)

AT 90-17-2

ANALYTICAL EVALUATION OF OUTDOOR AIRFLOW RATE VARIATION VS. SUPPLY AIRFLOW RATE VARIATION IN VARIABLE­AIR-VOLUME SYSTEMS WHEN THE

OUTDOOR AIR DAMPER POSITION IS FIXED

S.A. Mumma, Ph.D., P.E. Member ASHRAE

ABSTRACT Most commercial buildings designed and built

today employ variable-air-volume (VAV) systems. For further energy conservation, the systems incorporate "free cooling" by using either a variable outside air (OA) (minimum to 100% of supply) economizer cycle or by using cooling tower water directly or indirectly in the cooling coils (referred to as water-side free cool­ing). The latter method does not provide controls to modulate the quantity of outside air.

The central thrust of this work was to investigate the variation in OA quantity vs. supply air quantity dur­ing periods when it was too warm or cold for efficient use of OA quantities above the minimum in systems with OA economizer cycles, or at any time in systems using water-side free cooling and fixed OA dampers. In other words, the specific periods of concern were when the OA economizer dampers had modulated to their minimum position and there was no cooling by outside air.

The OA deficiency problem occurred when the three dampers (outdoor, recirculation, and relief air) were in their fixed or minimum position. Systems are set up so that the desired flow rates occur at design conditions. At off design (i.e., supply air less than 100%), with no variation in the damper positions, the reduced quantity of recirculated air that should pass through the fixed-position recirculation air damper ex­periences a smaller pressure drop. As a result, the pressure both before and after the recirculation air damper changes. These changed pressures directly impact the pressure difference across the relief and outdoor air dampers. Such pressure changes ad­versely impact the flow of outdoor air into the building. Each of these pressures and the associated volumetric flow through each of the three dampers was modeled, assuming a tracking return fan which maintained a constant offset relative to the supply.

The analytical approach employed to model this

~':)~;"~~~ber ASHRAE Al ~I phenomenon was to first mathematically characterize the pressure difference vs. velocity for commercial mu/ti-louver positioning dampers. Second, the family of equations that characterizes the damper/system performance was written and solved using the New­ton-Raphson technique. The results up to this point revealed that a 30% reduction below the required min­imum OA occurred when the supply air quantity was reduced by 40%.

The second phase of the analysis involved the study of a commercial building in six geographic lo­cations to determine the hours of operation vs. supply air quantity during times when the economizer cycle control would call for minimum air. This phase pre­dicted a substantial number of hours with deficient quantities of OA in all six climate zones. While some engineers realize that the OA quantity in VAV systems varies with supply quantity, to the authors' knowledge this is the first time the extent of the problem has been documented for the practicing engineer.

The conclusion of this research clearly points to the need for modulating control of just the recirculated air damper (not the relief and outdoor dampers) in systems with economizer cycles when in the minimum air mode. Systems that use the cooling tower to pro­duce free cooling also need modulating control of the recirculated air damper.

INTRODUCTION

Heating, ventilating, and air-conditioning (HVAC) system designers have generally been aware that the flow rate of outdoor air in VAV systems responds some­what in proportion to the flow rate of supply air. In fact, tor many years VAV systems have been designed for 10% to 20% outdoor air. However, in recent years, considerable emphasis has been placed on indoor air quality (IAQ), with the result that the new ASHRAE Standard 62 (ASHRAE 1989) specifies a fixed cubic feet per minute (cfm) per person. Since the outdoor

Stanley A. Mumma is a Professor, and V.M. Wong is a Graduate Assistant in the Department of Architectural Engineering Pennsylvania State University, University Park, PA. ' THIS PREPRINT IS FOR DISCUSSION PURPOSES ONLY, FOR INCLUSION IN ASHRAE TRANSACTIONS 1990, V. 96, Pt. 1. Not to be reprinted in whole or 1n part wit~out written permissio~ of the American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc .. 1791 Tullie Circle, NE, Atlanta, GA 30329. Opinions, findings, conclusions, or recommendations expressed in this paper are those of the author(s) and do not necessarily reflect the views of ASHRAE.

air varies in proportion to the supply air, there will be times when Standard 62 cannot be met by conven­tional VAV systems. The central thrusts of this work were to quantify the magnitude of the outdoor air de­ficiency in VAV systems and to recommend resolution.

In the execution of this work, it was recognized that many VAV systems designed today employ either an outdoor air economizer cycle or a water-side free cooling cycle. Systems that utilize outdoor air econ­omizer cycles were considered to provide more than the Standard 62 requirement for outdoor air during free cooling regardless of the supply air.

PROBLEM DEFINITION AND APPROACH The basic central VAV system configuration con­

sidered in this research was as illustrated in Figure 1. Many variations to the illustrated cont iguration do exist, and most operat~ with the same inherent deficiencies as the one lllU tnned. The system consisted of a VAV SUfl>J!i>IY air fan w h a: tracking return air fan and three dampers I beled outside air, relief air, and recirculated air. It was assumed that the control dampers were properly sized and positioned such that at design sup­ply, the desired minimum outside air was drawn through the outdoor air damper.

Now, to characterize the problem, assume that at design the conditions of Figure 2 prevailed. Further assume that the dampers were positioned for minimum outdoor air (that is, the outdoor air temperature was either too hot or too cold for free cooling to be eco­nomical). Under these conditions, a negative pressure would exist just inside the outdoor air damper and a positive pressure would exist just before the relief air damper. These pressure conditions would cause a fixed pressure drop across the recirculated air dam­per. The problem of reduced outdoor air at reduced supply air may be dramatized by considering a drop in supply air. If the supply air were to drop to 75,000 from the design rate of 121,500, then the designed quantity of recirculated air would drop from 99,500 to 53,000 while the desired outdoor air quantity remained constant. However, during minimum outdoor air set­tings when the three dampers remained fixed, the flow rate across the fixed recirculated air damper would be substantially higher than desired and the quantity of

Flow

Relief Air Damper

Return Air Chamber

• 'R- ci rcul non

Damper

Alr

Mixing Air Chamber

Recitcualtion Air Damper

Low Limil

Return Fan

Supply Fan

Figure 1 Dampers set up for conventional VAV systems

Design CondiLion (supply Air = 121500 CFM)

Relier Air D1mper Pcrfotm111.ce : Pr·P• • Fun'tion(Rclicf Air flo.,. Vdochy, D1mpu l'Olil ioft)

lt~t •It J.lit• Rm~ g:;:; ~~i':',; ~q1Ft • 11000 CFM

Ycloril)'. :ioo FP~ ,. , .o.'i lft W& - ~lfl11flfl ~If '°"'""'" Air loiWi°' lb~•• ~I <P-• • ·- "-• ft ...... l .._..., - ....... . f.o~I

•nmt.UM i\lf ,, •• II •

1 • 119500 C""4 a.i.W Eak.wfll Alir ·~ lfut

P1 1 Jlllfurehll.. Vcloo::ity • 1776 FPM • 10000 CFM ........

~~~~;;!h~ri~n ~;mure) : tJl'p I j\~ ~0.,,1lidc Air) -,

I Reclrcul&1i011 Air Damper Pt'ffomunce : ~~;/;J- Pr·P• • Functian(Recircul" ion Air Flaw Vetaci1y, Damper P0:111li011)

01mpu POI ii Ion .. 0 67 Damp« Area • S6 Sq Fl.

Out.tide Air Aow lhv:

•llOOOCFM -

Vl'.locily • 144 FPM ~"------i.iJ-,U_=_:_ ........... '-/

' Ouuide Air O.mpcr Perf°'m1nu : Pi·l'I • FUM(li«!(Out.tide Air Flow Vdocily, O.mpcr Potilion) D1mper P1>1i1ion "'04 D1mpu Are.o • 64 Sq Fl.

Parl Load Condilion (Supply Air ~educed Lo 75000 CFM)

Relief Air D1mper Pcrform1nce: Pr.Pi • F11n~uon(Rd ief Air Aaw Velaci1y. D1mpe1 Pntilion)

Relief air Aow Rile:"-. g;;: ~-:'~ 1°:' ~~)Pl • S272 CfM

Vc:lodly'" Hll FPM ,.____ Pr•O l l In Wi - Rc1urn Air

r,.,,...,.. •• ..,,_.t_...,., tp ......... '"'""" "• o. .... , <J"PP11 ""' .. , ... , u~ ..... ,

" ' (A1moapheric Preuure) : Auumcd O lnWa 1

l)lialdir A.If i-iowlt.nt • IS272CFM -

ledr<lll11'-i Alir f.loow •"' • S91l1 CFM

Velocity• 2311 FPM .....-111o-----_,J.l-C>L--_........__.,

' Ouuide Air D1mper Pcrf.,munce : P1-Pt • ftJ11aion (Ou1.1 ide Air flow Vdacity, Dvnpc:r Poti1!an) Ovnpcr POlllian • 0.4 D1mpcr Atta• 64 Sqf\

Figure 2 Modeling process overview for Dallas/Fort Worth

outdoor air would drop below the desired figure. It is important to note that with a fixed outdoor air damper position, the only way to maintain a constant quantity of outdoor air in VAV systems as illustrated in this paper (Figure 1) would be to maintain a constant pres­sure drop across the outdoor damper. This must occur independently of the building pressure, supply fan op­eration, return air operation, or any other flow or pres­sure variable.

The approach used to analytically characterize the problem defined above consisted of the following four steps:

1. Mathematically characterize the damper per­formance, i.e., the pressure drop vs. flow for various damper positions.

2. Model the performance of the three dampers acting together to determine the relationship between outdoor air and supply air.

3. For a prototypical building, determine the hours of operation at the various supply airflow rates during both economizer and non-economizer cycle periods.

4. Put the results of steps 2 and 3 together to determine the total hours for which a given percentage or greater outdoor deficiency occurs.

)

OPPOSED BLADE DAMPER PERFORMANCE CHARACTERIZATION

Manufacturer's data (Johnson Controls Inc. 1988), as illustrated in Figure 3, were used to develop a third-order polynomial equation relating pressure drop to velocity and damper position. The manufac­turer's published curves were digitized and the least squares method (Stoecker 1980) was employed to generate the constant coefficients in the polynomial equation . In order to obtain a good fit it was necessary to break the data into two damper position regions­one for a damper position between 10° and 60° and the other for a damper position between 50° and 90°.

SIMULATION OF THE THREE DAMPERS ACTING TOGETHER

Given a particular set of designed airflow rates, damper positions, pressures, and damper perfor­mance characteristics, it was possible to analytically determine new values of flows and pressures for a given supply airflow rate. Specifically, the variables in the simulation were supply airflow rate, building ex­haust airflow rate, return airflow rate, return air cham­ber static pressure, relief airflow rate, recirculated airflow rate, outside airflow rate, and mixed air cham­ber static pressure.

Figure 3 Flow characteristics of control damper

The eight equations necessary to mathematically describe the system were solved by the Newton-Raph­son (Stoecker 1980) method. Aided by the simulation tool, it was possible to analytically predict the actual flow rate of outdoor air vs. supply airflow rate and the resultant percent deficiency. For example, when the supply airflow rate dropped from 121 ,500 to 75,000, the outdoor air brought into the building dropped from the required 22,000 to 15,221. That represents a 31 % reduction in the required outdoor air. The question at this point was, how many hours per year would the supply air quantity be at 75,000 or, for that matter, any other reduced rate. To answer that question, a pro­totypical commercial office building placed In six geo­graphical locations was studied.

SUPPLY AIRFLOW RATE VS. HOURS OF OCCURRENCE FOR A PROTOTYPICAL COMMERCIAL BUILDING IN SIX GEOGRAPHICAL LOCATIONS USING DOE 2.1 C

The building selected as the prototypical com­mercial building for evaluation in this research was a seven-story building with a single VAV system sup­plying all the floors. Figure 4 illustrates the plan and elevation views of the building. The building descrip­tion follows:

Window Heighl

Typical Floor Heighl

1-Rool Floor 6-Typical Floors

20·0· ~ __________ ~:.SJ<i.r!Q.• ___________ _ ,

Basemen!

Elevation

1· .. ..

··• .....

200'

Roof-Floor-North (RZ2) }15' ...... •••· u; OS

L\l 0 .Q u.. "ii 0 a:

7th Floor Zone Descripitioo

Typical-Floor-Norlh (TZ2) J1 5' •..... / J:i w

Typical Floor Zone Description

100'

100'

Figure 4 Seven-story sample office building, plan and elevation views

Shape:

Height:

Structure: Mechanical Space:

Exterior Walls: Roof:

Windows:

Ceiling:

Partitions:

Lighting:

People:

Equipment:

Rectangular 100 by 200 ft with the long facade facing north 7 stories, 11 .25 ft floor to floor, 8.5 ft ceilings, 2 stories below grade for storage Steel frame with concrete floors

Penthouse contained all mechanical, electrical, and elevator equipment. Penthouse was not modeled. Insulated curtain wall Built-up roofing on concrete, insulated Double glazed, 4 ft high on every floor, and ran the full width of each wall Suspended acoustical tile with lay-in fluorescent light fixtures Effective U-value = 1 .5 Btu/ h·ft2·°F Recessed fluorescent, 2.5 W/ft2, 20% heat to ceiling plenum 100 ft2 per person in exterior zones, 160 ft2 per person in interior zones, storage areas; unoccupied 0.5 W/ft2

The minimum outside air requirement was set at 20 per person according to ASHRAE Standard 62-1981 R. Basically, the system was operated with a tem­perature-controlled outside air economizer or no econ-

omizer at all. This approach allowed data to be generated which were used to study the outside air deficiency problem, not only from a system with an economizer cycle, but also the system using water­side free cooling instead of the economizer cycle. The system description is as follows:

• System One variable-air-volume system supplies the entire seven stories. Thermostatically controlled baseboards provide heating requirements. The system had a dry-bulb-controlled economizer cycle with a limit temperature of 72°F. The VAV boxes had a

• Design temperature

• Supply air Ventilation

• Fans

minimum stop of 20%. Cooling 78°F, heating 70°F.

55° at all times. 20 per person of outside air when the supply fan was on. Operated from 8:00 a.m. to 6:00 p.m. weekdays; off weekends and holidays.

This study, which was performed for six geo­graphic locations, included: Chicago, IL; Pittsburgh, PA; Miami, FL; Fort Worth, TX; Los Angeles, CA; and Phoenix, AZ.

The results of the annual supply airflow rate vs. the hours of occurrence for the six geographic loca­tions (Acrosoft International 1987) are presented in Figures 5 through 10. Figure 9, for Fort Worth, TX, illustrates the information obtained. Specifically, for the

M'NVALSL:PPLYAIR Fl.OW RATE Dl.STRml.JilON MlA-"11, FL

: •oo+---------_,,~ ~--------­;:ii" J5o~---------i u "'Joo+----------.-1. 250 j,.......-'-----'"--1

H 200+-------­o u 150.j.....;-----~

: 100 +-------

" +----=---:i 50 55 110 es 10 n eo 115 uo u 100 10s 110 115 120

SUPPi.. Y AIR Cf'M rTHOUSAND) 0 MIN. OA OPEil • EQ)NOMIZD.

Figure 5 Annual supply airflow rate distribution in Miami, FL

s , ... u "" ... L

II •oo 0 u j OO

• '

ANNUAL SlJPf'LY AIR Fl.OW RATE Dl.STRIBll'TION CHICAGO, IL

H 10 H oo d Ml II .. II "' 1' Ill II • .. •• 101 Ill Ill llP

SUPPLY MR C'FM (THOUSAND) 0 ~. OA OPEil • ecoNOMmR

Figure 7 Annual supply airflow rate distribution in Chicago, IL

' '• ' ::-; .. L

A ' " L

H 0 u •

'

' -..

&

A:-..-...1JAL SL'l'PLY AIR FLOW RATE 01.STRIBl.JllO,'I FQRnloQRTll, TX

;.

"" - ,,:_ I ~ ~ ·- - ... ... -... - - .... ... ,_

I< I' ,... .. ... ... ... ,...

' ., I I• ... ... ... ... ... ...

,, t• I" I'

I r ... ....... ... ...

I I I i ,_ ,_

·- ..

" -..

'l

... ,_

ITTft ...

~

.... I°"

SUPPLY AIR CFM (TllOUSAND) 0 MIN, OA OPEil • ecoNOMiml

Figure 9 Annual supply airflow rate distribution in Fort Worth, TX

supply airflow range of 65,000 to 70,000 (shown as 70,000 cfm) 121 hours would be under the economizer cycle and 64 hours would be under minimum outdoor air conditions, for a total of 185 hours. Coupling this information with that of outdoor air deficiency vs. sup­ply air, a tangible handle on the seriousness of the outdoor air deficiency problem was obtained.

HOURS OF OUTDOOR AIR DEFICIENCY

Integrating the work of the three previous sec­tions, one obtains the data in Tables 1 through 6 for the six geographic locations. Much of the information was also presented in Figures 11 through 22. Consider specifically Figures 19 and 20 for Fort Worth, TX, which illustrate the performance of a variable-air-volume sys­tem with and without an outdoor economizer cycle. That system (Figure 19) would operate with a 75,000 or less supply air quantity with the economizer cycle calling for minimum OA for 358 occupied hours per year (or the equivalent of approximately 45 working days) with at least a 30% or more outdoor air defi­ciency. On the other hand, a variable-air-volume sys-

" • u -• '

ANNUAL SUPPLY AUi. fU)W RATE DISTRIBl.ITION PHOE>llX.AZ

SUPPLY AD.CfM (THOUSAND) 0 MIN. OA OPEil • eCDNOMlZ!Jl

Figure 6 Annual supply airflow rate distribution in Phoenix, AZ

. H

... L

H -0

u •• • '

ANNUAL SUPPLY AIR FLOW RA'ln DLSTRIBUTION PrmiBURGll, f'A

11 ... 0 .. .. ... .. ,

SUPPLYAJRCFM('THOUSAND) OMIN. OAOPEJL .BCDtlOMZla.

Figure B Annual supply airflow rate distribution in Pitts­burgh, PA

ANNUALSUPl'LY AIR Ft.OW RATE 01.STRIBlITION.

u , ...

: ... µ~~~~::.a:~ ; 100 r--.:...--==..-:?9-+~

SUPPLY AIR CfM (THOUSAND) 0 MIN OA OPEil • eCDl«lMiml

Figure 10 Annual supply airflow rate distribution in Los Angeles, CA

tern without an outdoor economizer cycle (Figure 20) would operate with a 75,000 or less supply air quantity for 1540 occupied hours per year (the equivalent of approximately 192.5 working days) with at least a 30% or more outdoor air deficiency. As expected, the prob­lem was bad in VAV systems equipped with econo­mizer cycles, but was even worse for VAV systems without the economizer cycle. Another point worth not­ing from Figures 19 and 20 was that at 40,000 supply air and below, the quantity of outdoor air drawn in through the outdoor damper would be less than that exhausted in the building. Under the low supply airflow rate, outdoor air would be drawn in through the relief air damper because the pressure in the relief chamber had reversed from positive to negative.

CONCLUSIONS AND RECOMMENDATIONS

The conclusion of this work was that convention­ally designed, set-up, and operated VAV systems would experience the equivalent of months during which the flow rate of outdoor ventilation air would be well below that specified by ASHRAE Standard 62.

TABLE 1 Percent of Outside Air Deficiency vs. Annual Operating Hours in Miami, FL

Miami, FL

Supply Air Actual OA o/o OA Deficiency Annual Min. OA Cumulative + Cumulative * Quantity Quantity Off Design Operating Hours Annual Min. OA Annual Total Annual Total

(CFM) (CFM) Condition (Exclude Econo.) Ooerating Hours Operating Hours Operating Hours 120000 22000 0% 119 2433 120 2794 115000 21244 3% 100 2314 103 2674 110000 20517 7% 82 2214 84 2571 105000 19786 10% 120 2132 123 2487 100000 19054 13% 213 2012 218 2364 95000 18319 17% 373 1799 382 2146 90000 17583 20% 470 1426 483 1764 85000 16847 23% 415 956 443 1281 80000 16109 27% 297 541 343 838 75000 15373 30% 147 244 198 495 70000 14640 33% 72 97 123 297 65000 13910 37% 16 25 67 174 60000 13187 40% 7 9 54 107 55000 12473 43% 1 2 22 53 50000 11772 46% 1 1 31 31

+ This data reflects the time, for a system employing an air side OA economizer cycle, that the economizer is off and the dampers would be in the minimum position.

*This data reflects a system that does not employ an air side OA economizer cycle, and reflects the performance of a system that may employ water side free cooling.

TABLE 2 Percent of Outside Air Deficiency vs. Annual Operating Hours in Phoenix, AZ

Phoenix, AZ

Supply Air Actual OA o/o OA Deficiency Annual Min. OA Cumulative + Cumulative • Quantity Quantity Off Design Operating Hours Annual Min. OA Annual Total Annual Total (CFM) (CFM) Condition (Exclude Econo.) Ooeratlng Hours Ooerating Hours Operating Hours 145000 22000 0% 154 1649 154 2794 140000 21375 3% 113 1495 114 2640 135000 20765 6% 81 1382 83 2526 130000 20152 8% 99 1301 107 2443 125000 19536 11% 142 1202 151 2336 120000 18918 14% 182 1060 192 2185 115000 18299 17% 123 878 132 1993 110000 17678 20% 153 755 162 1861 105000 17059 22% 152 603 171 1699 100000 16439 25% 136 451 155 1528 95000 15822 28% 94 315 124 1373 90000 15209 31% 93 221 144 1249 85000 14601 34% 61 128 129 1105 80000 14001 36% 32 67 117 976 75000 13411 39% 25 35 124 859 70000 12833 42% 7 10 155 735 65000 12272 44% 3 3 167 580 60000 11732 47% 0 0 131 413 55000 10990 50% 0 0 135 282 50000 10335 53% 0 0 147 147

+ This data reflects the time, for a system employing an air side OA economizer cycle, that the economizer is off and the dampers would be in the minimum position.

•This data reflects a system that does not employ an air side OA economizer cycle, and reflects the performance of a system that may employ water side free cooling

TABLE 3 Percent of Outside Air Deficiency vs. Annual Operating Hours In Chicago, IL

Chicago, IL

Supply Air Actual OA % OA Deficiency Annual Min. OA Cumulative + Cumulative * Quantity Quantity Off Design Operating Hours Annual Min. OA Annual Total Annual Total

(CFM) (CFM) Condition (Exclude Econo.) Operating Hours Operating Hours Operating Hours 117500 22000 0% 12 1517 12 2794 115000 21581 2% 29 1505 44 2784 110000 20858 5% 25 1476 30 2738 105000 20129 9% 42 1451 47 2708 100000 19395 12% 64 1409 71 2661 95000 18650 15% 90 1345 104 2590 90000 17912 19% 111 1255 140 2486 85000 17164 22% 66 1144 116 2346 80000 16414 25% 48 1074 118 2230 75000 15662 29% 26 1026 130 2112 70000 14910 32% 20 1000 136 1982 65000 14159 36% 11 980 138 1846 60000 13413 39% 7 969 145 1708 55000 12676 42% 4 962 123 1563 50000 11952 46% 3 958 141 1440 45000 11249 49% 2 955 129 1299 40000 10550 52% 4 953 125 1170 35000 10000 55% 50 949 130 1045 30000 10000 55% 204 899 220 915 25000 10000 55% 695 695 695 695

+ This data reflects the time, for a system employing an air side OA economizer cycle, that the economizer is off and the dampers would be in the minimum position.

•This data reflects a system that does not employ an air side OA economizer cycle, and reflects the performance of a system that may employ water side free cooling .

TABLE 4

Pittsburgh, PA Percent of Outside Air Deficiency vs. Annual Operating Hours In Pittsburgh, PA

Supply Air Actual OA % OA Deficiency Annual Min. OA Cumulative + Cumulative • Quantity Quantity Off Design Operating Hours Annual Min. OA Annual Total Annual Total

(CFM) (CFM) Condition (Exclude Econo.) Operating Hours Operating Hours Operating Hours 115000 22000 0% 6 1578 8 2794 110000 21306 3% 18 1572 25 2786 105000 20548 7% 30 1554 38 2761 100000 19787 10% 42 1524 53 2723 95000 19022 14% 87 1482 102 2670 90000 18254 17% 116 1395 138 2568 85000 17484 21% 118 1279 163 2430 80000 16711 24% 124 1161 204 2267 75000 15937 28% 65 1037 141 2063 70000 15163 31 % 30 972 138 1922 65000 14389 35% 7 942 117 1784 60000 13619 38% 0 935 114 1667 55000 12853 42% 0 935 120 1553 50000 12096 45% 0 935 108 1433 45000 11351 48% 0 935 104 1325 40000 10600 52% 0 935 137 1221 35000 10000 55% 55 935 168 1084 30000 10000 55% 237 880 266 916 25000 10000 55% 643 643 643 650

+ This data reflects the time, for a system employing an air side OA economizer cycle, that the economizer is off and the dampers would be in the minimum position.

*This data reflects a system that does not employ an air side OA economizer cycle, and reflects the performance of a system that may employ waler side free coo Ii ng.

7 ·

TABLE 5 Percent of Outside Air Deficiency vs. Annual Operating Hours in Fort Worth, TX

Fort Worth, TX

Supply Air Actual OA % OA Deficiency Annual Min. OA Cumulative + Cumulatlve • Quantity Quantity Off Design Operating Hours Annual Min. OA Annual Total Annual Total

(CFM) (CFM) Condition (Exclude Econo.) Operating Hours Operating Hours Operating Hours 121500 22000 0% 86 1522 86 2794 120000 21715 1% 53 1436 54 2708 115000 21001 5% 54 1383 55 2654 110000 20283 8% 77 1329 78 2599 105000 19564 11% 136 1252 137 2521 100000 18842 14% 168 1116 171 2384 95000 18119 18% 160 948 169 2213 90000 17394 21% 188 788 205 2044 85000 16669 24% 123 600 151 1839 80000 15944 28% 119 477 148 1688 75000 15221 31% 97 358 148 1540 70000 14500 34% 64 261 185 1392 65000 13784 37% 27 197 141 1207 60000 13076 41% 7 170 145 1066 55000 12377 44% 2 163 147 921 50000 11693 47% 1 161 170 774 45000 10780 51% 0 160 142 604 40000 10000 55% 1 160 173 462 35000 10000 55% 30 159 127 289 30000 10000 55% 36 129 51 162 25000 10000 55% 93 93 111 111

+ This data reflects the time, for a system employing an air side OA economizer cycle, that the economizer is off and the dampers would be in the minimum position.

•This data reflects a system that does not employ an air side OA economizer cycle, and reflects the performance of a system that may employ water side free cooling.

TABLE 6 Percent of Outside Air Deficiency vs. Annual Operating Hours in Los Angeles, CA

Los Angeles, CA

Supply Air Actual OA % OA Deficiency Annual Min. OA Cumulative + Quantity Quantity Off Design Operating Hours Annual Min. OA Annual Total

(CFM) (CFM) Condition (Exclude Econo.) Operating Hours Operating Hours

112500 22000 0% 3 477 5 110000 21584 2% 1 474 4 105000 20825 5% 7 473 15 100000 20060 9% 21 466 25 95000 19289 12% 25 445 37 90000 18511 16% 57 420 74 85000 17729 19% 124 363 191 80000 16942 23% 171 239 324 75000 16152 27% 48 68 321 70000 15360 30% 19 20 495 65000 14568 34% 1 1 423 60000 13780 37% 0 0 344 55000 13205 40% 0 0 292 50000 12526 43% 0 0 144 45000 11890 46% 0 0 59 40000 11203 49% 0 0 29 35000 10549 52% 0 0 10 30000 10000 55% 0 0 2

Cumulative • Annual Total

Operating Hours

2794 2789 2785 2770 2745 2708 2634 2443 2119 1798 1303 880 536 244 100

41 12 2

+ This data reflects the time, for a system employing an air side OA economizer cycle, that the economizer is off and the dampers would be in the minimum position.

• This data reflects a system that does not employ an air side OA economizer cycle, and reflects the performance of a system that may employ water side free cooling.

)

.. f ' · rli \Rlllr.I~

' s - +--------~ r.

II

0

... 0 ... " ' • • 1 •• u •• 01 ... ' l ••

~ .. . ~ . . ..

I

:< u • •*+---'~::---------........ =::'."':"-~ L

1

~ ·­• H -0 u

.. .. , "o

• .. , I

"I 1 , , ' ., • ' .

Sll'PI v AIR CF\4 rntoL·sA. .. DJ St..'PPLY AIRCFM m1ot:S"1'1D)

Figure 11 Percent of outside air deficiency vs. cumulative annual minimum OA operating hours in Miami, FL, for a V AV system employing OA economizer cycle

0 - +----+--- • ... ,-------~-----; ' ' •

~ -- +----------- ·.-.1-111-111-mrm~ G

'I . ' Sll'PL\"AIACF'"1 1TIIOI SA.'1:0\

Figure 13 Percent of outside air deficiency vs. cumulative annual minimum OA operating hours in Phoenix, AZ, for a VAV system employing OA economizer cycle

' ' -' -­' 0 ·­, L

• I \: ..... •

·-·-· ...... ·- / 'OFO...DUKJI.:

..... Jit"" ., ·~ ......

' ' _!:,,,. A.: ~CA'-C\..'\tlt.An\'lffD(U= •,

/ "' ... •

I SL1'PlY ,\IRCF'.\1rTIIOL'SA . ..;O\

-·

• • • • • •

.. 1

,, " I

'i .. I

Figure 15 Percent of outside air deficiency vs. cumulative annual minimum OA operating hours in Chi­cago, IL, for a VAV system employing OA econ­omizer cycle

While a number of methods for correcting a problem as extensive and serious as the outdoor air deficiency presented in this paper do exist (Avery 1984 }, the au­thors will now present a method that has, to their knowl­edge, never been tried, used, or discussed in the literature (Haines 1982, 1986; Morris and Wiggin 1985; ASHRAE 1987; Dickson 1987; Alley 1988; Avery 1984). The outdoor air deficiency problem occurs because the pressure drop across the fixed-position outdoor damper does not remain constant in VAV systems. The authors propose that an effective way of maintaining the desired pressure drop, and hence the flow rate of outdoor air, would be to modulate the recirculated air

9

Figure 12 Percent of outside air deficiency vs. cumulative annual total operating hours in Miami, FL, for a VAV system employing water-side free cooling

• "

" -• • r I y ·-" .. n u

•

• Sl!PflLY A111. CFM(TI KMJSA.N'D)

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·• n I

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Figure 14 Percent of outside air deficiency vs. cumulative annual total operating hours in Phoenix, AZ, for a V AV system employing water-side free cooling

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nC I

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0 ~

~~~~ ... l-ll-llJ.11~~~11'1 ... iftfl~ .. t-mt-· ~

• • SUPflLY Anl CFM (TlfOOSAND)

y

Figure 16 Percent of outside air deficiency vs. cumulative annual total operating hours in Chicago, IL, for a V AV system employing water-side free cooling

damper. This may be accomplished (see Figure 23) in systems with an economizer cycle by opening a relay in the control signal to the outdoor and relief air dampers while controlling the recirculated air damper. Many control signals are possible for controlling a re­circulated air damper, including the following for feed­back control : • the pressure drop across the outdoor air damper, • the pressure drop across the relief air damper, • the pressure drop across the recirculated air

damper, • an outdoor airflow station, and • tracking control using the supply and recirculating

flow stations to maintain a fixed differential. A final open-loop control may utilize an algorithm

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0 -p

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T I .'II . ...,

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I Sll'PLY AIR Cf}4(l110CSAND)

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Figure 17 Percent of outside air deficiency vs. cumulative annual minimum OA operating hours in Pitts­burgh, PA, for a VAV system employing OA economizer cycle

A

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0

' " A ' T I

I c

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ft •I •I .. tt .. II "' " • ti • M , .. ... , .. " • ~H 411

SUPPLY AIM. CFM (THOUSAND)

Figure 19 Percent of outside air deficiency vs. cumulative annual minimum OA operating hours in Fort Worth, TX, for a VAV system employing OA economizer cycle

A •-¥-;:__:_e~,__...::::~-----__:___:~--_;_~ L

0

A -!---------'"'~----------+ '" F.

T ' I .,t N ·-+---"' _____ .._....__.. ____ """"=""-------t C

G I

SUPPLYAIRCFM(llf()USAND)

'Of.

' ' c y

Figure 21 Percent of outside air deficiency vs. cumulative annual minimum OA operating hours in Los An­geles, CA, for a VAV system employing OA economizer cycle

relating the recirculating air damper position to the supply air quantity, as illustrated in Figure 24.

At first glance it may appear that a fourth inter­acting local control loop would be added to the other three interacting local control loops (supply fan static pressure loop, the supply-return fan tracking loop, and the mixed air economizer dampers loop) which could contribute to further control instability. However, by taking a closer look, it may be seen that during the time the recirculating air damper would be controlled by itself, the economizer local control loop would be out of the picture. Therefore, this concept would not be expected to contribute further to control instability,

l <:>

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Figure 18 Percent of outside air deficiency vs. cumulative annual total operating hours in Pittsburgh, PA, for a VAV system employing water-side free cooling

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L F

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SUPPLY ArRCfN'(THOOSANO)

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Figure 20 Percent of outside air deficiency vs. cumulative annual total operating hours in Forth Worth, TX, for a VAV system employing water-side free cooling

. ·­L

~ -0 u • •

SUPPLY AIR CFM (llfOOSANOJ

HO

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A

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Figure 22 Percent of outside air deficiency vs. cumulative annual total operating hours in Los Angeles, CA, for a VAV system employing water-side free cooling

but would significantly improve the air quality in build­ings.

REFERENCES Acrosoft International , Inc. 1987. MICRO-DOE2 user's

guide. Denver, CO: Acrosoft International, Inc. Alley, R.L. 1988. "Selecting and sizing outside and return

air damper for VAV economizer systems. " ASHRAE Transactions, Vol. 94, Part 2.

ASHRAE. 1987. ASHRAE handbook-1987 systems and ap­plications, chapter 51 . Atlanta: American Society of Heat­ing, Refrigerating, and Air-Conditioning Engineers, Inc.

ASHRAE. 1989. ASHRAE Standard 62-1989, "Ventilation for acceptable indoor air quality." Atlanta: American Society

Relief

Flow

High Limit

Relief Air Damper SP

Return Air Chamber

I

~ I • - - IM~_._.__....,

I

IB1

Mixing Air Chamber

Outside Air Damper

Return Fan

S~Iy

Low Limit

Figure 23 Modulating control of the recirculation air damper to maintain minimum outside air

Modulating Positions of Recirculation Air Damper

0.7

D 0.6

a m

0.5 p e

0.4

p

0 0.3

0.2

0

n 0.1

S;.ilJply Air CFM (Thousand)

Figure 24 Modulating positions of recirculation air damper to maintain minimum outside air

of Heating, Refrigerating, and Air-Conditioning Engi­neers, Inc.

Avery, G. 1984. "VAV economizer cycle: don't use a return fan." Heating/Piping/Air Conditioning, August, p. 91.

Dickson, D.K. 1987. "Control dampers." ASHRAE Journal, June, p. 40.

Haines, R.W. 1982. Control systems for heating, ventilation, and air conditioning. New York: Van Nostrand Reinhold Co.

Haines, R.W. 1986. "Outside air volume control in VAV sys­tem.:· Heating/Piping/Air Conditioning, October, p. 130.

I?-

Johnson Controls, Inc. n.d. Flow rate characteristics of dam­pers, Db: 5 dampers. Milwaukee, WI: Johnson Controls, Inc.

Los Alamos Scientific Laboratory. 1980. DOE-2 reference manual. Los Alamos, NM: Group WX-4, Program Support, Los Alamos Scientific Laboratory.

Morris, R.H ., and Wiggin, M.E. 1985. "Indoor air pollution." Heating/Piping/Air Conditioning, February, p. 73.

Stoecker, W.F. 1980. Design of thermal systems. New York: McGraw-Hill.