Disp ChaptersBook

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3 BASIC KNOWLEDGE ABOUTDISPLACEMENT VENTILATION

3.1 SummaryThis chapter presents the basics for calculating the temperature gradient andestimating the contaminant concentration ina displacement ventilated room. Main itemsin this chapter are: Air flow patterns Temperature distribution Convection flows

Contaminant distribution Thermal comfort

3.2 Conclusions The contaminant concentration is always

better in the occupied zone in adisplacement-ventilated room than in aroom ventilated by mixing ventilation.

Theoretically we need a supply air volume flow of 20 l/s per person to keepthe occupied zone free fromcontaminants. However due to the freeconvection around a person a smaller supply air volume flow gives a much

better air quality in the breathing zone. Asupply air volume flow of 10 l/s per

person gives e.g. a concentration that isonly 20% of the concentration in theambient at the same level.

The vertical temperature distribution hasto be given attention. Make sure that a

suitable diffuser is utilised in order toavoid cold air along the floor.

3.3 Principles of DisplacementVentilation

The air-flow pattern in a ventilated room ismainly divided into two different types,mixing (dilution) ventilation anddisplacement ventilation. In mixingventilation the air is supplied in such a waythat the room air is fully mixed and thecontaminant concentration is the same in thewhole room. In displacement ventilation,

which is the subject of this book, a stratifiedflow is created using the buoyancy forces inthe room. The air quality in the occupiedzone is then generally better than withmixing ventilation. The ventilation systemsupplying the air to the room is notconsidered in this book, only the air flowwithin the room.

Figure 3.1 Schematic illustration of the air flow that might be found in a roomventilated by displacement ventilation

Displacement ventilation has for many years been used in industrial premises with highthermal loads. Since mid 80s it has also

been used in non-industrial premises to alarge extent, especially in the Scandinaviancountries. In recent years the interest indisplacement ventilation has increased allover the world. Displacement ventilation

presents the opportunity to improve both thetemperature effectiveness and the ventilation

effectiveness. The principle is based on air density differences where the room air separates into two layers, an upper pollutedzone and a lower clean zone, see Figure 3.1.This is achieved by supplying cool air with alow velocity in the lower zone andextracting the air in the upper zone. Freeconvection from heat sources creates avertical air movement in the room. When theconvection heat sources in the room are alsothe contamination sources, the convectionflows transport the warm polluted air up tothe upper zone. The convection flow rates


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Rehva Displacement Ventilation Guidebook

10

relative to the ventilation flow ratedetermine the height of the boundary

between the two zones. The sum of thewarm convection flow rates to the upper zone minus the downward directed flows

from cold surfaces to the lower zone is equalto the ventilation flow rate in the room. Anincreased ventilation flow rate thus movesthe boundary upwards and a decreased flowrate moves the boundary downwards at fixedconvection flow rates.

3.4 Air flow patternIn a displacement ventilated room the air flow pattern is governed by the convectionflows from heat sources and sinks present inthe room. This means that a distinctivefeature of displacement ventilation is theformation of horizontal air layers. Thewarmest air layers are at the top and thecoolest air layers are at the bottom. The air moves easily within a horizontal layer butthe transportation between the layers needs astronger force. See Figure 3.2. This meansthat the extract should be positioned at thelayer in which the pollutions are or where

the highest temperatures occur. In mostcases this means that the extract should be inthe upper part of the room.

Figure 3.2 Horizontal air movement inconnection with the extract.

The vertical air movement is caused byconvection flows from warm sources or coldsinks. Warm objects such as people,computers, lamps etc. create risingconvection flows. Depending on the power and geometry of the heat source theconvection flows will rise all the way to the

ceiling or settle at a lower height see Figure3.3.

Figure 3.3 Vertical air movement caused byconvection.

The supply air temperature must be lower than the room air temperature, which isnormally given when there is a heat surplusin the room. If the supply air temperature iswarmer there will be a short-circuit, seeFigure 3.4. However the vertical air flow hasa certain amount of entrainment, whichcauses some circulation in the rest of theroom, this is sometimes used for heating anempty room before occupational time .

Figure 3.4 Short-circuit of air flow in aroom when the supply air temperature iswarmer than the room air temperature.

3.5 Temperature distributionSince displacement ventilation supplies coldfresh air directly to the occupied zone, a

potential draught risk exists at floor level. Inaddition, the temperature stratification maycause discomfort. See Figure 3.5. Thetemperature will, however, not vary much inthe horizontal direction, except close to thediffuser.


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3 BASIC KNOWLEDGE ABOUT DISPLACEMENT VENTILATION

11

H e i g h t a b o v e

f l o o r

[ m ]

0 0,2 0,4 0,6 0,8 1 1,20,0

0,5

1,0

1,5

2,0

2,5

Temperature ratio -

- s

se

Figure 3.5 Temperature stratification in adisplacement ventilated room.

3.5.1 Temperature at the floor

The temperature of the supply air in thefloor area rises due to induction andconvection, as radiation from the other warmer surfaces in the room in turn heatsthe floor. A dimensionless temperature of the air near the floor is often presented as

se

s f

= (3.1)

where: f = the air temperature near the floor s = the supply air temperature e = the exhaust air temperature

The total temperature difference givestogether with the air volume flow rate theamount of heat removed from the space:

( ) tot se pv cq = 310 (3.2)

where:qv = the volume air flow rate [l/s]

= the air density = 1,2 kg/mc p= the specific heat of the air =

1000 J/kgK tot = the heat removed from the space [W]

Based on a literature review (Mundt, 1990)the following equation can be used toestimate the dimensionless temperature of the air near the floor.

11110

13

+

+

=

cf r

pv

A

cq

(3.3)

where A = the floor area [m] = the heat transfer coefficient due to

radiation [ 5 W/mK] cf = the heat transfer coefficient at the

floor due to convection [ 4 W/mK]

In Figure 3.6 the dimensionless temperatureof the air near the floor is shown as afunction of the ventilation flow rate per m 2 floor area. The points shown in the figureare from measurements with distributed heatsources presented in eleven differentreferences (Mundt, 1996).

Ventilation flow rate per m floor area,q / A [l/sm]

0

0,2

0,4

0,6

0,8

1,0

0 1 2 3 4 5 6 7 8

= 5 W/mK

= 3 W/mK

cf

cf

=

( -

) / ( -

)

f

s

e

s

v

Figure 3.6 Dimensionless temperature of theair near the floor as a function of theventilation flow rate per m 2 floor area withdifferent heat transfer coefficients due toconvection.

3.5.2 Vertical temperaturedistribution

The vertical temperature distribution in theroom depends on the vertical location of theheat sources. When the heat sources are inthe lower part of the room the temperaturegradient is larger in the lower part and thetemperature more constant in the upper part.On the other hand, when the heat sources are


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12

located mostly in the upper zone, thetemperature gradient is smaller in the lower

part and increases in the upper part, seeFigure 3.7.

Heat sources in thelower part of the

room

Heat sources in theupper part of the

room


f l o o r

[ m ]

0 0,2 0,4 0,6 0,8 1 1,20,0

0,5

1,0

1,5

2,0

2,5

Temperature ratio -

- s

e

Figure 3.7 Temperature gradient in adisplacement ventilated room with the heat

sources at different levels.

For a given arrangement of heat sources, therelative temperature distribution is relativelyindependent of the heat load.

The temperature gradient is stronglyinfluenced by the elevation of the heatsources. In rooms where the heat sources arelocated at a high level, displacementventilation is efficient for keeping the

occupied spaces cool. See Figure 3.8.

However, the air temperatures near thefloor, f ,and the vertical temperaturegradient are not only a function of flow rateand load, they are also a function of the typeof heat source in the room.According to Nielsen (1996) and Brohus andRyberg (1999) the relative air temperaturenear the floor, (see equation 3.1) varies

between 0,3 and 0,65 for different types of heat sources. See Figure 3.9.

A concentrated heat load as e.g. a smallfurnace in an industrial environment cangive a value of 0,3. Ceiling light will givea vertical temperature gradient with a floor temperature of = 0,5, which is generated

by radiation from the light source. When persons are the primary heat source willhave a value of 0,58, and evenly distributed

heat sources will give a value of 0,65. It isobvious that this variation can be of thesame magnitude as the one found at differentflow rates.

H e i g h

t a b o v e

f l o o r

Temperature

Figure 3.8 Roof heated by sun - an example where displacement ventilation is efficient.


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Distributed heat sourcesSedentary persons

Ceiling light Point heat source

H e i g h

t a b o v e

f l o o r

0 0,3 0,5 0,65 10,58

Temperature ratio -

- s

se

Figure 3.9 Vertical temperature distribution

for different types of heat loads.

The different temperature gradients areshown in Figure 3.9 where it is assumed thatthe vertical temperature distribution is alinear function of the height. If manydifferent heat sources are present in theroom it is advised to use the 50% rule(Chapter 3.6).

3.5.3 Temperature effectivenessAs the exhaust temperature is higher thanthe air temperature in the occupied zone, atemperature effectiveness can be defined:

soz

se

= (3.4)

where oz = the mean temperature in the occupied

zone

3.6 Practical assumptions for the temperature distribution

As shown in Figure 3.5 and Figure 3.7, thetemperature increases with height, and thetemperature profile depends on the locationof the heat sources and the flow rate. For most practical purposes, we may assume atemperature profile as shown in Figure 3.10.

Extract air temperature,

Supply air

temperature,

Air

temperatureat floor,

50% 50%

Temperature

H e i g h

t a b o v e

f l o o r

s f

e

Figure 3.10 The "50%-rule" for vertical temperature distribution.

The 50%-rule for the vertical temperaturedistribution says that the air temperature atfloor is half-way between the supply air temperature and the extract air temperature.This is a general experience that may beused as a first approximation for mostnormal rooms and normal air diffusers.

Example: If the heat balance and air flow rate in theroom yields a temperature increase of e - s = 10K , then the temperature at floor level will become approximately 5K higher than the supply air temperature.

3.7 The Archimedes number Several phenomena in a ventilated room,like the vertical temperature gradient,velocity levels in stratification flow,stratification level and ventilationeffectiveness can all be described byArchimedes number. The Archimedesnumber is simply a ratio between the

buoyancy forces and the inertia forces. Inits original form it is defined as:


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2v L g

Ar = (3.5)

where: = density difference between the

colder and the warmer air [kg/m] g = acceleration of gravity = 9,81 m/s L = a characteristic length [m] = air density [kg/m]v = air velocity [m/s]

The Archimedes number can be expressed ina number of ways, using temperaturedifferences to express density differencesetc. But the basic fact is always the same: Larger numbers mean that the buoyancy

forces are dominant Smaller numbers mean that inertia forces

(velocities) are dominant

3.8 Convection flows theengines of displacementventilation

Figure 3.11 Convection flows - the engine of displacement ventilation.

Natural convection flows are the engines of

displacement ventilation. A naturalconvection flow is the air current that risesabove warm objects like people or computers, rises along a warm wall, or descends from cold objects like windows or outer walls, due to buoyancy. See Figures3.11 - 3.13. To understand displacementventilation, one has to understand the natureof the natural convection flows, and to knowthe magnitude of these flows. Theconvection flow rising above a hot object is

called a thermal plume, or simply a plume.Empirical, analytical and computational

fluid dynamics are the commonly usedapproaches to evaluate air temperatures,velocities and air flow rates in thermal

plumes above different heat sources andconvection flows at vertical surfaces.

All plumes encountered in practicalventilation are turbulent flows, and followthe similarity laws for fully turbulent flows.

The amount of air in the convection flowsincreases with height due to entrainment of the surrounding air. The amount of air transported in a natural convection flowdepends on the temperature and thegeometry of the source and the temperatureof the surrounding air. As the driving forcein convection flows is the buoyancy forcecaused by the density difference (i.e. thetemperature difference) a temperaturegradient in the room influences the plumerise height.

Hot wall >

Cold wall <

su

su su

su

Flowqv

Flowqv

Figure 3.12 Convection flows at vertical surfaces.


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z

Flowqv

Figure 3.13 Thermal plume above ahorizontal source.

3.8.1 Point and line sourcesThermal plumes above point and linesources (Figure 3.14) have been studied for many years. Among the earliest publicationsare those from Zeldovich (1937) andSchmidt (1941). Turner (1973) gives acomprehensive record of most of the

phenomena encountered in connection with buoyancy effects in fluids. Analyticalequations to calculate velocities,temperatures and air flow rates in thermal

plumes over point and line heat sources withgiven heat loads were derived based on themomentum and energy conservationequations and assuming Gaussian velocityand excessive temperature distribution inthermal plume cross-sections (Mundt, 1996).These equations correspond with those

received experimentally by other

researchers (Mierzwinski, 1981, Popiolek,1981) and are listed in Table 3.1. Theequations in Table 3.1 were derived with theassumption that the heat source size wasvery small and did not account for the actual

source dimensions.

Point source Line source

Figure 3.14 Plumes from a point source and from a line source.

The coefficients in the equations differ slightly in different references depending onthe entrainment coefficients used. is the

convective heat flux in W or W/m from theheat source and z is the height above thelevel of the heat source. The convective heatflux can be estimated from the energyconsumption of the heat source tot by

= k tot (3.6)

The value of the coefficient k is 0,7-0,9 for pipes and ducts, 0,4-0,6 for smaller components and 0,3-0,5 for larger machinesand components (Nielsen, 1993 B).

Table 3.1 Characteristics of thermal plumes above point and line sources.

Parameter Point source Line sourceCentreline velocity, v z [m/s] v z = 0,128 1/3 z 1/3 v z = 0,067 1/3 Centreline excessive temperature, z [K] z = 0,329 2/3 z 5/3 z = 0,094 2/3 z 1 Air flow rate, qv,z [l/s for point source, l/sm for line source] qv,z = 5

1/3 z 5/3 qv,z = 13 1/3 z


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3.8.2 Convection flow alongvertical and horizontalsurfaces

Convection flow along vertical surfaces is

also of major interest. When the verticalextension of the surface is small theconvection flow is mainly laminar and atlarger extensions the flow is turbulent. The

basic equations for a surface with aconstant temperature are given in Table 3.2(Jaluria, 1980, Etheridge and Sandberg,1996).

is the temperature difference betweenthe surface and the surrounding air and z isthe height from the bottom of the surface.The flow changes from laminar toturbulent at GrPr =710 8, which for air andmoderate temperature differences meansaround z = 1 m and for air at higher temperatures around z = 0,5 m.

Convection flows from horizontal surfacesare very difficult to determine in the same

basic way as for point, line or verticalsources. The reason is that the flows

behave in a very unstable way and leavesthe flat surface from different positions atdifferent times, partly depending on thetotal air movement in the room. Thesesurfaces are mostly treated as plumes fromextended sources see chapter 3.8.3.

3.8.3 Extended sourcesIn reality heat sources are seldom a point, aline or a plane vertical surface. The most

common approach to account for the realsource dimensions is to use a virtual sourcefrom which the air flow rates arecalculated (Elterman 1980, Mundt 1992,Skistad

1994), see Figure 3.15. The virtual originis located along the plume axis at adistance z 0 on the other side of the realsource surface.

b) Extended source

Virtualsource

z

a) Point source

z 0

Flowqv

Figure 3.15 Illustration of the position of the virtual source

The adjustment of the point source modelto the realistic sources using the virtual

source method gives a reasonable estimateof the air flow rate in thermal plumes.

The weak part of this method is how toestimate the location of the virtual located

point source. The method of a "maximumcase" and a "minimum case" provides atool for such estimation. See Figure 3.16(Skistad 1994). According to the"maximum case", the real source isreplaced by the point source such that the

border of the plume above the point source passes through the top edge of the real

Table 3.2 Characteristics of convection flows along vertical surfaces

Parameter Laminar region Turbulent regionMaximum velocity, z v [m/s] z v z = 1,0 z v z = 1,0

Thickness of boundary layer [m] 25,025,005,0 z = 7,01,011,0 z =

Air flow rate, z vq , [l/sm width]75,025,0

, 87,2 z q z v = 2,14,0

, 75,2 z q z v =


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source (e.g., cylinder). The "minimum case"is when the diameter of vena contracta of the

plume is about 80% of the upper surfacediameter and is located approximately 1/3diameter above the source. The spreading

angle of the plume is set to 25. For the low-temperature sources, Skistad (1994)recommends the "maximum case", whereasthe "minimum case" best fits themeasurements for larger, high temperaturesources. The maximum case gives z 0 =2,3 D and the minimum case z 0 = 1,8 D with z 0 defined in Figure 3.16.

For a flat heat source Morton (1956)suggests the position of the virtual source to

be located at z 0 = 1,7-2,1 D below the realsource. Mundt (1996) calculates thethickness of the boundary layer (see Table3.2) at the top of a vertical extended heatsource and adds this to the source radii andthen calculates the position of the virtualsource as z 0 = 2,1( D+2 ) before using the

point source equation. According to Bach etal (1993) the volume flow from the verticalsurfaces should be added to the volume flowcalculated by the equations for point or linesources.

Minimumcase

Maximumcase

D

z

H

h

D

z

H

h D/3d 0 d 0

z 0 z

0

Figure 3.16 Convection flow above avertical cylinder

ExampleCalculate the convection flow rate 0,5 mabove a cylinder with height 1 m anddiameter 0,4 m. The convective heat flux is50 W.

In the maximum case we get

( ) m9,0255,25,12tan20 === D D z o

and

m4,15,09,00 =+=+= h z z

from Table 3.1 we use

3531, 5 z q z v =

which gives

s/l324,1505 3531, == z vq

In the minimum case we get

( ) m72,0804,15,12tan28,00 === D D z o

and

m09,15,03,072,030 =+=+= h D z z

which givess/l2109,1505 3531, == z vq

(The position of the virtual source is in this

case ( ) D D=

47,131804,1 below theupper edge of the source)


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3.8.4 Plume interactionWhen a heat source is located close to a wallthe plume may be attached to the wall,Figure 3.17. In this case the entrainment will

be reduced compared to the entrainment in afree plume. The air flow rate from a heatsource can then be calculated as half of theflow from a source with a heat emission of 2 (Nielsen, 1993 B).

( ) 35313531, 2,32

25 z

z q z v

=

= (3.7)

a) Plume attachedto a wall

b) Interaction betweentwo plumes

Figure 3.17 Thermal plumes.

If the heat source is located in a corner theair flow rate is equal to 25% of the air flowfrom a heat source with a heat emission of 4 (Kofoed, 1991)

3531, 2 z q z v = (3.8)

When several heat sources are positionedclose to each other the plumes merge into asingle plume, see Figure 3.17 . The total flowfrom N identical sources is then given by,(Nielsen, 1993 B)

z v N z v q N q ,31

,,= (3.9)

where

qv, z = the flow in the plume from one of thesources

When the heat sources are more separatedthe total flow is equal to the sum of the

flows from each heat source.

3.8.5 Plumes and temperaturegradients

When there is temperature stratification in aroom, like in a room ventilated bydisplacement ventilation, the plumes areinfluenced by the temperature stratification.The driving force for the plume is thetemperature difference between the plumeand the surroundings and when thisdifference diminishes the plumes willdisintegrate and spread horizontally in theroom, see Figure 3.18.

Plume1

Plume2

Plume3

Plume 1

Plume 2 Room

Figure 3.18 Schematic illustration of the air flow pattern in a room ventilated bydisplacement.

Batchelor (1954) noticed the influence of atemperature gradient in the surroundings andMorton et al (1956) gave a solution for calculating the maximum plume rise from a

point source in surroundings with atemperature gradient. The volume flow ratesin the plumes in a room with temperaturestratification is slightly decreased comparedto the volume flow rates calculated with theequations presented for a non stratifiedmedia, Mundt (1992). Jin, (1993) studiedthe maximum plume rise height for plumesabove welding arcs.


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In the presence of a temperature gradient,the convective plume reaches theequilibrium height ( z t ) where thetemperature difference between the plumeand the ambient air disappears, see Figure

3.19. Also there is another level in the plume, where the air velocity equals to zero.This is referred to as the maximum height of the plume ( z max ).

z*

2,1

2,8

z**

2,0

2,95

Pointsource

Linesource

s = > 0d dz

z t z max

Figure 3.19 Vertical plume in a room withtemperature gradients and stratification

The plume spreads horizontally betweenthese two heights. The convective flow

below z t can be calculated from thefollowing model (Mundt, 1996).

Point sourceThe position of the virtual source iscalculated. A dimensionless height z * abovethe virtual source is calculated

4/18/3*

86,2

=cf s z z (3.10)

where: s = vertical temperature gradient ( / z)

in the room [K/m] cf = convective heat from the source [W]

As can be seen from Figure 3.19 only z * values less than 2,1 are relevant to further calculations. The volume flow rate at theheight z * is then given by

1854338,2 Z sq cf v =

with

3*2**1 062,0380,0039,0004,0 z z z Z ++=

(3.11)

where: qv = the volume flow rate in l/s

The maximum height z max is given byEquation (3.10) for z * = 2,8

8341max 98,0

= s z cf (3.12)

and the height z t by Equation (3.10) for z * = 2,1

834174,0 = s z cf t (3.13)

Line sourceThe position of the virtual source iscalculated. A dimensionless height z ** abovethe virtual source is calculated

3/12/1** 78,5 = cf s z z

(3.14)where:

s = vertical temperature gradient ( / z inthe room [K/m]

cf = convective heat from thesource [W/m]

As can be seen from Figure 3.19 only z ** values less than 2,0 are relevant to further calculations. The volume flow rate at theheight z ** is then given by

22132

, 82,4 Z sq cf l v =

with3**2****

2 018,0029,0477,0004,0 z z z Z ++= (3.15)

whereqv, 1 = the volume flow rate in l/(s m)


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The maximum height z max is given byEquation (3.14) for z **=2,95

2/13/1max 51,0

= s z cf (3.16)

and the height z t by Equation (3.14) for z **=2,0

213135,0 = s z cf t (3.17)

3.8.6 Convection flows from realobjects

Personalcomputer 75W

Fluorecentlamp36W

Desk lamp60 W

0,3 0,5 1,0 1,2 1,4Height above object, z [m]

3

5

10

30

50

80

Height above floor, z [m]

10

30

50

80100

20

200

1,0 2,0 3,0 4,0 5,0

Vertical temp.gradient:

s = 0,3 C/m

s = 0,09 C/m

Equation, Table 3.1

C o n v e c t

i o n

f l o w r a

t e , q

[ l / s ]

v z

C o n v e c t

i o n

f l o w r a

t e , q

[ l / s ]

v z

Figure 3.20 Convection volume flow at normal room temperatures above a

sedentary person and above some objects. From Mundt, 1992/Nielsen, 1993 B.

From the theories above and practicalexperiments, Nielsen (1993 B) hassummarised the convection flows abovesome common objects found in non-industrial environments, see Figure 3.20.

The line drawn in the figure to the left iscalculated by the equation for the air flowrate in Table 3.1. The convection flow abovea sitting person is thus approximately 20 l/s,see Figure 3.21. In order to keep the inhaledair at a lower concentration than the ambienta lower air flow may however be used incalculations, see Chapter 3.10.

H e i g h

t a b o v e

f l o o r

[ m ]

0

0,5

1,0

1,5

2,0

2,5

s = d /dz = 1,5 C/m

q = 20 l/svz

Figure 3.21 Convection flow in plume abovea sedentary person in a normal environment.

3.9 Contamination distributionThe contamination distribution in adisplacement-ventilated room depends onthe position of the contamination sourcesand if the heat sources are also thecontamination sources. In the ideal case withwarm concentrated contamination sources

all contaminants are transported directly intothe upper zone by the convection flows, seeFigure 3.22. If the contamination sources arecold and evenly distributed at the floor, thecontamination distribution will be like thetemperature distribution (see Figure 3.10)according to Krhne and Fitzner (1995).However if the source is too weak, the

plume might disintegrate at a lower leveland the contaminants will then be trapped atthis level, see Figure 3.23, and only slowly

transported indirectly by the stronger convection flows to the upper zone.


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21

0 0,2 0,4 0,6 0,8 1,0

H e i g h

t a b o v e

f l o o r , z

[ m ]

0

0,5

1,0

1,5

2,0

2,5

Contamination ratio, c /croom e

Figure 3.22 Schematic illustration of the contamination distribution in a room ventilated bydisplacement ventilation and with warm contaminant sources.


f l o o r , z

[ m ]

Temperature,

0

0,5

1,0

1,5

2,0

2,5

Contamination, c room

room plume 1

plume 2

croom

Figure 3.23 Schematic illustration of the contamination distribution in a room ventilated bydisplacement ventilation, when the contaminant source (the person) is not the warmest

source.

The contaminant concentration is of coursealso influenced by the downward directedconvection flows that might occur at theouter walls in cold seasons, especiallywhen the walls are poorly insulated. Thesedownward flows will then transport thecontaminants from the upper zone back tothe lower zone. However as long as thereis a positive concentration gradient in theroom, the contaminant concentration in the

occupied zone will always be lower than by mixing ventilation.

The influence of a poorly insulated roof will, in the cold season decrease theconcentration gradient, due to the downfallof cold air, just like with the cold walls(See Figure 3.24). However if the roof isheated by the sun this will help stabilisethe displacement ventilation as it heats theair in the upper zone. (See Figure 3.8).


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Figure 3.24 Poor building air tightnessand insulation may reduce the benefit of displacement ventilation, and make it morelike mixing ventilation.

3.10 Ventilation effectivenessDifferent definitions of ventilationeffectiveness have been introduced. Indefining ventilation efficiency, adistinction must be made between twoterms: the contaminant removal effectiveness ,

c, which is a measure of how quicklyan airborne contaminant is removedfrom the room (Brouns and Waters,1991) and

the air change efficiency, a , which is ameasure of how quickly the air in theroom is replaced (Sutcliffe, 1990).

In a displacement ventilated room the air change efficiency is mostly higher ( a 60-

70 %) than in a room ventilated by mixingventilation ( a 50 %), (Mundt, 1994). Agood survey of the relation between thedifferent versions of ventilationeffectiveness is given by Nielsen (1993B),

pp. 17 19. The most relevant versions of ventilation effectiveness for displacementventilation in non-commercial premises aretreated below.

3.10.1 Contaminant removaleffectiveness

The contaminant removal effectiveness isdefined by:

smean

sec

cccc

= (3.18)wherece = the contaminant concentration in

the exhaustc s = the contaminant concentration in

the supplycmean = the mean contaminant concentration

in the room

or for the occupied zone:

soz

sec

cccc

= (3.19)

wherecoz = the mean contaminant

concentration in the occupied zone

3.10.2 Personal exposure index.Thermal flow around a person and flowgenerated by the movement of a personmay give an inhaled concentration that is

different from the concentration in headheight if the measurements are madewithout a person.

This can be expressed by the following personal exposure index, Brohus and Nielsen (1996 A):

s

se

cccc

=

expexp (3.20)

wherecexp = the inhaled concentration.

It is possible to work with a stratificationheight that is lower than the height of the

breathing zone. The personal exposureindex will often be larger than the localventilation index because clean air ismoved from the lower part of the room upto the breathing zone by the free-

convection boundary layer around the person, see Figure 3.25 and Figure 3.26.


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23

Figure 3.25 Thermal flow around a personmay give cleaner breathing air.

Measurements of the personal exposureindex made in situations with air movement in the occupied zone andcontaminant sources close to a person cangive rise to a very small exposure index,see Brohus and Nielsen (1996 B).

Figure 3.26 Iso-concentration map showing the dispersion pattern of a tracer gas emitted directly above a 4 W heat source in the lower zone. (Stymne et al,1991)

Although the personal exposure indexshows the ability of improved air quality inthe inhaled air displacement ventilationshould not be used when the contaminationsources are mostly cold.

Flow rate

litre/s person

roomvolh

0,0

0,2

0,3

0,4

0,5

0,6

0,1

0 1 2 3 4

0 5 10 15 20

C o n c e n t r a

t i o n

i n t h e a m

b i e n

t ( )

C o n c e n t r a

t i o n

i n t h e

b r e a

t h i n g z o n e

( )

=

1 e x p

Figure 3.27 The ratio between theconcentration in the breathing zone and inthe ambient air at the same height (Etheridge and Sandberg, 1996)

As pointed out above, the ventilation flowrate must not always be set to cover theconvection flows above the persons

present in a room. Figure 3.27 shows theimprovement in inhaled air quality relativeto the air quality in the ambient as afunction of the ventilation flow rate per

person.

With a ventilation flow rate of 20l/(s,person) the border is above the person.A ventilation flow rate of 10 l/(s,person)gives however a concentration which isonly 20% of the concentration in theambient at the same level. Measurements

by Mundt (1994) also showed the rapidalmost instantaneous recreation of thethermal flow around a person when the

person moves from one place to another ina room.


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24

Particle transportation in a displacement-ventilated room was studied by Mundt(2000), the results indicate that there seemto be little risk of resuspension of particlesfrom the floor into the supply air flow. The

sizes studied were however only particleslarger than 0,5 m and more research isneeded for smaller particles

3.11 Thermal considerationsOne of the limiting factors for the thermalcomfort in displacement ventilation is theair velocity in the floor area. There is anear zone close to the air supply devicewhere the velocities will be greater thanwhat is recommended, 0,15 m/s in winter time and 0,25 m/s in summer time (ISO7730). The extent of this zone is dependingon the air supply device and the thermalload, and should be presented in cataloguesfrom the manufacturer.

The other limiting factor is the temperaturegradient, which should be less then 3 K/m

between 0,1 m and 1,1 m above the floor (ISO 7730). In some countries the limit is

set to 2 K/m.

3.12 Room heatingRadiators

Radiator

Figure 3.28 Radiator well suited for displacement ventilation.

Using radiators is a good method for roomheating with displacement ventilation. The

radiators should preferably be located below the coldest elements in the room, i.e.the windows and the outer walls. Thelarger part of the heat emission is theradiation. The minor part is convection,

which will counteract the colddowndraught from windows and coldwalls.

Convectors

Convector

Figure 3.29 Convector ok when located below the cold walls or windows.

Convectors placed below windows goeswell with displacement ventilation, whenthe heat is distributed along the coldwalls/windows. One concentratedconvector may case mixing of the roomair.

Heating by ceiling panels Heating by ceiling panels is very suitablefor displacement ventilation. In normalconditions, without heating demand, theceiling is 3 4K warmer than the floor,yielding a heat transfer from ceiling tofloor of about 20 W/m. Thus, a slightincrease in ceiling temperature will

provide sufficient heat for room heating.The convection part of the heat from theceiling panels will counteract the heat lossthrough the ceiling.

Heated ceiling panels stabilises the thermalstratification, and thus benefitdisplacement ventilation.


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25

Heating panels

Figure 3.30 Heating by ceiling panels.

Floor heating In normal conditions, floor heating is wellsuited for displacement ventilation. Part of

the heat transmission from the floor isradiation towards the cold surfaces of theroom. The convective heat transfer willheat the supply air that spreads across thefloor.

If the floor is too hot, it will heat the air and make it rise so that it causes mixing, atleast in the lower part of the room.However, practice has shown that with afloor temperature below approximately25C and the supply air being some 2K or more colder than the room air, the supplyair spreads along the floor.

Heated floor Figure 3.31 Floor heating goes well withdisplacement ventilation.

Heating the room by warm air is notrecommended. The warm supply air willrise, and cause mixing of the room air. Itmay also cause short-circuiting of theventilation air, as shown in Figure 3.4.

Warmsupply air

Figure 3.32 Warm air supply not recommended for displacement ventilation

Room heating by the supply air is used tosome extent to heat the rooms in themorning before personnel enter the room.When personnel has entered the room, andwork starts, the heat gain from persons andequipment normally outweighs the coolingfrom the surroundings.


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41

5 DESIGN PROCEDURES

5.1 SummaryThe design of a ventilation system shouldalways follow a systematic procedure asfollows: First, choose a suitable ventilation

principle. (Displacement ventilation isnot always the best for all purposes!)

If displacement ventilation is chosen,calculate the required ventilation air flowrate with regard to air quality andtemperature conditions.

Select suitable diffusers with regard tovertical temperature distribution andadjacent zones.

5.2 Strategic design of theroom air conditioningprocess

a) Target levels The aim of the room air conditioning is tomaintain desired conditions, i.e. target

levels, in the room during different operating

conditions in the most economical way(energy usage, cost efficiency). Depending

on the design criteria the designer hasdifferent strategies to choose from in order to achieve specified targets. The room air conditioning design and evaluation processis illustrated in Figure 5.1.

b) StrategyThe room air conditioning strategy is afundamental scheme that describes thetargeted temperature, humidity andcontaminant distributions as well as air flow

patterns within the air-conditioned room.The room air conditioning system consistsof different methods and their controls thatall together create the system performance.The system performance is evaluated bycomparing the achieved conditions to thechosen strategy. Both the methods (room air distribution, exhaust, room heating andcooling, etc.) and processes and disturbancesinside the room influence the resultingconditions.

a) TARGET OF THE

INDOOR AIR CONDITIONS

a) TARGET OF THE

INDOOR AIR CONDITIONS

c)AIR CONDITIONINGSYSTEM & CONTROL

c)AIR CONDITIONINGSYSTEM & CONTROL

AIR CONDITIONINGSYSTEM PERFORMANCE

VALUATION

AIR CONDITIONINGSYSTEM PERFORMANCE

VALUATION

ROOMINDOOR / OUTDOOR

LOADSDISTURBANCES

ROOMINDOOR / OUTDOOR

LOADSDISTURBANCES

Temperature [C]Humidity [%RH]Air quality [ppm]

Investment cost [ ]Running Cost [ ]

Temperature [C]Humidity [%RH]Air quality [ppm]

Investment cost [ ]Running Cost [ ]

oC oC

b)ROOM AIR CONDITIONING

STRATEGY

b)ROOM AIR CONDITIONING

STRATEGY

Figure 5.1 The Room Air conditioning and Evaluation Process. (Hagstrm 2000)


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42

PISTON STRATIFICATIONZONING MIXINGStrategy

Description Unidirectionalflow through theroom

Utilise densitydifferences

Air flow fromclean zones tocontaminatedzones

Uniformconditions in all

parts of theroom

Air quality;temp, humidity RF contaminants, c

, RF, c

Room dimension

, RF, c

Room dimension

, RF, c

Room dimension

, RF, c

Room dimension

s = supplye = exhaust

s

e

s

e

s

e

s

e

Maincharacteristics

Flow patterncontrolled by lowmomentum supplyair, strong enoughto overcomedisturbances

Flow patterncontrolled by

buoyancy

Flow patterncontrolled byhigh momentumsupply air

Flow patterncontrolled partly

by buoyancy and partly by supplyair momentum

Ventilationeffectiveness

1

DISPLACEMENT

- = -

oz s

e s c = c - cc - c

oz s

e s

Figure 5.2 The summary of the ideal room air conditioning strategies.(Hagstrm 2000)

c) SystemThe room air distribution method is oftenconsidered as a principal parameter to applya certain room air conditioning strategy andheating and cooling as assisting methods.However, it must be noted that in somecases a strategy can be fulfilled also withoutany mechanical air distribution installationsusing buoyancy forces. The classification of ideal room air conditioning strategies issummarized in Figure 5.2. Note that pistonflow requires large amounts of air. Fitzner (1996) points out that piston flow from thefloor and upwards exists for Archimedesnumbers less than 360:

3602 360, buoyancy forces will

dominate, and make a thermally stratifiedflow.


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5 DESIGN PROCEDURES

43

5.3 Ventilation and room air conditioning strategies

Displacement is an efficient air distributionmethod when:

the aim is air quality in rooms wherethe contaminants are warm,

large heat surpluses are required beremoved by large quantities of air (more than about 60 - 70 W/m 2 or more than around 10 l/sm 2 (36m3/hm 2)).

The design criteria for these cases differ andthey are discussed later in this chapter.

It is necessary to emphasize the difference between the room air conditioning systemand the air distribution method. Choosingdisplacement ventilation as an air distribution method does not by itself resultin a stratification strategy, if the whole roomair conditioning system is not designed for that purpose. As an example overheatedsupply air through displacement units resultsin close to mixed conditions (Halton Oy,

(2000)). Thus, it is possible to usedisplacement ventilation for example pre-heating of the space in the morning.However, due to the short-circuiting effect,constant heating of a room by hot ventilationair should not be used in connection withdisplacement ventilation.

Cooled ceiling panels combined withdisplacement ventilationAnother example is a system consisting of displacement ventilation and cooled ceilings.Low velocity air supply and cooled ceilingsystems behave like mixing systems whenthe cooled ceiling provides a substantial partof the cooling. See Figure 5.3 (Tan (1998) etal.)

0,0

1,0

1,5

2,5

H e i g h

t a b o v e

f l o o r

l e v e

l , z

[ m ]

0,8 1,0 1,2 1,4

Relative air temperature(relative to temp. at 0,1 m above the floor)

= 0

= 0,4

= 0,5

= 0,6

Cooled ceiling

2,0

0,5

= ratio of the cooled ceiling cooling output to the total cooling output (Tan 1998)

Figure 5.3 Vertical air temperaturedistribution in a room with cooled ceiling.Temperatures relative to temp. 0,1 metreabove the floor. Tan (1998)

The stratification effect decreases graduallyas the relative cooling load of the ceiling, ,increases. When the ratio is less than about0,6 there still is some thermal stratification

in the room (Figure 5.3). Similar type of behaviour hasalso been found with the contaminantstratification (Krhne 1995). This is shownin Figure 5.4.

0,0 0,2 0,4 0,6 0,8 1,00,0

0,2

0,4

0,6

0,8

1,0

1,2

/A = 12 W/m

/A = 20 W/m /A = 30 W/m /A = 50 W/m /A = 65 - 93 W/m /A = 50 W/m

Heat surplus per unit floor area

Relative cooling load of cooled ceiling, C o n

t a m

i n a t

i o n r a

t i o

i n o c c u p

i e d z o n e , c

/ c o z

e

Figure 5.4 Contamination ratio in occupied space versus cooling effect from ceiling panels. (Krhne, 1995).


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5.4 Factors influencing thethermal stratification andthe design methods

While the contaminant stratification level is

mainly affected by the relation of supply air flow rate and convective air flow rate,thermal stratification is also affected bythermal radiation exchange betweendifferent room zones. The thermal radiationfrom upper zone warms up the air temperature at floor level.From this fact, it follows that if the supplyair flow rate in the room is decreased whichleads to an increase in the temperaturestratification and in the ceiling temperature.

This implies that the thermal radiation fromupper zone to lower zone will also increaseand thus increase the air temperature at thefloor level. This in turn will decrease thetemperature stratification. This process has

been presented by Mundt (1996) in her doctoral thesis. When the verticaltemperature gradient has reached itsmaximum, the temperature in the wholeroom will start to rise. This is demonstratedin Figure 5.5.

The first displacement ventilation design

methods applicable for manual calculationsare based on the empirical coefficients, inwhich the influence of the thermal radiationexchange between upper and lower part of the room is built in. Such methods are

presented as an example by Halton (2000)and Skistad (1994). The value of thesemethods is their ease of use and also theaccuracy of the estimation which in manycases is still reasonable.

More detailed methods allowingcomputational treatment of radiationexchange and situations beyond thetraditional cases have been presented byLivtchak (2001) and Mundt (1996).However, these methods are iterative andtoo complex to be used manually, and needto be coded into software.

It is also possible to use computational fluiddynamics (CFD) software to simulate large,complex spaces. However, one needs to payspecial attention to description of radiationexchange and the right interpretation of

boundary conditions in heat and contaminantsources and also in supply air units.

17 19 21 23 250

0,5

1,0

1,5

2,0

2,5

Temperature, [C]

H e i g h

t a b o v e

f l o o r , z

[ m ]

16 18 20 22 24 26

9AM 10AM 12AM 1 PM

Supply air temperature, s = 17.8 C

Rule-of-thumbcurve

Figure 5.5 Temperature profiles measured at various times during a meeting in a room withconstant supply air temperature. (Skistad 1994).


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5 DESIGN PROCEDURES

45

5.5 Displacement VentilationDesign Procedure

5.5.1 Air quality: Design criteria for

contaminant stratificationThe design criterion for the air quality baseddesign is that the supply air flow rate isequal to the sum of convective flows at thestratification height (shift zone). Moreover,it should be ensured that any contaminantsthat are carried upwards by the convectiondo not re-circulate into the occupied zone.Once the required supply air flow rate isdefined it is necessary to check that both thecontaminant concentration and the thermalconditions requirements are satisfied withinthe occupied zone.

It must be noted that the verticalstratification of the contaminants occursonly when the contaminant source is insidethe warm convective current or thecontaminant is lighter than air. If the heatand contaminant sources are separate thereis a risk that contaminants are not carried out

from the occupied zone.

5.5.2 Temperatures: Design criteriafor thermal comfort

The design criteria for temperature-baseddesign are the removal of excess heat fromthe occupied zone and thermal comfort.Thus, the supply air flow rate is not chosen

based on the convective flows but on: occupied zone temperature requirement

(Minimum temperature at floor level andmaximum temperature at the edge of theoccupied zone)

maximum vertical temperaturedifferential within the room.

5.5.3 Design Procedure flow chartThe displacement ventilation design

procedure taking into account bothcontaminant and temperature stratification is

presented as a flow chart in Figure 5.5. Theapplication of the design procedure isdemonstrated with practical examples inchapter 8. The following notes apply to thechart:

S1: Typically stratification layer isselected slightly above the breathingzone.

S2: Take into account both ascending anddescending air currents.

S5: According to (Nielsen 1993) the lower (occupied) zone concentration is 0.1-0,3 times the exhaust air concentration. Using a conservativeestimate of 0,3 it can be checkedwhether the occupied zoneconcentration is below acceptablelevel. If the occupied zoneconcentration is higher than required,

then increase the supply air flow rateaccordingly.

T1: -Occupied zone temperaturerequirement ( qmin at floor level andqmax at the edge of the occupied zone)-Maximum temperature gradient

T3: Use the comfort criteria:a) The vertical stratification iscalculated by multiplying the

maximum temperature gradient withthe room height. b) Estimate the air temperature at thefloor level using 50% rule.

T5: Use equation 3.2

T6: This can be done for example usingdimensionless temperature methodthat was introduced in chapter 3.3.


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46

Air Quality

Select stratification height

Determine the convective flow ratesthrough the stratification height

Calculate the exhaust contaminantconcentration, c

Evaluate the concentration in theoccupied zone, c

Choose supply air flow rate,q = sum of convective flows

Temperature

Select thermalcomfort criteria

Calculate the heat surplus to beremoved by the ventilating air

Calculate the supply air temperature,

Calculate the supply air volume flowrate, q

Calculate the maximum temperatureincrease from supply to exhaust air,

-

Re-evaluate the air temperatureincrease at floor level,

S1

S2

S3

S4

S5

T1

T2

T3

T4

T5

Result

Check that the air flow rate is sufficient according to codes andstandards.

Choose the air volume flow q with regard totemperatures, air quality and regulations.

Re-calculate the vertical temperature distribution in the room,and estimate the pollutant stratification height.

Select diffusers and ensure that the adjacent zones are acceptable.

T6

R1

R2

R3

R4

s

e

oz

e s

s

s

f

s

Figure 5.6 Displacement Ventilation Design Procedure.


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58

8 CASE STUDIES

8.1 Restaurant

Figure 8.1 The restaurant that shall be ventilated.

The first case study is a restaurant withdense seating and both smoking and non-smoking areas. This is a typical casewhere air quality is a major issue.

8.1.1 Description

Table 8.1 Data for the restaurant.

Room dimensions

Floor area 132,0 mHeight 3,0 mRoom volume 396,0 m

Max. number of people in the restaurantSmokers 48 pers.

Non-smokers 48 pers.Employees 6 pers.Max occupancy 102 persons

Floor area per customer 1,38 m/customer The restaurant to be ventilated, shown inFigure 8.1, has a floor area of 132 m. Itwill accommodate a maximum of 96

customers, 48 in the smoking area and 48in the non-smoking area.

8.1.2 Design criteriaThe thermal comfort criteria are asfollows:

Table 8.2 Thermal comfort requirements.

Temperatures in the occupied spaceMax. temp. 26 C

Min. temp. 20 CTarget temperature 23 CMaximum verticaltemperature gradient 2,0 K/m

The air quality demands of the restaurantowner are:

it complies with the governmentalregulations (*)

the air quality is good the non-smokers are affected by

tobacco smoke as little as possible


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(*) The majority of countries have nationalrequirements for the ventilation air flowrate per customer in restaurants.

8.1.3 Ventilation strategy

In the smoking area, air quality is a major concern. Although displacementventilation is the chosen method, by itself it may not be adequate to ensure good air quality for the non-smokers. Theventilation should be designed so that littleor no air from the smoking area creeps intothe non-smoking area. This can beachieved by supplying as much air as

possible into the non-smoking area, andextracting air from the smoking area. Thesmoking zone and the building elementsshould be arranged so that smoke-contaminated air does not infiltrate thenon-smoking zone. This is illustrated inFigure 8.2.

8.1.4 Design for air qualityThe maximum number of people in theroom is 102. Using a ventilation rate of 20l/s per person, the ventilation flow required

for contaminant stratification above thesedentary peoples heads is calculated (seechapter 3):

q s = 102 20 l/s = 2 040 l/s (= 7 344 m/h)

8.1.5 Design for thermal comfortThe heat gain of the room is given in Table8.3 and shows that almost 90% of the heatsurplus comes from the customers.On a typical day, the restaurant is open for a couple of hours at lunchtime and a muchlonger period in the evening. Thenecessary airflow rate to remove the heatsurplus is determined by taking intoconsideration the heat accumulation in the

building fabric.

Wardrobe

Supply air

Extract air

Main air flowdirection

S m o k

i n g a r

e a

Non-smoking area

Kitchen

Doors for waiters

Column(unmovable)

(1)

(2)

(3)

Entrancedoors

(4)

(5)

Figure 8.2 Seating arrangement and ventilation strategy for the restaurant.


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Table 8.3 Heat gain without regard to heat accumulation in building elements.

People: 102 persons 85 W/person 8 670 W 65,7 W/mLighting: 12 lamps 100 W/lamp 1 200 W 9,1 W/m

Sum: 9 870 W 74,8 W/m

For this example it is assumed that the heataccumulation in the building fabric reducesthe need for air-cooling by 40%. Thus, thenet requirement for cooling by theventilation air becomes:

net = 0,6 9,87 kW 6 kW

This corresponds to a specific heat load of 45,5 W/m.

The room is 3 metres high. A verticaltemperature gradient of 2K/m correspondsto a temperature difference between floor and ceiling of 6C. By the 50% rulethere is a temperature difference of 12K

between extract air and supply air.However, this is more than most air diffusers can handle without causingdraughts along the floor, so a maximum

temperature difference between the extractand supply air is calculated as follows

= e s = 10K

A temperature-diagram for this case isshown in Figure 8.3.

This gives a ventilation rate of:

q s = 478 l/s ( = 1 720 m/h)

Comment: The maximum temperaturedifference of 10C between extract andsupply air is similar to that normally usedfor mixing ventilation. Thus, the air volume flow for removal of heat surpluswill be the same for both displacement andmixing ventilation.

0

0,5

1,0

1,5

2,0

2,5

Temperature, [C]

H e i g

h t a b o v e

f l o o r

[ m ]

50% 50%

18 19 20 21 22 23 24 25 26 2717161514

3,0

Room air temperature ~ 23C1,1

Extract air temperature = 26Ce

Air temperature atfloor level, = 21C f

Supply air temperature =16C s

Figure 8.3 Temperature diagram at maximum temperature difference between extract air and

supply air.


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8 CASE STUDIES - RESTAURANT

61

8.1.6 Resulting ventilation data

Table 8.4 Ventilation rates.

Ventilation rates total per person per floor area

q s q s /n q s /A f l/s l/s pers l/s m (m/h m)

With respect to air quality 2040 20,0 15,45 (55,6)With respect to thermal comfort 478 4,7 3,62 (13,0)

National regulations - - - -Choice 1020 7,73 (27,8)Air change rate: 9,3

Ventilation air flow rateWhen comparing the ventilation rate fromair quality considerations and from thermalcomfort considerations, the ideal flow rateof 20 l/s per person gives a very highventilation rate, while the ventilation ratefor thermal comfort is very low.A compromise of q s = 9,5 l/s per person ischosen, giving a temperature difference of 5K between supply and extract air. This

provides a reasonable air quality when therestaurant is full, and a reasonabletemperature difference between the supply

air and the extract air. However, most of the time, the restaurant is only half full, inwhich case, the air flow rate will be up to20 l/s per person. This will give excellentair quality.

Design air qualityThe CO 2-emission of a seated person isabout 20 l/h = 0,006 l/s (Recknagel et.al.2001). Assuming that the CO 2-concentration in the supply air is c s = 350

ppm (the outdoor concentration), we cancalculate the concentration in the extractair, ce. The air quality in the breathing zonewill be better than in the extract air.

Design temperaturesWith this choice of ventilation rate, thedesign data is shown in Table 8.6, and the

temperature diagram in Figure 8.4.

Table 8.5 Design air quality for the ventilation of the restaurant.

Min. vent. Max. ventAir volume flow per person: q s /n 10 l/s 20 l/sCO 2-concentration increase ce - c s 556 ppm 278 ppm

CO 2-concentration in extract air ce 906 ppm 556 ppm

Table 8.6 Design temperatures for the ventilation of the restaurant.

Air flow rate q s 1000 l/s = 3600 m/hTemp. difference extract - supply e - s 5,0 K Air temperature at floor level: f 22,1 CSupply air temperature: s 20,1 CExtract air temperature: e 25,1 CAverage vertical temp. gradient s 0,8 K/mTemperature 1,1 m above floor 1,1m 23 K Temp-diff. 1m - supply 1m- s 2,9 K


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0

0,5

1,0

1,5

2,0

2,5

50% 50%

18 19 20 21 22 23 24 25 26 2717161514

3,0

Temperature, [C]

H e i g h

t a b o v e

f l o o r

[ m ]

Room air temperature ~ 23C1,1

Extract air temperature = 24,5Ce

Air temperature atfloor level, = 22C f

Supply air temperature =19,5C s

Figure 8.4 Temperature diagram at the chosen ventilation rate and maximum heat gain.

8.1.7 ArrangementsLocation of air diffusersWhen locating the air diffuser, remember that many restaurants are refurbished andmodified several times during the lifetimeof a ventilation system. Thus, air terminaldevices and ducts should be located insuch a way that only minor changes arerequired when the room is refurbished.Moreover, the ventilation system shouldnot require that the furniture should bekept away from those areas where arestaurant operator would find it natural to

place furniture.

The air diffusers are located beside twocolumns (pos 1 and pos 2 inFigure 8.2) and in the passage outside thedoor between the kitchen and therestaurant (pos 3 inFigure 8.2). There will probably be noseating in these areas, so that the adjacentzone can be large, if necessary. Diffusershave not been placed along the walls,

because they would be too close to some of the seats.

Units 1 and 2Units (1) and (2) are located as shown inFigure 8.5. Two semi-circular wall unitswith the same diameter and width of thecolumns i.e. 0,7 m are installed in theselocations. The unit is shown in Figure 8.6.

The adjacent zone diagram for this unit isshown below. Looking at the seating plan,it can be seen that the nearest ankles are

about 1,5 metres from the diffusers. FromFigure 8.7 it can be seen that more than350 l/s can be supplied from each unit. Toallow for some margin of safety, choose

q s = 320 l/s from each of unit1 and unit 2

Some seats are closer than 1,5 metres fromdiffuser no.2. To protect these seats fromdraughts, partitions are placed betweenthem and diffuser 2. This is shown in

Figure 8.8.


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63

(1)

(2)

Figure 8.5 Air diffuser (1) and (2).

Figure 8.6 Semicircular wall unit for places 1 and 2.


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0 100 200 300 400 500 600 700 800 900

Supply air flow, q [l/s]

0

1

2

3

4

5

Supply unit 1 and 2

B = 0,7m, H = 1,8m W = 0,65m

A d j a c e n

t z o n e

l e n g

t h , l

[ m ]

n - = 3 K oz s

s

Figure 8.7 Adjacent zone length for units 1 and 2.

Figure 8.8 Partitions between the diffuser and the closest seats protecting the customers fromdraughts.


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Unit 3For unit 3, the available space is shown inFigure 8.9. The front surface area is 2,5metres wide and 1,25 metres high. The

distance from the diffuser to the ankles of the nearest customers is about 1,5 metres.In this space two plane units were installedas shown in Figure 8.10. The dimensionsof this unit are: height = 1,2 m, width = 1,1m and depth 0,3 m.

The length of the adjacent zone for anunder-temperature of 3K is shown inFigure 8.11 as a function of the supply

volume flow. For an adjacent zone of 1,5metres, each unit supplies approximately180 l/s. Two units are placed at thislocation, giving total supply airflow of 360 l/s.

2,5m

1,25 m

1,5 m

Figure 8.9 Air diffuser no. 3.

Figure 8.10 Diffuser chosen for place 3.


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66

0 100 200 300 400 5000

1

2

3

4

5

Supply unit 3

B = 1 m, H = 1,2 m W = 0,3 m

- = 3 K oz s

A d j a c e n

t z o n e

l e n g

t h , l

[ m ]

n

Supply air flow, q [l/s] Figure 8.11 Adjacent zone length for the units at place 3.

(1)

(2)(3)

(4)

(5)

q = 330 l/s

Threshold below ceiling

q = 330 l/s

q = 330 l/s

q = 330 l/s

q = 670 l/se,4

s,1

s,2

s,3

e,5

Figure 8.12 Location of extract units.

Warning: When putting two or more units beside each other, the airflow from thediffusers merge, and the adjacent zone

becomes larger than that from one singleunit. To allow for this effect, the supply

rate is reduced from 2 x 180 l/s = 360 l/s to2 x 160 l/s = 320 l/s from unit 3.

A total of 1000 l/s is extracted from theroom. More air is extracted from thesmoking zone than from the non-smokingzone, to prevent smoke drifting into thenon-smoking area. In addition, a thresholdis put below the ceiling between the two

zones. Its location is shown in Figure 8.12.Also, see Figure 8.13 below.


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9 REFERENCES

Alamdari, F (1998). DisplacementVentilation and Cooled Ceilings.Proceedings of Roomvent '98, Stockholm.

Bach, H et al 1993. Gezielte belftung der Arbeitsbereiche in produktionshallen zumAbbau der Schadstoffbelstung.Forschungsbericht HLK-1-92 ISSN 0943-013 X.

Batchelor, G.K. 1954 Heat convection and

buoyancy effects in fluids. Quart. J. Roy.Met. Soc. 80, pp. 339-358.

Brohus, H. and Nielsen, P. V. (1996 A),Personal Exposure in DisplacementVentilated Rooms. In: Indoor Air:International Journal of Indoor Air Qualityand Climate, Vol. 6, No. 3. - pp 157-167.

Brohus H (1998), Influence of the cooledceiling on indoor air quality in adisplacement ventilated space examined bymeans of computational fluid dynamics.Proceedings of Roomvent '98, Stockholm.

Brohus, H. and Nielsen, P. V. (1996 B),CFD Models of Persons Evaluated by Full-Scale Wind Channel Experiments. In:ROOMVENT96. Proceedings of the 5thInternational Conference on Air Distribution in Rooms, Yokohama, Japan,

July 17-19, 1996 / Murakami, S. (ed.). -Vol. 2, pp. 137-144.

Brohus, H and Ryberg, H. (1999)TheKappa Model a Simple Model for theApproximate Determination of VerticalTemperature Distribution in Rooms.Department of Building Technology andStructural Engineering, AalborgUniversity.

Brouns, C., Waters, B.: A guide tocontaminant removal effectiveness.Technical note AIVC 28:2, AIVC, 1991.

Butler, D. (2002), Air conditioning usingdisplacement ventilation to maximise freecooling. BRE, UK. To be published.

CEN Report, 1997: CR 12792 Ventilationfor buildings - Terminology, symbols andunits. (revised and completed version

expected to be upgraded into EuropeanStandard by year 2002)

CEN Report, 1998: CR 1752 Designcriteria for indoor environment.

Elterman, V.M. 1980. Ventilation of Chemical Plants. Moscow: KHIMIA.

Etheridge, D., Sandberg, M. 1996.Building Ventilation - Theory and

Measurement, Wiley.Fitzner, K. 1991. Quellftung in Theorieund Praxis, Japanisches Journal fr Heizungs- und Klimatechnik, Heft 1, 1994;und CCI, Heft 5, 1991

Fitzner, K. 1996. Displacement ventilationand cooled ceilings, results of laboratorytests and practical installations Proc.Indoor Air 1996, Nagoya.

Fitzner, K. 2001. Private communications.

Goodfellow, Howard and Esko Thti(2001). Industrial Ventilation DesignGuidebook. Academic Press, 2001, 1520

pages.

Halton OY (2000). DisplacementVentilation Design Guide, 01.2000


34/35

REFERENCES

97

Hagstrm K, Sandberg E, Koskela H,Hautalampi T (2000). Classification for theroom air conditioning strategies. Buildingand Environment, to be published in 2000.

ISO 7730, 1994. Moderate thermalenvironments - Determination of the PMVand PPD indices and specification of theconditions for thermal comfort.

Jacobsen, T.V. and P.V.Nielsen (1992).Velocity and temperature distribution inflow from an air inlet device withdisplacement ventilation. Roomvent92,

p.21, ISBN 87-982652-6-2

Jaluria, Y. 1980. Natural convection, heatand mass transfer. Pergamon Press.

Jin, Y.1993. Particle transport in turbulent buoyant plumes rising in a stably stratifiedenvironment. Ph.D. Thesis, Dept. of Building Services Engineering, KTH,Stockholm.

Kofoed, P. 1991. Thermal plumes inventilated rooms. Ph.D. Thesis, Aalborg.

Krhne, H. 1995. Experimentelle undtheoretische Untersuchungen zur Quelluftstrmung, Dissertation TUBerlin.

Krhne, H., Fitzner,K. 1995. Luftqualittin der Atemzone von Personen mitQuelluftstrmung. Luft- und Kltetechnik 12/1995

Livtchak A, Nall D, Displacement

Ventilation (2001). Application for HotandHumid Climate, Submitted for Proceedingsof Clima 2000/Napoli 2001.

Mierzwinski, S . 1981. Air motion andtemperature distribution above a human

body in result of natural convection. A4-serien no. 45, Inst. fr Uppv.- o Vent.teknik, KTH, Stockholm.

Morton, B.R., Taylor, G., Turner, J.S.1956. Turbulent gravitational convectionfrom maintained and instantaneoussources. Proc. Royal Soc., Vol. 234A, p.1.

Mundt, E. 1990; Convection flows abovecommon heat sources in rooms withdisplacement ventilation. Proc.ROOMVENT '90, Oslo.

Mundt, E. 1992. Convection Flows inRooms with Temperature Gradients -Theory and Measurements.ROOMVENT'92. Proceedings of the ThirdInternational Conference on Air Distribution in Rooms. Vol. 3. Aalborg.

Mundt, E. 1994 ContaminationDistribution in Displacement Ventilation -Influence of Disturbances. Building andEnvironment, Vol 29, No. 3

Mundt, E. 1996. The Performance of Displacement Ventilation Systems-Experimental and Theoretical studies,Ph.D. Thesis, Bulletin no 38, Building

Services Engineering, KTH, Stockholm.Mundt, E. 2001. Non-buoyant pollutantsources and particles in displacementventilation. Building and Environment,Vol 36, No. 7.

Nickel, J., 1990. Air Distribution inDisplacement Ventilation (in Danish).VVS, Teknisk Forlag A/S, Copenhagen,March.

Nielsen, P. V., 2000, Velocity Distributionin a Room Ventilated by DisplacementVentilation and Wall-Mounted Air Terminal Devices. Energy and Building.Vol. 31/3, pp. 179-187.

Nielsen, P.V., 1996 TemperatureDistribution in a Displacement VentilatedRoom. In ROOMVENT 96: Proceedingsof the 5th International Conference on Air

Distribution in rooms, Yokohama, Japan,July 17-19, 1996 / Ed. by Murakami, S.


35/35


Vol. 3, pp. 323-330. ISBN: 4-924557-01-3.

Nielsen, P.V., 1994(A), Velocity Distri- bution in a Room with Displacement

Ventilation and Low-Level Diffusers.International Report, IEA Annex 20,Aalborg University.

Nielsen, P. V., 1994 (B), Stratified Flow ina Room with Displacement Ventilation andWall-Mounted Air Terminal Devices.ASHRAE transaction, Vol. 100, Part 1.

Nielsen, P.V., 1993 (A). Air Distributionin Rooms - Room Air Movement and

Ventilation Effectiveness. InternationalSymposium on Room Convection andVentilation Effectiveness, ISRACVE,ASHRAE. Tokyo.

Nielsen, P.V. ,1993 (B). DisplacementVentilation - Theory and Design,Aalborg university. ISSN 0902-8002U9306.

Nielsen, P.V. , 1992, Velocity Distributionin the Flow from a Wall-Mounted Diffuser in Rooms with Displacement Ventilation.Presented at ROOMVENT'92. Third Int.Conf. on Air Distribution in Rooms,Aalborg, Denmark.

Nielsen, P.V. , L. Hoff and L. G. Pedersen(1988), Displacement Ventilation byDifferent Types of Diffusers. 9th AIVCConference on Effective Ventilation,

Ghent, Belgium. Nordtest method, Air Terminal Devices:Aerodynamic Testing and Rating at LowVelocity, NT VVS 083:A, ISSN 0283-7226, Finland, 2002.

Popiolec, Z. 1981. Problems of testing andmathematical modeling of plumes abovehuman body and other extensive heatsources. A4-serien no. 54, Inst. fr Uppv.-

o Vent. teknik, KTH, Stockholm.

Recknagel, Sprenger and Schramek (Taschenbuch fr Heizung und Klima-technik 01/02 ISBN 3-486-26450-8.

Schmidt, W. 1941. Turbulente Ausbreitung

eines Stromes erhitzer Luft ZAMM. Bd. 21#5.

Skistad, H. 1994. DisplacementVentilation. Research Studies Press, JohnWiley & Sons, Ltd., West Sussex. UK.

Skret, E. 2000. Ventilasjonsteknisk hand- bok. Hndbok 48, Norges Bygg-forskningsinstitutt. ISBN 82-536-0714-8.

Stymne, H., Sandberg, M., Mattsson, M.:Dispersion pattern of contaminants in adisplacement ventilated room- implicationsfor demand control. 12th AIVCconference, Ottawa, Canada, 1991.

Sutcliffe, H.: A guide to air changeefficiency. Technical note AIVC 28,AIVC, 1990.

Tan H, Murata T, Aoki K, Kurabuchi T

(1998). Cooled ceilings / displacementventilation hybrid air conditioning system -Design Criteria. Proceedings of Roomvent'98, Stockholm.

Tapola (1983). Dimensioning and controlof Displacement ventilation in IndustrialPremises: Ministry of Trade and industry.Energy department, serie D:173, Helsinki1983.

Turner, J.S. (1973), Buoyancy effects influids. Cambridge University Press.ISBN 0 521 08623 X.

Zeldovitch, Y.B. 1937. FundamentalPrinciples for Free Convective Plumes.Journal of the Experimental and TechnicalPhysics. Vol.7(12). Moscow.

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