exposure takes place at a distance zone with high contaminant removal effec- of 110 cm. tiveness (εexp ~ 70) when the source patient is lying on their back or the exhaled air can stratify in the ward in the inhalation height with a very low contaminant removal effectiveness (εexp ~ 0,7) when the source patient is lying on their side (Qian et al. 2006).

The increased cross infection risk that can take place in the macro environment in a displacement ventilated room due to stratification of the exhalation is a serious problem in rooms occupied with people having air borne disease. Displacement ventilation can therefore not be recommended in this case (Bjørn and Nielsen. 2002 and Li et al. 2011).

9.4.2 Exposure in stratified flow in the microenvironment

Figure 9.11 shows measurements in the microenvironment between two people (simulated with thermal manikins) who are standing in four different positions, namely: face to face, face to back, face to Figure 9.11. Experiments with cross infection risk side and sitting source person. The expo- in a room with displacement ventilation. A) Illus- sure of the target person to the air exhaled tration of the four positions of the two people; B) by the source person is given as cexp/ce, Exposure versus distance between people. where cexp is the concentration in the target manikin’s inhalation, and ce is the concen- There is a remarkable increase in the direct tration at the exhaust. When the distance exposure when the distance between the between the manikins is 110 cm, the target people is less than 80 cm for the cases face manikin inhales a concentration which is to face and face to side of the target person. equal to the background concentration in The exposure increases up to 12 times the the room. The two people do not have a concentration in a fully mixed situation, in common microenvironment with respect to the face to face situation, and up to 7 times cross-infection considerations. The concen- in the face to the side of the target person tration cexp/ce is ~0,5 for face to the side and situation, when the distance is 35 cm. With face to the back, which is typical of dis- respect to the protection against cross-in- placement ventilation where the inhalation fection this is a serious setback for systems contains air from the lower zone in the generating a vertical temperature gradient room (Brohus and Nielsen 1996b). cexp/ce is (Nielsen et al. 2012 and Olmedo et al. ~1,0 for the face to face situation, and the 2012).

Protection from cross-infection seems to be due to vertical temperature difference be- high in the face to back situation. The ex- tween neck level (1,1 m height) and ankle posure cexp/ce does only reach 0,75 at a dis- level (0,1 m height above floor). Both air tance of 35 cm, which is below a fully velocity (highest near the floor) and air mixed case. It should be noted that the peo- temperature (lowest near the floor) are im- ple have the same height. A difference in portant for avoiding draught discomfort and height with a tall target person may give a they are a function of the supply air temper- larger exposure. ature and flow rate. At fixed heat load the same room air temperature (1,1 m above Many parameters may influence the cross- the floor) can be obtained by different com- infection risk between people situated close binations of supply air temperature and to each other in a ventilated room and they flow rate, i.e. small temperature difference can be summarized in the following: between the room air and supply air and high supply flow rate or vice versa. Distances between the people, positions and orientations of the people, breathing process A small temperature difference between (breathing through the mouth or through the supply and room air temperature and high nose, opening of mouth, coughing, speak- supply flow rate will lead to elevated veloc- ing), difference in the height of the people, ity at floor level and thus increased risk of activity levels of the people, number of peo- draught (draught risk is discussed in Chapter ple, temperature and vertical temperature 3). However, this will reduce the risk of local gradient in the microenvironment around the thermal discomfort due to vertical tempera- people, air velocity (speed and direction) in ture difference, because the vertical temper- the microenvironment around the people, ature difference will be small. Large temper- and turbulence level of the air flow in the mi- ature difference and small flow rate will lead croenvironment around the people. to reduction of the velocity but will increase the vertical temperature difference. Figure 9.11 gives the results for displacement ventilation in the room when the total Perceived air quality must also be consid- heat load is 500 W and the air change rate ered during design because it is directly re- is 5,6 h-1. The exhalation frequency is 19 lated to room air temperature, supply air exh/min for the source person and 15,5 for temperature and flow rate. Inhaling warm the target person. The exhaled flow is and polluted air will negatively impact per- 11 L/min for the source person and ceived air quality (ASHRAE Guide 11, 10 L/min for the target person. The source 2011, see also Chapter 3). In rooms with dis- person produces a total heat realise of 94 W placement ventilation an increase of the sup- and the target person 102 W. ply flow rate will move the stratification height to a higher level which is expected to 9.5 Displacement ventilation improve perceived air quality (inhaled air design based on occupants’ will be cleaner and cooler) and vice versa in response the case when the flow rate is decreased but the temperature difference is increased. In rooms with displacement air distribution thermal comfort concerns are focused on Thus, different approaches in the design of draught at the feet/lower leg and discomfort displacement ventilation can be adopted

Single user license only, copying and networking prohibited. All rights reserved by REHVA. REHVA Displacement Ventilation Guidebook with advantages and disadvantages. With- • operation modes: five combinations of out considering the energy implications, room air temperature, difference be- the questions are: What is more beneficial tween room temperature and supply air for occupants, high supply flow rate and temperature (ΔTs) and supply flow rate small difference between room tempera- were applied in the case of displacement ture and supply air temperature or low sup- ventilation and one combination of these ply flow rate and high temperature differ- parameters was studied in the case of ence? How does the selection of these two mixing ventilation and used for compar- parameters depend on the room air temper- ison (Table 9.1): ature? This has been studied by Dalewski • Thirty-two subjects were exposed to each et al. (2014) and is briefly explained in the of the six conditions randomly; two sub- following. jects at the time seated at the workstations performing office work on computer; Design details: • Subjects responded to questionnaires on • full-scale room (3,6 x 4,8 x 2,6 m³) fur- thermal comfort, perceived air quality, air nished to simulate an office with two movement sensation, etc. workstations (Fig. 9.12); • each workstation consisted of a desk, an adjustable chair, a desk lamp (20 W) and a laptop PC (50 W); • one semi-circular floor standing air supply diffuser for displacement ventilation (DV); • one ceiling air supply diffuser for mixing ventilation (MV); • one ceiling exhaust diffuser; • room air temperature maintained at 1,1 m Figure 9.12. Test room set-up. height;

Table 9.1. Operating modes (conditions): DV – displacement ventilation; MV – mixing ventilation.

Condition Room air ΔTs Air flow supplied by Ventilation temperature at 1,1 m DV or MV system [°C] [K] [m³/(s·person)] tested 23 °C 3 K DV 3 DV 23 0,045 23 °C 3 K MV 3 MV 26 °C 3 K DV 3 0,045 DV 26 26 °C 5 K DV 5 0,027 DV 29 °C 3 K DV 3 0,045 DV 29 29 °C 6 K DV 6 0,023 DV

Results and 26 °C, the thermal environment was acceptable for most of the subjects (only 2 % Physical measurements complained of thermal discomfort) regard- The velocity field measurements revealed less the air distribution (DV or MV), ΔTs that the workstations were not situated and the supplied flow rate (Figure 9.15). At within the adjacent zone defined as the area 29 °C more than 15 % of the subjects were near the supply air diffuser where air veloc- dissatisfied with the thermal environment. ity exceeds 0,2 m/s at a height of 0,1 m More subjects (28 %) were dissatisfied at above the floor. The adjacent zone extended ΔTs = 3 K and high flow rate (0,090 m³/s) up to 1,4 m from the DV diffuser when ΔTs than at ΔTs =6 K and flow rate of was 3 K (supply air flow of 0,090 m³/s), and 0,046 m³/s, 18 % of the subjects. was shorter, up to 1,0 m, when ΔTs was 5 K or 6 K (supply airflow of 0,054 m³/s or 0,046 m³/s respectively). In the case of mixing ventilation, air velocity at 0,1 m and 1,1 m above the floor was below 0,2 m/s.

The vertical temperature gradient existed in the case of DV, indicating stratification in the room (Figure 9.13). The vertical temperature difference between 1,1 and 0,1 m Figure 9.13. Measured vertical temperature profiles. was 2,0 K, when ΔTs was 3 K. This com- plies with environmental category A according to EN15251 (2007). Vertical temperature differences increased to 3,0 K and 3,5 K, as ΔTs was changed to 5 K and 6 K respectively. In the case of MV, no vertical temperature gradient was found, indicating that the air in the room was well mixed.

Subjective response The percentage of subjects dissatisfied with Figure 9.14. Percentage dissatisfied people perceived air quality (PAQ) was lowest at with the PAQ at the studied operating modes of room ventilation. 23 °C and was the same with displacement and mixing ventilation (Figure 9.14). This result was expected because there was not a strong pollution source in the room. The increase of the room temperature and ΔTs and decrease of the supply flow rate caused an increase in the percentage dissatisfied subjects.

The operating mode of the displacement ventilation had different impact on the people’s thermal comfort compared to the im- Figure 9.15. Percentage dissatisfied people with the thermal environment at the studied op- pact on PAQ. At room temperatures of 23 erating modes of room ventilation.

The impact of the operating mode on the air • Increase of room air temperature movement perception of the people was above 26 °C has a negative impact on also different (Figure 9.16). At 23 °C only occupants’ comfort: Increase of the few people (1–2 % of the subjects) were room air temperature reduces the strength dissatisfied with the air movement gener- of the free convection flow around the hu- ated by displacement or mixing ventilation. man body and thus its ability to entrain In the case of DV, the percentage subjects room air. The PAQ will not be felt of high dissatisfied with the air movement (too low quality even when the supply flow rate is air movement) increased with the increase increased (small temperature difference) of the room air temperature. The change of because the warm air of the free convec- ΔTs and supply flow rate had a different im- tion flow around the human body is in- pact on thermal sensation: more subjects haled. Complaints due to a warm environ- (44 %) were dissatisfied and requested ment and lack of air movement will in- more air movement at ΔTs =6 K and low crease. When DV is used at room air tem- supply flow rate (0,046 m³/s) than at ΔTs perature above 26 °C supply of cooler air =3 K and at a high flow rate of 0,090 m³/s, will help to improve occupants’ thermal 32 % of the subjects. sensation compared to increasing the supply flow rate; • No substantial difference between MV and DV in terms of the thermal comfort and PAQ: At a comfortable air temperature and without highly polluted room air DV and MV perform similarly with regard to occupants’ thermal comfort and PAQ. Request for more air movement may be reported with DV. No differences in subjects’ response exists between MV and DV alone at 23 °C in

terms of the thermal comfort, PAQ or Figure 9.16. Percentage dissatisfied people with the air movement sensation at the studied SBS (sick building symptom) symptoms. operating modes of room ventilation. However, a need for more air movement is reported with DV; • Use of DV in a warm environment with additionally provided local convective Useful outcomes for occupant based cooling is inefficient: The standards rec- design of displacement ventilation ommend an energy saving strategy by maintaining relative high room tempera- • The rate at which clean outdoor air is ture and improving occupants’ comfort supplied is more important for PAQ by locally applied air movement with el- than decrease of its temperature: The evated velocity. This strategy is not a fea- stratification height will be lowered when sible application in the case of DV. In this the supply flow rate is decreased and this case the airflow interaction will cause will have negative impact on the inhaled mixing of the room air and will destroy and perceived air quality; the stratification.

9.6 Convective boundary layer surrounding temperature (Figure 9.17a). around human body The thickness of the velocity and temperature boundary layers varies from less than People are important heat sources in occu- 5 mm at the lower legs up to 150 mm and pied spaces. In a comfortable and uniform more at the head height (Clark and Toy indoor thermal environment, the skin and 1975, Homma and Yakiyama 1988, Özcan clothing surface temperatures are higher et al. 2005, Licina et al. 2014, Voelker et al. than the indoor air temperature. The air in 2014). contact with the skin and clothing becomes warmer than the surrounding air and a temperature gradient with a resulting density gradient is established in the layer of air in the vicinity of the human body. The effect is a buoyancy force which induces upward airflow known as free (natural) convection flow.

The free convection flow starts with a convective boundary layer around the body which is transformed in a thermal plume above the body. The convective boundary a) b) layer transports pollution generated by the human body and in its surroundings to the breathing zone and therefore is important for occupants’ exposure and inhaled air quality. The pollution is transported further by the thermal plume and is mixed with the background room air. The importance of the thermal plume for the performance of displacement ventilation was already discussed in Chapter 4 and Chapter 9 (Case Figure 9.17. a) Temperature field of the CBL study 9.3). In the following the convective around seated person – the colours show the boundary layer (CBL) is discussed in the difference between the local temperature in the light of its practical importance. CBL and the room air temperature (Homma and Yakiyama 1988); b) Profiles of velocity measured with laser Doppler anemometer at the The CBL is slow, laminar and thin over the front/centre of a nude and clothed seated ther- lower parts of the body but becomes faster, mal manikin with a female body shape. Calm turbulent and thick at the height of the head. environment at 19 °C (Melikov 2015). Velocity and temperature distribution in the CBL is similar to that in a free convection flow over a heated vertical surface: near to At a comfortable room air temperature, the the surface the velocity increases from zero maximum velocity in the CBL may be as to a maximum followed by a decrease; the high as 0,25 – 0,30 m/s. It decreases when temperature decreases with increasing dis- the difference of the body surface tempera- tance from the surface until it reaches the ture and the surrounding air temperature

Single user license only, copying and networking prohibited. All rights reserved by REHVA. REHVA Displacement Ventilation Guidebook decreases, i.e. when the room temperature The CBL entrains and transports air and increases (Licina et al. 2014). The velocity pollution from the surrounding upward. In and temperature distribution in the CBL, as rooms with displacement ventilation the well as the thickness of the boundary layer largest proportion of the inhaled air origi- is influenced by the body posture, clothing nates from the CBL (Clark and Cox 1973, thermal resistance and its design, presence Zhu et al. 2005). The CBL transports pollu- of obstacles in the vicinity of the body (for tion generated in the human body micro-en- example, a desk that greatly reduces the vironment (bio-effluents, secondary prod- strength of the natural convection flow), ucts of chemical reaction between ozone etc. (Licina et al. 2014, 2015). and skin oil, pollution generated from the clothing, etc.) to the breathing zone (Licina Breathing also influences the natural con- et al. 2015, Bivolarova et al. 2017). It en- vection flow (Özcan et al. 2003, 2005, trains and transports pollution generated in- Bivolarova et al. 2017). The CBL at the doors or infiltrated from outdoors when it breathing zone of a person sitting with has arrived in the proximity of the human slightly open legs is result of interaction of body (Rim and Novoselac 2009). The CBL the convection flow which starts to develop interacts with the transient flow of breath- at the groins with the thermal flows gener- ing. The interaction of the CBL with the ated by the thighs and the legs. Such flow flow of exhalation (mouth or nose) is com- interaction is not present with a standing plex and important for exposure to pollu- posture. Leaning backwards induces a peak tion generated by the body itself (Bivo- velocity in the CBL that is substantially larova et al. 2017). (40 %) higher than when leaning forward (Licina et al. 2014). The performance of displacement ventilation with regard to providing clean air to the Clothing weakens the CBL and reduces its breathing zone depends on the location of thickness because its surface temperature is the pollution source. When the pollution lower than the skin temperature, though source is located at the floor or lower level still higher than the surrounding air temper- in the room the CBL will bring it upward to ature (Figure 9.17a). At the same room the breathing zone (Rim and Novoselac temperature covering the body with cloth- 2009, Licina et al. 2015a, 2015b). In this ing will reduce the velocity by half (Me- case the best performance of displacement likov 2015). Clothing insulation is non-uni- ventilation will be the same as mixing ven- formly distributed over the body surface tilation (Brohus and Nielsen 1994, Cermak and thus introduces non-uniformity in the et al. 2006). generated CBL. For a seated person the chair isolates part of the body from the sur- The reduction of exposure to pollution from rounding air and locally blocks the estab- a point source located near the feet can be lishment of the CBL. Changes in the CBL achieved by control of the CBL. Active and will have impact on the heat exchange be- passive control can be applied (Bolashikov tween the body and the surrounding and et al. 2010). The passive control is based on thus on occupant’s thermal comfort (Licina breaking the CBL at the lower chest level et al. 2016, Melikov 2015). with a movable board as a part of the desk

Single user license only, copying and networking prohibited. All rights reserved by REHVA. 9. RESEARCH FINDINGS design (Figure 9.18b). The active control its performance. For example, the genera- method is based on local suction of the tion of the CBL will cease locally for areas CBL below the desk (Figure 9.18c). In where due to local radiant cooling the cloth- both cases the transport of pollution from ing surface temperature has decreased to the lower level is terminated and new CBL the level of the surrounding air temperature is developed in front of the body above the (Melikov 2015). The CBL will therefore be desk level. Closing the gap between the ta- weakened (Figure 9.19 middle). When the ble and the abdomen blocks the CBL that clothing surface temperature becomes has developed over the legs and reduces the lower than the surrounding air temperature, maximum velocity of the CBL (developed local downward flows opposing and dis- above the board) at the level of the mouth turbing the main upward flow of the CBL from 0,17 m/s to 0,11 m/s (Licina et al. may occur (Figure 9.19 right). As a result, 2014). the transport of clean air from near the floor to the breathing zone will stop and more of the surrounding polluted air will be inhaled. The CBL and clothing are in continuous contact. Clothing made of deodorant materials can be used to clean and disinfect the air of the CBL and thus to improve inhaled air quality (Melikov 2015).

Figure 9.18. Control of the CBL in front of sit- The thermal plume above a person has an ting person: a) without control, b) passive con- upward velocity of approx. 0,25 m/s (25 cm trol; c) active control (Bolashikov et al. 2010). above head), and this flow often prevents draught at head height. Figure 9.20 shows the thermal boundary layer around a standing thermal manikin with size and heat production as an “average” person. The free convection flow around the manikin is vis- ualized by smoke. The interaction between the thermal plume generated by the manikin and a downward flow with different velocities is demonstrated with six photo- graphs. The boundary layer is preserved Figure 9.19. Effect of local radiant cooling on with a downward velocity lower than the development of the CBL: a) strong CBL; b) 0,25 m/s (Nielsen 2009). A similar result is weak CBL when the local clothing surface temperature is equal to the room air temperature; reported by Licina et al. (2015c). The direc- c) Downward CBL establishes locally when the tion and magnitude of the surrounding air- clothing surface temperature is lower the room flows considerably influence the airflow temperature (Melikov 2015). distribution around the human body. Downward flow with velocity of 0,175 m/s The combination of the DV with other does not influence the convective flow in methods for generating high quality indoor the breathing zone, while flow at 0,30 m/s environment should be considered care- collides with the CBL at the nose level re- fully because it may have negative effect on ducing the peak velocity from 0,185 to

0,10 m/s. Transverse horizontal flow from velocity in the breathing zone and changes the front disturbs the CBL at the breathing the flow pattern around the body, com- zone even at 0,175 m/s. In case of a mani- pared to the assisting flow of 0,175 m/s or kin sitting on a chair, airflow from below quiescent conditions. In this case, the air- (assisting the CBL) with velocity of be- flow interaction is strongly affected by the tween 0,30 and 0,45 m/s reduces the peak presence of the chair.

Figure 9.20. A thermal manikin 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.

10 References

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Single user license only, copying and networking prohibited. All rights reserved by REHVA. Single user license only, copying and networking prohibited. All rights reserved by REHVA. REHVA Guidebooks: No 1 Displacement Ventilation in Non-industrial Premises (Out of print) No 2 Ventilation Effectiveness No 3 Electrostatic Precipitators for Industrial Applications No 4 Ventilation and Smoking (Out of print) No 5 Chilled Beam Cooling No 6 Indoor Climate and Productivity in Offices No 7 Low Temperature Heating And High Temperature Cooling No 8 Cleanliness of Ventilation Systems No 9 Hygiene Requirement for Ventilation and Air-conditioning No 10 Computational Fluid Dynamics in Ventilation Design No 11 Air Filtration in HVAC Systems No 12 Solar Shading – How to integrate solar shading in sustainable buildings No 13 Indoor Environment and Energy Efficiency in Schools – Part 1 Principles No 14 Indoor Climate Quality Assessment No 15 Energy Efficient Heating and Ventilation of Large Halls No 16 HVAC in Sustainable Office Buildings – A bridge between owners and engineers No 17 Design of energy efficient ventilation and air-conditioning systems No 18 Legionellosis Prevention in Building Water and HVAC Systems No 19 Mixing Ventilation – Guide on mixing air distribution design No 20 Advanced system design and operation of GEOTABS buildings No 21 Active and Passive Beam Application Design Guide – For Global Application No 22 Introduction to Building Automation, Controls and Technical Building Management No 23 Displacement Ventilation

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