AIVC 11896 4259 (RP-949)

Performance Evaluation and Design Guidelines for Displacement Ventilation

Xiaoxlong Yuan, Ph.D. Qlngyan Chen, Ph.D. Member ASHRAE

Leon R. Glicksman, Ph.D. Member ASHRAE

ABST RACT

This paper evaluates the peiformance of traditional displacement ventilation systems for small offices, large offices with partitions, classrooms, and industrial workshops under U.S. thermal and flow boundary conditions, such as a high cooling load. With proper design, displacement ventila­tion can maintain a thermally comfortable environment that has a low air velocity, a small temperature difference between the head and foot level, and a low percentage of dissatisfied people.

Compared with conventional mixing ventilation, displacement ventilation may provide better indoor air quality in the occupied zone when the contaminant sources are asso­ciated with the heat sources. The mean age of air is younger, and the ventilation effectiveness is higher.

Based on results from Scandinavian countries and the authors' investigation of U.S. buildings, this paper presents guidelines for designing displacement ventilation in the United States.

INTRODUCTION

Displacement ventilation has been used quite commonly in Scandinavia during the past 20 years for ventilation of industrial facilities to provide good indoor air quality and save energy. In 1989 in Nordic countries, it was estimated that displacement ventilation accounted for a 50% market share of industrial applications and 25% of office applications (Svens­son 1989). However, very few applications of displacement ventilation can be found in the United States. U.S. designers are not very familiar with the performance of displacement ventilation and do not have suitable design guidelines.

Although a lot of research on displacement ventilation has been conducted in Scandinavian countries, as reviewed by Yuan et al. (1998), direct application of Scandinavian results for U.S. design may not be feasible because U.S. buildings have different layouts and higher cooling loads than Scandi­navian buildings.

This paper evaluates the performance of displacement ventilation systems for small offices, large offices with parti­tions, classrooms, and industrial workshops under U.S. ther­mal and flow boundary conditions such as a high cooling load. Based on the results from Scandinavian countries and the authors' investigation of U.S. buildings, this paper presents guidelines for designing displacement ventilation in the United States.

PERFORMANCE EVALUATION OF DISPLACEMENT VENTILATION

At the design stage, a designer needs to predict the perfor­mance of displacement ventilation. The performance is often evaluated by the thermal comfort level, indoor air quality, energy consumption of the HVAC system, and the first cost and maintenance costs of the system. The air temperature distribution, the percentage of dissatisfied people due to draft (PD), and the predicted percentage of dissatisfied for thermal comfort (PPD) are widely used as criteria to evaluate thermal comfort. Contaminant concentration distributions and the mean age of air are often good indicators for indoor air quality. Energy consumption is related to air temperature distribution and ventilation rate. All these performance parameters are determined by the thermal and flow boundary conditions, such as the size and geometry of the space, heat sources, and contaminant sources.

Xiaoxiong Yuan was a post-doctoral associate and is currently a senior engineer at Applied Materials, Inc., Santa Clara, Calif. Qingyan (Yan) Chen is an associate professor and Leon R. Glicksman is a professor in the Department of Architecture, Massachusetts Institute of Technol­ogy, Cambridge.

THIS PREPRINT IS FOR DISCUSSION PURPOSES ONLY, FOR INCLUSION IN ASHRAE TRANSACTIONS 1999, V. 105, Pt. 1. Not to be reprinted in whole or in part without written permiseion of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., 1791 Tullie Circle, NE, Atlanta, GA 30329. Opinions, findings, conclualons, or recommendatione expreued In this paper are those of the author(s) and do not necessarily reflect the views of ASHRAE. Written question• and comments regarding this paper should be received at ASHRAE no later than February 13, 1919.

Evaluation Criteria This secti1;m will evaluate the performan�e of displace­

ment ventilation for individual offict<s, large offices with parti­tions, cla��rooms, and industrial workshops. Jhe investigation uses the CFD program validated in a companion pap�r (Yuan et al. l 999a). This program is used to calculate the flow.and thermal distribution for �.large number of different boll!ldary conditions. l:fowever, nor,mal CFD program�. po not calsulate PD, PPD, and the mean age of air. We have imple­men,ted the following models in the CFD program to calculate these parameters. , ,

Fanger et al. (19&9) developeq a model to calculate PD as

PD� o:o21(�3�2-1)(u-9.S)0·62(3.14 + 0.0019u Tu) [%l (IP') (la) PD= (34-T)(u - 0.05)0·62(3.14 + 0.37u Tu) [%l

(for PD> 100%, use PD= 100%)

(Sl)(lb)

where T u

= air temperature (0C), . '

= air velocity (fpm or mis) (for u < 10 fym or 0�05 ml s, use u = 10 fpm or 0.05 mis), "

Tu = turbulent intensity.

We use ·' (i)

where k = the turbulent �etjc energy.

: ' �1 . ' �

The PJ?E> can btf.calculated-via (ISO 1990): � . ' ·"

PPQ = 100-95 exp(-0, 3353PMV4-0.2 ·.9.PMV,2) (%q• 3) f'' I )' <' !

The pre<:licted mean vote, PMV, in the equation is· d-et��-. . ,... i rruned by 1 •

PMV = 3.155 [0.303 exp(--0.114M) + 0.02Sl L (IP) (4a) PMV = [0.303 exp(,,-0,036M) ,+ 0.02Sl L (SI) (4b)

where L=M-W

- { 1.196 x 10"9fc1[(Tc1+ 460)4 -(T, + 460)41 + fcA<Tc1- n + 0.97 [5.73 - 0.022 (M -W)- 6.9 Pal + 0.42 (M- W-lS.43) + O.Q173 M (5.S7 - 0.69 Pa) -0.00077 M (93.2 -1) h (IP) (Sa)

L=M-,_W ') :"' - { 3.96 X' l 0-8fcif<Tc1 + 273)4 - (Tr,� 2.J3§1J+ fc�iTc1 -.1)

+ 3:�� x 10·3 [5733-6.2 •. �M"ll �:_Pal

with M w fc1

2

+ 0.42 (M-' W,.-5.8.15) + 1.7 x 10:5M°(58G1T P0) + d'.O,_ l4M (34 -1) I ,.

' . (SI) (Sb)

. ' '11 ''i\h�·

= metabolism (Btu/h, W/m2): ., , '··

= external work (Btu/h, W /m2), = cloth factor,

= loq<)l air temperature (°F, 0C), = cloth temperature (°F, 0C), : . ; = mean radiant temperature (°F, °C), = convective heat transfer coefficient between the

cloth and air (Btulh·fi2·°F, W/m2·K), = partial water vapor pr�sureJtP· water, Pa).

The fc1, Tc1, and he are determined py the following (fqua-tions:

fet = 1.05 + 0.645/c/

fc1 = 1.00 + 1.290/c/

for I ct <? 0.07S for Ic1 < 0.07S

'-•

Tc1=96.3- 0.156(M - W)-lcd 1.196 x 16c9Jc1[(Tc1+460)4 -

(6) (7)

(T, + 460)41 + fc1hc(Tc1 -1) (IP) (Sa)

Tc1 = 35.7 - 0.02S(M - W)-lc1{3.96 x 10-8.fc1[(Tc1 + 273)4 -(T, + 273)41 + fc1hc(Tc1 -1) (SI) (Sb)

The convective heat transfer coefficient, he> is determined from:

h = 0.361(T -1)0·25 C l . c/ ! L. �! for 0.361(Tc1.-c: T) 0}5 � p.151u0·5 ·,., (IP) (9a)

\; ..t•t:: ' h = 2.38(i' ....:. n°·25 e . ' .. cl

for 2.38(Tc1 - T)0·25 � 12.1 u0·5 (SI) (9b)

h = 0.151,.�0.5 C 'Sf ' {o 0.36l(t -�: ·�.�� o:. 1 1P·5 (IP) (JOa) . . .., el,.. I . .. Ai . ... •

•• ' .,. .. 1··'·1�•"' 'f 1�-' li·\1. ·"'d' . � )) 'I •

' '· ··1:'.I f : b ·=.1�. · u ·5 ,1 �... ' . ' C• I ;.)7). l�.\ • � I I "' -

· · fot 2.38(T..1-Tj 25 < 12.1 uO. . / (-.•,., · (SI) (IOb) ,.r. � . c.; , f � • , It J.' i;- ... h � . w ere

• ,., '1 • ., let =- �clothi�g insulation (°F'.ft2.h/Btu, °C·m?-!w), u = air velocity (fpm, mis).

The mean age o� arr, 't, is defined as tlle averaged time for all air molecules to ttavcl 'from ·we supply diffuser to that point. It can be derived from the measured transient history of the tracer-gas concentration. Li and Jiang (1996) show that the mean age of air is governed by a transport equation:

a 'a· . o ( ai:) a,(P'ti�+ ax (pu/t:), = ax r-tox. � p ..•. • . j '· • J •• J.

. . . with the following as•the b�unctary 'co'.nditiotis: I • .{

(11)

't = Q. '!�,.\he supply di�s,rr, t ; ' l �:j Q"t I.' I ' ' \, :, � ("' fl.°':t I·� ' •

:;- = d'lit!tll� eltl1. usf and W, · p.s. • 1 • ci,�. ax1 ,. , . 1 • , �·h !.;� .,< The ventilation effecqv.cneS's i ·d�f¢ep 'JS

... • ·.�

y.' ct-c J<

T\ = --c-cs

where '. 1·� \ ., {� o • 0\ r-)l'l , , '\'•\I" ' ... Tl

. . = ventilation effectiveness,

(12)

I'

4259 (RP-949)

= contaminant concentration at the exhaust air (ppm),

cs = contaminant concentration at the supply air (ppm),

c = contaminant concentration in the ro9m air (ppm). • ' ' 1.1-.

In addition a CFD program will calculate the airflow pattern and ilie distributions of air velocity, temperature, and contru:ninant concentration. We will be able to evaluate the ther­mal comfort and indoor air quality provided by displacement ventilation in terms of

•

•

airflow pattern, temperatu�e di�tribution, PD,

• PPD,

, .

I • contaminant concentration distributions, • mean age of air, and • ventilation effectiveness.

The energy consumption ,of the HVAC system and the first costs and maintenance costS of the system for displace­ment ventilation are al!io important parameter for evaluating the perfonnance of the ystem. We are conducting research on energy consumption and costs and will report the results in another·1Japer.

.. .

•• ) I• •t·, . ; ')

' /..•

(c)

Case Studied The present investigation studied displacement ventilation

systems without chilled ceiling panels for four types of paces: small offices, large offices with partitions, cla5srooms, and industrial workshops. Figure 1 illustrates a typical space config­uration for each type. In order to obtain sufficient information to evaluate the pe1fonna.hce of displacement ventilation; we have used the CFD program to study 56 cases. These inclutl�'18 cases of small offices, 12 cases of large offices with partitions; 14 cases of classrooms, and 12 cases of industrial workshop'ii.: The thermal and flow conditions for the cases are summarized in a companion paper (Yuan et al. 1999b). These thermal and flow poundary conditions cover a wide range ofU,S. buildings:, • 8 ft$ room height$ 18 ft (2.43 m $room height$ 5.5 m) • 2 ACH $ ventilation rate $ 15 ACH • 6.6 Btu/(h·ft2) $Qt IA$ 38 Btu/(b·ft2) (21 Wlm2 $ Q1 IA

$120 Wlm2) • 0.08 � Q0/Q, S 0.68 • Jl ,$ QlQt $ 0.43 )·1 . • b � QeJQ1 $ 0.92 where

'I

Q1 = total cooling load in the room; A = floor surface area; Q0e = heat generated by occupant, desk lamps, and

equipment;

(b)

. · co ·• ; •

Figure 1 Typical rooms studied: (a) a small office, (b) a 'large office with partitions, (c) a classroom, (d) a workshop. i. ·.. ·:1i

4259 (RP-949) 3

= heat genei;ated by lighting;

= heat from exterior walls and; windows and transmitted solar radiation.

Performance Evaluation of " r;i1splacement Ventilation For the 56 cases, we have calculated the airflow patterns

and the distributions of air temperature, PD, PPD, C02 concentration, the mean age of air, and ventilation effective­ness. This sc;c:tio:p,,uses the results for a classroom shown in Figµre le a,s; an example. Tile. classroom has a .teacher, 24 pµpils, an exterior window/wal_� •. . 9verheacL ligqting, and dif'fusers in all four comers. More, d�tailed information is

. . .

1:" i'

''·'

i � _,: .i J' . 11 •• •

I 1•.,.1

'' -�· t,.

I• " .• t ... i."

'i J

j-'....:::

provided it) ca e CR01 in Tab!<; 1 in Yuan et aj. (1999b). The results for other cases can lead to similar conclusions.

Airflow Pattern. The classroom uses one diffuser in each of the four comers. Figure 2 show the airflow p�tterns in three different sections of the room: a horizontal section at ankle level, section A-A, and section B-B (see Figure 2a for the sl!ction Ideation). The relatively cold air from the diffusf!ts falls toward the floor. The falling tl0ws behave like jets, and the jets can meet iri the center of tlte room. The wall temper­ature determines 'the flow near the w�l's:iTli'e heated objects, such as the occupants, generate strong

' plumes that bring the

c;on�minants from the lower zone to:the;1upper zone. If the walls are assumed to be isothermal and �aye the same temper-

IB

·:·:fi. ·L . ·�. . '

,{

•),.l .,. ..... . ,..,

, .....

::1.

J. •'

--- -0.'20�/s (40··fpm) · • __ :.__ �:·_:_

(c) --- 0.29�/s (40 fpm) Figure 2 The aiiflow pa{te!n in t�� classraorri_: r(q) 1;1e,f}r,,;a71�lel<tYe.l, �))t,pe,qtion A,�(\.,,( c� in se9ti<?[!i�B.

4 4259 (RP-949)

ature as the average room air temperature, the wall flow in the lower part is upward and in the upper part of the wall the flow is downward. This occurs because room air temperature in the lower part is lower than th�t of the walls and in the upper part, higher than th<it of the walls. However, in our experiment, we found that the temperature of the walls is higher than the room �: temper_ature. Therefore, the �?WS near tile wall are atl upward, a� shown in Figure 2. The air velocity in the room with displacement ventilation is generally small (less than 40 fpm or 0.2 m/s) except in the thermal P,lum_es and the flow near the floor anq ,walls.

- --

In the winter, the downward flow near the- exterior window and wall might bring pollutants from the upper zone to the lower zone. Preventing such a downward flow requires a heating device placed near the exterior window or wall.at the floor level. The heating capacity should be slightly larger than the heating load from the exterior window or wall.

Temperature Distribution. Figure 3 shows that the air _

temperature is nearly uniform in the horizontal direction except in the region close to an occupant. The supply_ air is heated first by the floor and mixed with room air by induction. As a result, the air temperature near the floor, T1, is higher than that of supply air. To avoid draft, the supply air temperature cannot be too low. The supply air temperature depends on the room geometry, cooling load, and heat source type. Generally, :.

lower. In general, radiation from the ceiling may not be felt unless the air temperature near the ceiling is very:high.

Percentage Dissatisfied Due to Draft (PD) and Predicted Percentage Dissatisfied (PPD). The PD and PPD are generally less than 15% in the occupied zone, if the design uses the guidelines to be shown in' the next section. Figure5'4 and 5 show that only in the region very close to the diffuser {l .6 ft or 0.5 m) the PD- and PPD are higher. Therefore displacement ventilation presents a sa·tisfactory comfort level.

Contaminant Concentration Distributions. The present study uses C02 as an indicator for contaminants. With displacement ventilation, the'C02 concentration in the lower zone is lower than that in thC- upper zone, as shown in Figure 6. The C02 level at the inlet is 400 ppm. The indoor C02 source� are from the oq:upants, and the occupants are the heat sources as well. The h(!at generates thermal plumes that can bring the C02 to the upper zone. In most buildings, many contaminant sources �e associated with the heat sources, such as. :printers, computers, and other heated equipment. Displace­ment ventilation.provides better indoor air quality than mixing ventilation for these cases.

The convective flow around a human body brings the air at a lower zone to the breathing zone. Therefore, the occupant actually breathes air with }ower contaminant concentrations

- . than tho_se at the nose level in the middle of the room. the supply air temperature is in the range of 65°F�68°F (l8°C- However, dispiacement ventilation may not provide 20°C). There is a vertical gradient. of the air. temperature, bett�i;:iridoor air cjuliJ.ity than mixing ventilation if the contam-which depends on the distribution of the heat smirces.-The inant sources are not' associated with heat sources, such as gradient in the lower part will be larger than that in_ the upper c'volatile;. organic .compounds from building materials. The part if most of the heat sources are in the fower part of the - � �onc�ntr'ation is i�deed vei-y sensitive to the locations of the room. Although Figure 3 shows a high air temperature near the contaminant and heat sources, wall thermal conditions, and ceiling, the ceiling surface temperature is several degrees disturbances, such as moving objects in the room.

- : .

M

�S" ;'·fVIS.2 F

MS 75.2 F (24 C) _ ': ' ·

� . · _ ··

C) _ '_13 .4°.F . ,·_ . . � )�·6 � - c=:Y-.

. / 7�; �-{U''' -:;3 �� - _ (a) :1r :

so·�§ F c21·-c

. • r. " . .:·! \, (b)

�F 0

Figure 3 The air temperatu1'e distribution 'in the cla'ssrdo"tt1:1(ti)lin•secti-on A�A; (b) in s'ection"BiB:"

4259 (RP-949) 5

0

-- .. �,._ .. _.. . ..

0

- -·--- - ·-·---

)"

(c) �\�\� :; t, �! 1/< l�\ ; \\ �··� ''� v �J :1: ,I .1:1 �\,, , ,•, \ .\ � ;\:.:� '�

Figure 4 The PD distribution in the classroom: (a) near the ankle level, (b) in section A-A, (c) in section B-B.

4a59. (Rl?-949)·

10

oD

1'· t ·; ,\

1' '

I

DB DB DB

DD D·D DD DD D - ·D D D DD DD DD 0� D .,D

•,

(a)

·, I" .

I , . . :

. (b) 1

- ... o- -

- - --

•J

- . .. - ·

10

Figure 5 The PPD distribution in the classroom: (a) near the ankle level, (b) in section A-A, (c) in section B-B. ' )�'" -� . \\� i.#\ �·-�) ,·-.:,• :l� ':t•.·1 .�_11·:\ ':, ,, � �-._-:\ (;.:11' : .:• -�· ' "\.•'"

4259 (RP-949) 71

(a)

Figure 6 The·C03 &istribution in th�·�lassroortz: (a)1ln sectionA�A. (b) irl s�ctio;f B-'B: •Ji :i:' ' .. �1" �.i. < • I h'J·'�:

. , ,.

.>

l � ..

' 1 ... � ....

'.l .,(b)

. _·( , ,{) ' l' ·u. I t 'h

· T� r

}\

'.

..

. ll . ___ .,

/" . [l ..

ff!: I

�:� _ 3rdv!:.H�, .tt 'if �)·:.:Jf': f Ji · . �·;1 J ":'-"''

- J').

--"''-C-"'"-'''-"1J5Qtf � 1 if 1 . ... ·; ,, , .:_".'"• I ;� ·u� � p1�1 i g�1.�

•r;;. .. ·1 . . � .: .

... , {; 'r � ;1 f1t ... , a •,Ii) H.ic-dz.� _ i ·.:.dt�·l.11:�l -� ; -

. .. . · \ • • � .,/ I .·t:

�!� _,_. •t ·L ·�!.' :,..�>�f 1trl ":Vlf"i?. :;-,1.!Ji.H!":J�_ ... ,, J'JLJ _ y..r '' \ i\.f: FigurB, 1.. 'flJt, d�tribution of the meal\1�ge pf qir (in seconds) in th� olas::.1mom:!�)1 i'n·>sl!Gtioh�'�A. W>·in ·section B-B.!

1-....

I' ,, ; I 4 o ; �J .� ·'. JJ? ', ''.:'1 • J (·) r. ll

.,,,_., .. ,: ....

; I -, .. .,

...

1: , •

"

8 4259 (RP-949r-

·ta)

(b) Figure 8 The ventilation effectiveness distribution in the clafisroom: (a) in section A-A, (b) in section B-B.

Mean Age of Air and Venti.Iatioq,Effectiveness. Figure 7 illustrates the mean age of air in the classroom. Clearly, th� mean age of air in the lower part of the room is much younger than that in the upper part of the room. The mean age of air at the breathing level in the classroom is about 600 s, when the same ventilation rate (4 ACH) is used. The mean age of air·in ·

the classroom with complete mixing is 900 s. The distribution _ _ . of ventilation effectiven!'!�s shown in-Figure 8 iilso indicates that the classroom with a displacement ventliation system has higher ventilation effectiveness. With a· perfect mixing venti­lation system, the ventilation effec.tiveness is 1. Perfect or complete mixing is impossible in practice. The corresponding - , mean age of air will b� older in the occupied zone than with displacement ventilation, and the ventilation 'effectiveness ____ _

will be lower than 1. Hence, we can conclude that displace­ment ventilation does provide milch better indoor air quality than mixing ventilation. - · ·-·- · - - --- - ·

DISCUSSION

The air tempera�e difference between the.he�d and foot levels is 3°F ( l .6°C). In this case, the ventilation rate is 15 ACH. The high ventilation rate would require a large fan and duct. It may not be economically feasible compared to mixing ventilation. In other words, energy efficiency and cost might limit the maximum cooling load. In ·addition, the .available area of the walls' fo-r' installing supply diffusers will limit the maximum c�oliiig_load. This deserves fur:tfier study.

DESIGN GUIDELINES · The above stud� .iihows that displacement ventilation is

v,erf promising and can be applied in the U.S. The authors haye. reviewed the design guidelines available to design displacement ventilation and concluded that the current guide­lines are not complete (Yuan et al. 1998). With measured data

_from the literature (Yuan et al. 1998) and from our environ­mental .chamber (Yuan et al .• 1999a) and the computed results ofthe.56 displacement ventilation cases, we have developed the fo1lowing ten guidelines for displacement ventilation.

We are particularly interested in studying cases with high cooling load. Our survey in several conventional buildings in _

the greater Boston area shows. that the cooling load -can be as· high as 40 Btu/(h·ft2) or 120 W/m2. Some investigators suggested that the maximum cooling loa'd the dis.place� ment ventilation system can handle is about 13 Btu/(h·ft2)

Step (1): Judge t�e Applicabillty of ·' Displacement Ventilation

( 40 W /m2). However, the present study shows that maximum cooling load depends S\fOngly,on,Uie,i;oom;h,eight, ventilation rate, and heat source type and location. A displacement venti­lation system may maintain a comfortable environment with a cooling load up to 40 Btu/(h-ft2) (120 W/m2).

We have studied a case with such a high cooling load. The PD and PPD in the occupied zone can still be less than 15%.

4259 (RP-949)

Displacement ventilation is suitable when the contam­inant sources are associated with heat sources and the ceiling

· height is no less than 8 ft. There is also a limitation on the cooling load that can be handled by displacement ventila-

• tion. The current·Stutly shows that•the m'axim�m can be as high as 38 Btu/(h·:ft2) (120 W/m2) if there is sufficient space for installing large diffusers. When the cooling load is high, the energy consumption with displacement ventilation will increase significantly. A more detailed energy analysis will be reported soon.

9

Step (2): Calculate Summer Design Cooling Load Use a cooling load program or the ASHRAE manual

method to calculate the design cooling load of the space in summer. If possible, assume a 1°F/ft (2°C/m) vertical temper­ature gradient in the space in the computer calculation because the air temperature in a room with displacement ventilation is not uniform.

•

•

Itemize the cooling load into

.the occupants, desk lamps, and equipment, Q0e (B tu/h, W); the overhead lighting, Q1 (Btu/h, W); the heat conduction thrbugh the room envelope and transmitted solar radiation, Qex (Btulh, W).

Step (3): Determine the Required Flow Rate of the Supply Air for Summer Cooling

Displacement ventilation creates a thermal stratification. To maintain a comf9rt level, the design air temperature differ­ence betw.�ei:l tl1'� head and fo'ot level of a·sedentari occupant, AThf' should be less than 3.5°F (2 K). The required v·entilation rate, n, can be determined according to (Yuan et al. 1999b):

where •'

, • 'a'';. 11, ... , = ventilatit;mrate.(ACH)e'I .,,, ·1 . AThf = 5°F (2 K), ". · Y p = air density (lb/ft3, kWm3), • : • �·

CP = specific heat of air (Btu/lb·°F,'Jficg·K): · . I .. ft. .

H = space height. (ft, 'm); '· · · ' ·

A = floor area (fi2, m2), and 'I

(13)

.1

I I�'

aoe• a1, and aex = 0:295, 0.132, and O�IB5, :reBpeC.tively. Tue. �oefficie1nw s��� for tt?e fractions·.Pf�$e co0Upg.l9ads enter­mg the space between tl;lf head an� f�t of a sedeqiflI'Y qcc_:��

t . pan . . . · ... ,, .. 1"'1 . ,, .

•; ANSllASHf?4$ Standard,5S-f.�9�.,Therrf/ql Enyiro'}'!}tgi-, tQj. Conditions for lf.uman Qccupa11cy (AS� 1992) ..

I \.,0,1 '" \ - • � �/' � requires the temperature differen«fe .bet��n the . .1;!1�d .;,ndJoot level of a standing person not to exceed.�0.f' (3 �) Siµc� .�he ,, t> 1 .. . , •. �

vertical ��mperature rad�ept in.;�e �p,9e b�t;r"e9n ?-!6 ft (�.l m) .,��d 5.5 ft ( .7 m) is ge. ner�H>; : �mal��r,;':9-·an th�� b�t�.e�n. P,.3 ft (0.1 m) and 3.6 ft (1. 1 ni),, jh� ,�e&1gri with, $b�1 abo�e'i"o nnula can also provid0e a conifoi1able' condition fOr a . ., . • Jff:.,, ff standing Rt?rson. .

The f'(ow rate re9hlred foi: summer cooling, 1V11, is ·.��n: •'" • I ' • ' ti ''

;: t J

vh = nAH/60 (cfin), ': .. L �'.i I : I

Vh = n AH/3(/JQ (Jn.%).", '.. :o:r:

. • i(JP) (l4a): • , -J : � • ) i '. � : •

.': .,. (SL}(J�)

· ... . :f: ' ' � '-��lJf;j,. :::;_ '. '.iJ :..... ··,�j St�m.(4):_,find tht;t �pgulred Floy.t .R.a�lil 9,,f're�ti,�Jrr., for Acceptable Indoor Air QulJ:llfy,.1 . .. :�' ;, ':.mil:

. i.,1Use ANSJJ/tSHRAE 'Sta�dard�62-J.989, Vtnrllation for Acceptable--.lndoo'fY_ Air Quality (�SHRAE J:989): to firid•the-1

10.

required flow rate for acceptable indoor !ilfr quality, V,.. This flow rate is based on mixing ventUatiori with a ventilation effectiveness of 1. As mentioned earlier, displacement venti­lation has higher ventilation effectiveness. For the same indoor air quality, displac,ement ventilation requires less fresh air than q1.i�ing ventilation. The required flow rate of fresh air ifor displacement ventilation, Vp can be calculated by

•,• . (15)

where 11 = the ventilati?n effectiveness at the breathing level.

The 11 can be determined by (Yuan et al. 1999b)

Step (5): Determine the Supply Airflow Rate ,.Ghoose the greater of the required flow rate for summer

cooling, Vh, and that for indoor air quality, Vp as the design flow rate of the supply air, V.

(17)

If V = v1, the air-handling system uses 100% fresh air.

Step (6): Calculate the Supply Air Temperature � can be determined by 'thtr-air temperature at the floor

level, Tp and the dimensionless temperature, 0/� ,·' . r.·; · 11;·;;, T !..'._ta Q,I (pC V) (18) �; L .... -.'.;:�· ,":1;: t·.'J";\;Jf. ''f · :;: . ,n. . r:··

where � . ',- . '.�·· :" · ., { .- · . !"l '( :l.' "· '

"!·,.·)l j '

· �:. Flefo/ T;. is the'ro6m ·desi;g'ri arr t�mper�tur�� ·'and ·i .'�, U .. fr ; !-; • •·,,_ •:·,"; ·j ' , _ ""1 , -�r <• I��-

•j ;, ii/r '"� '�'(0.29s'Q· '. +o.i3�cr :+:·o.1�sQ '/·n'. p(20) hf pC V . ,�, a,e · r · ' , .. I · ?." ,.eli, ... »· , •

p • I ··: I . . , • • l -···· . .

r • .,. I' . . 'b_ • - •,.... . .':' The e1can be caldUiated.by 'Mundt's fonnula (1'992):

,.(''

;: .. ::

where a.r

(IP) (21a)

·? '' • � • �·

:ii! -'kaiativb treat tRuisfer�effrcient'ironlcidiling1to · : ·

floor (Btu/ft2·h·°F, W/m2·K),\ ·

.,.,, ! ' • r

; =,.aonveo.tive,heat transrercoefficientfrom the floor sw.face:to room:air (Btuffi2,Jl·�F;.W/n12·K). :s

·:. '1I'hfladiatlve aitd'. coil.vective:h�hl tran�fer caefficients can be folitfd· rn 'ASHRAE F1miidmentals'{'liSHRAE"l997)'. As Q rough estimate, the)l'are clo!le to'l Btu/(h·i'f · 0F) or 5W/(m2. K). •

4259 (RP.949)

Then the exhaust a,i,f,temperature, Te, can be easily calcu-lated via an energy balc:wce for the entire space: · ,, ;

(22)

In practice, one air-handling system often serves several (N)'rooms (N> 1). For easy control, the supply air temperature should be the same for these rooms. To ensure the therrrial comfort in all the rooms, we choose the highest supply air temperature calculated by Equation 18 as the supply air temperature for all the rooms:

where 'I'ss = supply air temperature for all the rooms, 'I's,i = supply air temperature calculated for room i.

(23)

In this case, the supply flow rates for these rooms, Vi (i = 1, . . . , N), have to be recalculated. For simplification, V; can be estimated from

where Q,,i

!J:. ... j

:t::fcooling load in the itA. roontt .•..

(24)

Te. i ; := exh�ust air temperature in the ith room calculated by Equation 22. , ·· 1 '· 1 • •

Since Tss � Ts. 1 (i = 1, . , ..• N),. the �upply flow rates with Tss are larger than the ones with 'I's. i· The air temperature differ­ences between head and foot levels with Tss are smaller-titan those with Ts,i· The V1 calculated by Equation 24 is not very accurate, but it is acceptable fur most ·designs. T11 for the room should be r�.�<fulated. ��it is �<?19 ,h_igh compare� �o the �esign room air temperature, Vi should be increased in small incre­n}�nts until,T,, i�

_ac<;eptabl�. An exa,�t. � c�_be calcula,ted by

solving Equations 1 8 through 22 together. · .

The total supply flow i;�te of the sy&tem is .· . . , ' � I ' ": ' • • � : � If .,# : .-:::.: �' J \

N v .. =; ·� v,.. ·-:- ; : '

' ' :;_,;, 1 :. _.,.·-.

(25)

Note that if the difference between the 'I's,; is larger than �0f �3 .K), the design

.er �h?u�d co_!lS�d� _l_l�ing t�o supply air

temperatures, Tss· This would allow so,me·1ef the spaces to be supplied with a lower air temp(:rafui:eiancfa smaller amount of supply air.

SteP,J!).:, q,t�rmi'1'1�e��t'!e, lja_ijo.�f t�' Frest),,Alr: to the Supply Air,''··:� .. · . . '/. ·c ,. ' ,-, •. :,·:'.,

· For the mom with hazardous; or toxic €Ontaminants, such as biotech ahd1cheniist:r.J laboratories, ·displacement ventila­tion system� �s.e 1 OQ% fre,sh a.i,r. er,.ry=:;. l_PQ%. ;r11e. .�pei;gy loss throug� �-exhaust ��·c:an..l�e r�Ryered qy a heat e�cqanger. For oth�r, ,pflces, su� 1as ,o:ffices, vy-� �rec;:ommend, the ll�e. of·.

filtered return air to;save·energy. The fresh ·air ratio is deter­mined from

·. � r1 = max V. , i = 1, . .. , N I

(26)

where Vt i = the required flow rate for acceptable indoor air quality determined at Step 4.

Step (8): Select Air Supply Diffuser Size and Number ·111� supply air velocity has an upper limit to avoid draft.

Our investigation shows that the maximum face velocity of a diffuser is about 40 fpm (0.2 mis). Based on this face velocity, l .•.

the total face area of the diffusers for the ith room, A1, is

A;� V; I 4Q (fr), . .; ;n. A;= V/0.2 (m2). ��

(IP) (27a)

(SI) (27b) r"_ i t l �'�ft I

If each diffuser has a fa�e ar�a of Ad, the .tJ.µ\fl��r pf dif:filsc ers neeqed for the room, Nd• is · •?.

, . (28)

According to the space layout, the designer should place the diffusers in the room with the following rules:

•

•

•

•

There should not be large obstacles near the diffusers . The diffusers should be plaeed'on the· walls opposite to the exterior walls/windows. The diffusers can be placed in the center of a room, such as around a column.,, More diffusers should be placed in the spaces with higher cooling load.

)i Step (9): :check the Winter Heating Situation

The di pia1 em'ent velltilation system is ideal for'c��ling. The systerti ·should nod5e usect for he£ting bee�use the buoy�'. ancr and low air supply ;.:elocity will drive the hot supply air" to th ceiling lev�l. However, Straub (1962) used a diffuser to dischilrge:hot ai� horizontally at floor level for heating. With' a'blgh air su 1 ply ve.lpcicy,'·tbe air temperature in the room undet the l:teating condition can be rather uniform.

For a building·p'ehmeter where beating may be .IJ ecessary � .wiritit; �� �epara. C. 9eatihg systerh Js necessary to 8ffset.�j. heatfug load 'from exterior walls' and windows when Oisplace-rhent vehtilatioi{1s used. ,, : ,�,

,')( The peri,meter heating system can be baseboard convec-tots, 'rildiators, and fan coil units at floor ievel, etc. The heating capacity should be slightly larger than the winter design heat­ing load from exterior walls and windows. This will ensure an upward flow near the exterior·wallsiwmdows. The perimeter heating device should be placed near the exterio�. walls and window"s'4a.t floor 1�vel. if a fan ·coil nit is used, �the airtiaw direction should be upwWd. :,A :cc _,,.. 1-' .: ·-

A displacement ventilation system.for winter.heating will cQ.p.dition the!:t-Oom air as if.the room has only ari internal c�>0l-

11

ing load. Although heating (by a perimeter heating system) and cooling (by displacement ventilation) are used simulta­neously in the winter, energy i not wasted because the supply air temperature is still much higher than the outdoor air temperature. The outdoor air must be heated to the supply air temperature in any system. The ventilation air is heated, but not to as high a temperature as that in mixing ventilation.

Step (1 0): Estimate the First Costs and Annual Energy Consumption

We are analyzing first costs and energy consumption for four U.S. climatic regions and the result'> will be available soon. According to the literature review (Yuan et al. 1998), the first costs of displacement ventilation may b� slightly higher than those of mixing ventilation, and the energy consumption was found to depend very much on the conttol strategies and air-handling systems.

CONCLUSIONS With proper design, displacement ventilation can main­

tain a tl1ermally comfortable environment. The air velocity is smaller than 40 fpm or 0.2 mis. The temperature difference between the head and foot level of a sedentary occupant is less than 3.5°F (2 K). The percentage of dissatisfied people due to draft (PD) and the predicted percentage of dissatisfied are less than 15%.

Compared with conventional mixing ventilation, displacement ventilation provides better indoor air quality when the contaminant sources are associated with the heat sources. The thermal plumes bring the contaminants to the upper zone, and the contaminant concentrations in the lower zone are lower. The mean age of air is younger and the venti­lation effectiveness is higher in a room with displacement ventilation than in a room with mixing ventilation.

A high cooling load in a room would require a high venti­lation rate, which may limit the application of displacement ventilation. An air-conditioning system with a high ventilation rate is not energy-efficient and is expensive.

This paper presents a ten-step design guideline for displacement ventilation. The guideline was developed from computed results and measured data of high cooling loads and are, therefore, especially useful for U.S. buildings.

12

ACKNOWLEDGMENTS The U.S. National Science Foundation partially

supported this research through grant No. CMS-9623864.

REFERENCES

ASHRAE. 1989. ANSUASHRAE Standard 62-1989, Ventila­tion for acceptable indoor air quality. Atlanta: Ameri­can Society of Heating, Refrigerating and Air­Conditioning Engineers, Inc.

ASHRAE. 1992. ANSUASHRAE Standard 55-1992, Thermal environmental conditions for human occupancy. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

ASHRAE. 1997. 1997 ASHRAE Handbook-Fundamen­tals. Atlanta: American Society of Heating, Refrigerat­ing and Air-Conditioning Engineers, Inc.

Fanger, P.O., A.K. Melikov, H. Hanzawa, and J. Ring. 1989. Turbulence and draft. ASHRAE Journal 3 1 (7): 1 8-23.

ISO. 1990. ISO Standard 7730, Moderate thermal environ­ments-Determination of the PMV and PPD indices and specification of the conditions for thermal comfort. Geneva: International Standards Organization.

Li, X., and Y. Jiang. 1 996. Calculation of age-of-air with velocity field. Paper presented at Post-IAQ 96 Seminar, Beijing.

Straub, H.E. 1962. What should you know about room air distribution? Heating, Piping, Air Conditioning, Janu­ary, 209-220.

Svensson, A.GL. 1989. Nordic experiences of displacement ventilation systems. ASHRAE Transactions 95(2).

Yuan, X., Q. Chen, and L.R. Glicksman. 1998. A critical review of displacement ventilation. ASHRAE Transac­tions 104(1).

Yuan, X., Q. Chen, L.R. Glicksman, Y. Hu, and X. Yang. 1999a. Measurements and computations of room air­flow with displacement ventilation. ASHRAE Transac­tions 105(1 ).

Yuan, X., Q. Chen, and L.R. Glicksman. 1 999b. Models for prediction of temperature difference and ventilation effectiveness with displacement ventilation. ASHRAE Transactions 105(1).

4259 (RP-949)