AABCAssociated Air Balance Council

TABj o u r n a lTABj o u r n a l

Serving the HVAC Test and Balance and Engineering Industries

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Duct LeakageTesting

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AABCAssociated Air Balance Council

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11

Understanding Temperature and Altitude Corrections . . . . . . . . . . . . 2Ron Schilling

Air Basics - For Design Intent . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4Jerry Lavender and Mike Van Weichen

Parallel Pumping System Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8Mario L. Perez

Preliminary Testing Before Remodeling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11Bob Severin

Static Pressure Set Points for VAV Systems . . . . . . . . . . . . . . . . . . . . 12Derek R. Shupe

Circuit Setter Problems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13Richard MillerA Comparison of SMACNA New Duct Leakage Test Criteria. . . . . . . . 14Laszlo A. Lukacs

Fresh Air and Why it May Not be so Good for You . . . . . . . . . . . . . . 15James P. Bragg

Proportional Balancing Air Handling Systems . . . . . . . . . . . . . . . . . . 17Mike Nix

Smoke Dampers - The Pressure Drop Dilemma . . . . . . . . . . . . . . . . 21Albert L. Englehart

C O N T E N T S

Associated Air Balance Council

Board of Directors and Officers

PresidentWilliam A. DerseProfessional System Analysis, Inc.Executive Vice PresidentPatrick H. KellyAmerican Testing Inc.

Secretary/TreasurerRobert A. ConboyAmerican Air Balance Co., Inc.

Vice President/Eastern Zone-1Joseph E. Baumgartner, III, P.E.Baltimore Air Balance Company

Vice President/Central Zone-2Mike YoungTest and Balance Corp.

Vice President/Western Zone-3Michael RenovichRS Analysis, Inc.

Immediate Past PresidentWilliam K. Thomas, Sr. P.E.Thomas-Young Associates, Inc.

Director, Canadian ChapterEd St. LaurentA.H.S. Testing and Balancing, Ltd.

Executive DirectorKenneth M. Sufka

TAB Journal EditorBrian G. Hutchings

Editorial Office1518 K Street, N.W., Suite 503Washington, D.C. 20005(202) 737-0202FAX: (202) 638-4833E-mail: aabchq@aol.comWebSite:www.aabchq.com

TAB Journal is published quarterly bythe Associated Air Balance Council. It isdistributed free to AABC members andby subscription to non-members at $24per year.

TAB Journal is an open forum for thefree expression of opinions and information. The views expressed arenot necessarily those of AABC, its officers, directors, or staff.

Letters, manuscripts, and other submis-sions are welcome. However, TABJournal accepts no responsibility forunsolicited material.

All rights reserved. Copyright 2000by the Associated Air Balance Council.

TABj o u r n a l

From the PublisherThough not a required test, the Associated Air Balance Council Building recommends that all duct systems, including low-pressure systems, be sealed and tested in accordance with its NationalStandards. In fact, AABCs new 2000 National Standards (scheduled for release later this year)will feature an updated and revised chapter dedicated to Duct Leakage. Though most buildingcodes normally require that ducts be sufficiently airtight to ensure energy conservation and controlof the air movement, humidity, and temperature in the space, problems with excessive duct leakageare widespread.This issue of TAB Journal, entitled Duct Leakage Testing, contains several articles focusing on thistopic. In the articles, we see how duct leakage is affected by static pressure, openings in the duct(through joints, seams, access doors, rod penetrations, etc), and workmanship, and how duct testingcan save money and improve indoor air quality. Among these, Laszlo Lukacs with AerodynamicsInspecting Company compares the differences between SMACNAs New Duct Leakage TestCriteria against their old one and reports that newer may not always be better. Albert Englehart,Mechanical Testing, Inc., presents an informative case study on problems with static pressure dropsacross smoke dampers, and the drawbacks of using smoke dampers in small ducts. And finally, JerryLavender and Mike Van Weichen of AIRWASO, show the importance of duct leakage testing andhow testing for duct leakage can save the owner considerable expense over the long run.In the Forum section, James Braggs humorously titled article, Fresh Air and Why it May Not beso Good For You, addresses the serious issue how adding too much outside air can actually lead topoor indoor air quality. In other articles, Mike Nix, of Delta-T, Inc., takes a look at the benefits ofproportional balancing air handling systems, and Mario Perez, Precisionaire of Texas, explains whyit is important to perform a thorough analysis of each parallel pumping system application. Thisissue of TAB Journal also includes a new edition of AABCs technical newsletter TechTips, and areport by the Department of Energy revealing that most commercial HVAC auxiliary equipment isnot energy efficient. We thank all of the authors for their contributions, and for helping to make this another informativeand educational issue of TAB Journal. We welcome reader input at TAB Journal and encourageyou to provide us with your comments, letters, and articles.

2 TAB Journal

F A N P E R F O R M A N C E

Understanding Temperature and Altitude CorrectionsR o n S c h i l l i n gGreenheck Fan Company

The most common influences on airdensity are the effects of temperatureother than 70F and barometric pressuresother than 29.92 caused by elevationsabove sea level.

Ratings found in fan performance tablesand curves are based on standard air,which is defined as clean, dry air with adensity of .075 pounds per cubic foot,with the barometric pressure at sea levelof 29.92 inches of mercury and a tempera-ture of 70F. Selecting a fan to operate atconditions other than standard airrequires adjustment to both static pres-sure and brake horsepower. The volumeof air will not be effected in a givensystem because a fan will move the sameamount of air regardless of the air den-sity. In other words, if a fan will move3,000 CFM at 70F, it will also move3,000 CFM at 250F. Because 250F airweighs only 34% of 70F air, the fan willrequire less bhp, but it will also createless pressure than specified.

When a fan is specified for given CFMand static pressure (Ps) at conditions otherthan standard, the correction factors(shown in table) must be applied to selectthe proper size fan, fan speed and bhp tomeet the new condition. The best way tounderstand how the correction factors areused is to work out several examples.Lets look at an example using a specifica-tion for a fan to operate at 600F at sealevel. This example will clearly show thatthe fan must be selected to handle a muchgreater static pressure than specified.

Example #1

A 20" centrifugal fan (20" BISW) isrequired to deliver 5,000 CFM at 3.0 inchesstatic pressure. Elevation is 0 (sea level).Temperature is 600F.

1. Using the chart, the correction factoris 2.00.

2. Multiply the specified operating staticpressure by the correction factor todetermine the standard air densityequivalent static pressure. (Correctedstatic pressure = 3.0 x 2.00 = 6". Thefan must be selected for 6 in. of staticpressure.)

3. Based upon our performance table for a20" BISW fan at 5,000 CFM at 6 in

wg. 2,018 frpm is needed to producethe required performance. (This nowrequires a Class II fan. Before the cor-rection was made it would haveappeared to be a Class I selection.)

4. The bhp from the performance chart is6.76.

5. What is the operating bhp at 600F?

Since the horsepower shown in the perform-ance chart refers to standard air density, thisshould be corrected to reflect actual bhp at thelighter operating air. Operating bhp = standardbhp 2.00 or 6.76 2.00 = 3.38 bhp.

Important: We now know the operatingbhp. However, what motor horsepowershould be specified for this fan?

Figure 1: It is acceptable to interpolate when exact temperatures or elevations are not shown in chart.

Air Temp.

F 00.87

0.96

1.00

1.06

1.15

1.25

1.34

1.43

1.53

1.62

1.81

2.00

2.19

2.38

2.56

2.76

1000

9.90

1.00

1.04

1.10

1.19

1.29

1.38

1.49

1.58

1.68

1.88

2.07

2.27

2.48

2.66

2.87

2000

0.94

1.04

1.08

1.14

1.24

1.34

1.44

1.54

1.64

1.75

1.95

2.15

2.35

2.57

2.76

2.99

3000

0.97

1.06

1.12

1.18

1.30

1.40

1.50

1.50

1.71

1.81

2.02

2.23

2.44

2.67

2.87

3.09

4000

1.01

1.111

.16

1.22

1.33

1.44

1.55

1.66

1.77

1.88

2.10

2.31

2.53

2.76

2.97

3.20

5000

1.05

1.15

1.22

1.27

1.38

1.50

1.61

1.72

1.84

1.94

2.18

2.40

2.63

2.86

3.07

3.31

6000

1.08

1.20

1.25

1.32

1.44

1.56

1.67

1.79

1.91

2.03

2.26

2.50

2.73

2.98

3.20

3.45

7000

1.13

1.24

1.30

1.37

1.49

1.51

1.74

1.86

1.98

2.09

2.35

2.59

2.83

3.09

3.33

3.59

8000

1.17

1.30

1.35

1.42

1.55

1.58

1.80

1.93

2.06

2.19

2.44

2.69

2.94

3.21

3.46

3.73

9000

1.22

1.34

1.40

1.46

1.61

1.75

1.88

2.01

2.14

2.27

2.54

2.84

3.07

3.33

3.58

3.86

10000

1.26

1.40

1.45

1.54

1.67

1.81

1.95

2.08

2.22

2.37

2.63

2.91

3.17

3.45

3.71

4.00

11000

1.31

1.45

1.51

1.60

1.74

1.99

2.02

2.16

2.31

2.45

2.73

3.02

3.31

3.59

3.87

4.17

12000

1.37

1.51

1.57

1.66

1.81

1.96

2.10

2.25

2.40

2.54

2.84

3.14

3.44

3.74

4.02

4.33

13000

1.43

1.57

1.64

1.74

1.89

2.05

2.20

2.35

2.51

2.66

2.97

3.28

3.59

3.90

4.20

4.53

14000

1.48

1.63

1.70

1.80

1.96

2.13

2.28

2.43

2.60

2.75

3.08

3.40

3.72

4.05

4.35

4.69

15000

1.54

1.70

1.71

1.86

2.04

2.21

2.37

2.53

2.71

2.87

3.20

3.54

3.88

4.21

4.53

4.89

0

50

70

100

150

200

250

300

350

400

500

600

700

800

900

1000

Air Density Correction FactorsElevation [ feet above sea level ]

3TAB Journal

4. The bhp from the performance chart is2.40.

5. What is the operating bhp at 6,000-ft.elevation and 100F air?

Since the horsepower selected refers tostandard air density, this should be cor-rected to reflect actual bhp at the lighteroperating air. Operating bhp = standard bhp 1.32 or 2.40 1.32 = 1.82 bhp.

In this example, we can use the correctedbhp because the fan is located at a givenelevation and will not be turned on untilthe attic temperature reaches 100F. Theresult is a 2 hp motor can be specified inlieu of a 3 hp motor.

Communicate YourCorrections

When a specified fan appears on the fanschedule, it is important to determine if thespecifier has already made the required cor-rections for temperature and altitude. Avoidconfusion by specifying at what tempera-

ture or altitude (or both) the static pressurewas calculated.

For example: 5,000 CFM at 600F and 6 in.static pressure at 600F (or 3" Ps. at 70F).Electronic fan selection programs, such asGreenheck CAPs are excellent tools tosolve both the selection and specifyingproblems. CAPs prompts the user to enterthe air stream temperature, the start up temperature, and the altitude. The fan withthe corrected conditions is then automati-cally selected.

Using CAPs will also guard against makingselections for fan types or models that arenot appropriate for the condition. This isespecially important for selections atextreme temperatures that require specialconsiderations for materials, motors, bear-ings, drives, and speed derate factors.

As demonstrated in the previous examples,for optimum system design and performance,it is important to understand and make theproper temperature and altitude corrections.

Example 1: The fan curve represents the fansoperation at both the corrected and specifiedconditions. Curves are plotted at standard air.

Example 2: The curve above from CAPS represents the fan density correction for example #2.

If a fan is selected to operate at high tem-peratures, the motor must be of sufficienthorsepower to handle the increased load atany lower operating temperature where theair is denser. Assume the air entering thefan at start up is 70F, therefore no correc-tion should be made. The starting bhpremains at 6.76 and a 7.5 hp motor isrequired. Note: bhp corrections are mostcommonly used for altitude corrections (seenext example) or when the starting andoperating temperatures are the same.

Example #2

A fan used at 6,000-ft. elevation to exhaust100F air from an attic space. A 30" rooffan (GB-300) is required to move 10,400CFM at .25 inch static pressure.

1. Using the chart the correction factor is1.32.

2. Multiply the specified operating staticpressure by the correction factor todetermine the standard air densityequivalent static pressure (Correctedstatic pressure = .25" x 1.32 = 0.33"static pressure. The fan must beselected for .33" static pressure.)

3. Based upon our performance table fora 30" roof fan (GB-300), 698 frpm isneeded to produce the required per-formance.

4 TAB Journal

H VA C S Y S T E M S

Air Basics - For Design IntentJ e r r y L a v e n d e r a n d M i k e V a n W e i c h e nAIRWASO

he basic elements of HVAC air systems design have becomeincreasingly absent from tender documents over the past 10 years. Theend result is that many HVAC air systems cannot be made to operate atdesign intent. Based on our experience, some probable causes are:1. Not enough time is allowed for design and construction. 2. The low bidder gets the job, which causes the contractors to take

short cuts and buy cheaper material with the result being the projectdoes not meet design intent.

3. Not enough adequately trained design and construction personnel tounderstand their role in a properly designed and constructed project.

4. Reliance on the control contractor and the balancing agency to tryto make a deficient system work.

In the following article, we identify some basic criteria that is oftenoverlooked in hopes of shedding light on this growing problem.

Fan Performance

The image in Figure 1 illustrates a fan arrangement used to measureAMCA rated performance as published in fan catalogues and providedby computer generated fan curves. Systems effect S.P. values must beadded to the already accumulated S.P. for ducts, coils, filters, etc. Thisis not a precise science, but ignoring systems effects will result in per-formance at less than design intent.

For a centrifugal fan at 2500 FPM inlet or outlet velocity:Arrangement Causes additional In.w.c. S.P.No discharge duct 0.5Elbow near outlet 0.6Elbow near inlet 0.8Inlet < 1 dia. from plenum wall 0.2Bearing, OSHA belt guard near inlet 0.3V.I.V. s full open 0.4Added Static Pressure 2.8" w.c.

Too often, it is not possible to correct a selection by increasing the fanRPM because the motor will overload or the fan wheel will exceed its

T

Figure 1: AMCA STD.210 Laboratory method of testingfans for rating

5TAB Journal

plus allowance for elbows too close toeach other= Required system S.P. for AMCA rating.

The conditions shown in Figure 2 shouldalways result in fan selection, where thesupply fan airflow rate is considerablygreater than the return fan airflow rate, inorder to allow for adequate outdoor air-flow rate.

When a supply fan on a constant volumesystem must operate against a wide rangeof static pressures due to filter loading anda wet/dry cooling coil, variable inlet vanesand duct static pressure controls should beinstalled to maintain a constant flow rate.

Duct Fittings

Figure 3 shows the best use of balancingdampers in a duct system. Manual volumedampers in the O.A., R.A., and E.A.mains permit the setting of equal pressuredrops so the automatic dampers canbetter modulate the flows. Quadrantvolume dampers (i.e. dampers with aquadrant locking device) are located inmains and branches with the opposedblade damper in the terminal being used for final trimming with no noisegeneration.

Figure 2: Supply - Return = Outdoor Air

plus allowance for elbows too close toeach otherplus allowance for dirty duct mounted filters= Required system S.P. for AMCA rating.

Note: Select the fan rating for sum ofmaximum resistance. Determine the flowrate and motor load at the sum of mini-mum resistances. Apply appropriate con-trols so design intent is provided overentire static pressure range.

Return Flow

Supply fan flow rate less sum of exhaust fansless flow to pressurized building + 0.05" w.c.less additional outdoor air for acceptableair quality= Required AMCA flow rating.

Return Static Pressure

Sum of duct route and terminal resistancewhich combine for greatest resistance.Not necessarily the longest route.plus percentage allowance for systems effectplus allowance for extra fittings aroundunforeseen obstacles plus allowance for smoke exhaust rout-ing of total airflow

RPM limit. This is often the case withCentrifugal Plug Fans (no housing)which are frequently selected to reducecosts. But be aware, this type of fansstatic efficiency is about 60%, whereas abackward inclined air foil is about 80%.The ducts connected to the plug fansupply plenum are usually square edgeopenings at 1800 FPM entrance velocityand a S.P. loss of 0.5 x VP = 0.1 w.c. orgreater, depending on the turbulence atthe duct inlet.

Vane axial fans with a flexible duct con-nection at the inlet will suffer a flow lossof 5-12% of the AMCA rating. This isbecause the fabric flexes inward andblocks airflow to the outer diameter ofthe blades. As a result, the flexible con-nection should be located at least onediameter from the fan inlet. Additionalenergy costs and/or poor systems per-formance can be significant when sys-tems effects or plug fan characteristicsare not considered.

Fan Selection

Supply Flow Rate:Sum of terminals less possible diversity plus percentage allowance for duct leakage= Required AMCA flow rating.

Supply Duct StaticPressure:

The Sum of duct route and terminal resist-ance which combine for greatest resist-ance. Not necessarily the longest route.plus percentage allowance for systems effectplus resistance of mixing vortexplus resistance of dirt loaded filtersplus resistance of heating coilplus resistance of wet cooling coil withextractorplus entry loss of duct at plug fan plenumplus allowance for extra fittings aroundunforeseen obstacles

6 TAB Journal

Figure 4 shows construction details ofdampers and branch fitting which areeconomical and permit effective balanc-ing. SMACNA publications on testingand balancing state:

Splitter type dampers offer little or nocontrol of air volume in ducts. Theyshould be used as air diverters only.

Manually operated volume dampersshould be installed in each branchduct to control the amount of air.

Register or diffuser dampers cannotbe used for reducing high air volumeswithout inducing objectionable airnoise levels.

When referring to splitter dampers andextractors, the ASHRAE FundamentalsHandbook 1989 states that these devicesare poor, and should not be used.Instead, the ASHRAE Handbook recom-mends the diverging tee branch asshown in Figure 4. Also, common sensedictates that no manual damper of anytype or shape should exist in duct sys-tems from the supply fan discharge tothe inlet of a pressure independent box,but occasionally such balancing devicesdo occur on the contract documents.Though pressure independent boxes arerated up to 3" w.c. inlet S.P., a duct with3" w.c. loss from the first to last boxwould be at least 1000 feet long, whichis highly unlikely.

The following are a few additional fittings that will create an unexpectedstatic pressure.

1. Two or more 90% elbows installedclose to each other.

2. A 90% elbow with a high aspectratio, i.e. very wide and very shal-low.

3. An elbow at the inlet of silencer.4. An open end suction duct without

flanges.

Figure 3: The best use of balancing dampers in a duct system.

Figure 4: Balancing dampers- diverging tee branch.

Transition when 90 branch airflow is 10% or more of main

Balancing damper herewhen this duct serves two or

more terminals

Handles on Adjacent sides withquadrant, locking device, and

mounting screws clear ofdamper movement

All drive shafts to extend 2beyondduct wall to prevent insulator from

hiding damper and also for mountingquadrant clear of vapor barrier

insulation on tube spacers.

Tee branch for supply return and exhaust systemsoperating up to 3 w.c. sp and 1800 fpm velocityon supply systems do not install in low pressure

locations downstream on main dust fittings.

Location w from main or downstream of reheat coil

APPLICATION1. Up to 1.5 sq.ft. blade area, 1000 fpm branch

velocity and 1/2w.c. sp across throttled damper,use single blade 24 max length and end bearing.

2. For airway areas from 1.5 to 3.0 sq.ft. use sev-eral single blades of 6 max width and 24 maxlength individually operated and located to func-tion in an opposed manner.

3. For branch conditions greater than either 3 sq. ft.airway area, 24 wide, 1000 fpm or 1/2w.c. spacross a throttled damper, ise gang operatedopposed blade damper and frame assembly withsingle exterior locking quadrant.

7TAB Journal

Duct Leakage

A duct or airway system consists of allsurfaces that enclose the airflow from thefan suction or discharge to the face of thediffuser, register or duct opening. Almostall duct systems leak, the amount of leak-age depends on the specified sealing andproven implementation of testing proce-dures. The allowable leakage shouldonly be in terms of the percentage of totalairflow at a pressure equal to highestoperating pressure that will occur in thetested section. A standard tested sectionwould be the main, branches and runouts,all capped but without cut outs for dropsto diffusers or registers.

A practical formula for duct leakage testing is:

x TOTAL SYSTEM OPERATING AIR FLOW RATE IN CFMx SELECTED DESIGN % LEAKAGE

This formula allows an assessment beforethe tendering of the related costs for extrafan power, cooling and heating of leakageair, added equipment capital costs, andbad effects on air quality. Unfortunately,leakage discovered during the balancingprocess is usually beyond remedy.

In my opinion, duct leakage test proce-dures should not be referenced toSMACNAs HVAC Air Duct Leakage TestManual, because it does not determine theleakage as a percentage of total flow.Also, the manual states that it is not rec-ommended that duct systems constructedto 3" w.c. class or lower be tested since itis generally recognized as not being costeffective. But, overall, duct leakage test-ing is very cost effective. This wasrecently proven on a project where a sep-arate price of $10,000 was tendered forleak testing of all duct work on a $5 mil-

lion hospital. This cost was only 1/50 of1% of the project cost. Costs savings inenergy provided a total payback after only14 months of operation.

Figure 5 illustrates how the amount ofduct leakage varies with the operatingstatic pressure in the duct. Therefore, it ispractical to leak test duct work at themaximum duct operating pressure, but notgreater. All duct work with a duct operat-ing pressure in excess of 0.25" w.c. S.P.and a length greater than 50 ft. should beleak tested, and the rate of leakage provento be within the specified allowable per-centage of total flow. This should be donebefore the ducts become inaccessible.

The bottom line is that excessive air leak-age from ducts will prevent HVAC sys-tems from reaching design intent. Afrequent location for excessive leakage isthe clearance between the terminal (dif-fuser or register) and the connected duct.This is especially true at exhaust/returnregisters where the O.B. damper in theregister must be used for balancing.

Figure 6 illustrates how an additional noleak fitting can stop all leakage betweenthe connecting collar duct and the register.Another location for excessive leakage isthe duct fittings around a reheat coil.Large air flow leakage will occur acrossheat wheel seals. A large pressure differ-ential up to 7" w.c. can occur when thewheel is installed on the discharge side ofthe supply fan and the suction side of thereturn/exhaust fan.

Conclusion

To avoid the problems addressed in thisarticle, it is important that the entireprocess of engaging professional services,systems design, tendering, construction,and testing and balancing needs to bedone with care at all stages by persons ofconsiderable experience and professionalintegrity. Otherwise, the end result willnot meet design intent, and that will provecostly in the long run.

CFM of allowable leakage

Sq. Ft. surface area of tested duct work

Sq. Ft. surface area of ductwork in entire system

Figure 5: Leakage per sq. ft. of accumulatedopenings

Figure 6: Usual leaky fit vs. No leak fitting

=

8 TAB Journal

Parallel Pumping System AnalysisM a r i o L . P e r e zPrecisionaire of Texas

P

P U M P I N G S Y S T E M S

arallel pumping systems are usedthroughout the HVAC industry as a meansto deliver specified flow rates when multi-ple chillers or boilers are utilized in thecentral plant design, and in secondarypumping systems with significant varia-tions in the building load. Parallel pump-ing systems can utilize two or morepumps to deliver the required total flowrate for the building. An analysis of theparalleled pumping systems needs to beperformed to ensure proper operation ofthe system during all modes of operation.In order to simplify matters, this articlewill focus on a two pump system, but thissame method can be used to determine theproper selection and operation of anynumber of parallel pumps.

In a two pump parallel pumping system,each pump is specified to deliver half thetotal flow rate required by the pumpingsystem at the same specified head pres-sure. Curve A in Figure 1 represents thepump curve for a single pump curve athalf the total system flow rate with Point 1as the operating point for this pump. Fromthe single pump curve, a paralleled pumpcurve can be generated by doubling theflow rate of the single pump curve at thesame head pressure. This method generatesCurve B along the paralleled pump curvewith Point 2 as the balance point with bothpumps in operation. (See Figure 1)

The next step in the analysis is to generate asystem curve using the design operatingconditions. This curve represents the flowrate versus head relationship for a particular

installation and set of pumps. The systemcurve is generated based on the equation:

H2/H1 = (Q2/Q1)2Where:H1 = Known HeadH2 = Desired HeadQ1 = Known Flow Rate Q2 = Desired Flow Rate

The application of this equation generatesa table which can be plotted directly ontothe specific pump curve being used in theinstallation. In our example, a pumping

system has been specified to deliver 750GPM at 100 ft. Hd. Therefore, each pumpwould need to deliver 375 GPM at 100 ft.Hd. The table for this example would beas shown below:

Flow Rate Head in Feet0 020 33540 47460 58180 671100 750120 822

motor horsepower selection to be differ-ent than if the pump were selected for asingle pump application.

In Figure 4, the motor selection for eachpump would be 15 HP with both pumpsin operation. A closer analysis of thesystem when a single pump is in opera-tion reveals that a 20 HP motor selectionis required, as indicated by Point 3,

during a single pump operation whenboth pumps have been balanced simultaneously. If a single pump opera-tion is anticipated for a specific applica-tion, then a 20 HP motor would allow the pumps to operate within the entirerange of this particular application. (SeeFigure 4)

Parallel pumping systems are a usefuland common application used in todaysHVAC systems. Misapplications andoversights can occur which could causeproblems in the operation of the system.A thorough analysis of each parallelpumping system application followingthe steps outline herein, can preventsome common mistakes encountered with these systems.

9TAB Journal

Plotting the data from this table on thepump curve, Figure 2, generates Curve Cwhich is the system curve for this partic-ular set of pumps. Notice that the systemcurve intersects the paralleled pumpcurve at Point 2 which is the systemoperating point with both pumps in oper-ation. Curve A represents the impellerdiameter for each single pump. (SeeFigure 2)

An analysis must now be performed todetermine the performance of the systemwhen it operates with a single pump.Single pump operation can be caused byroutine maintenance, control strategiesfor standby operation, or mechanicalfailure of one of the pumps. The opera-tion of the single pump which remains inoperation is determined by the intersec-tion of the single pump Curve A, and thesystem Curve C as illustrated by Point 3in Figure 3. As this intersecting pointillustrates, the flow rate of the singlepump has increased and the head pres-sure decreased. The new balance pointfor a single pump in operation for thisparticular installation is 615 GPM at 67ft. Hd. (See Figure 3)

The Point 3 on the impeller Curve Aneeds to intersect the system curve. If the selection falls below the endpointof the impeller curve, then the singlepoint operation can cause serious prob-lems such as cavitation, unstable opera-tion, or excessive vibration. Operatingthe pump beyond the endpoint of theimpeller curve could have serious conse-quences including voiding the manufac-turers warranty.

The analysis of the single pump balancepoint is also important due to the inher-ent characteristics of some pump motorsthat cause the amperage to increase withincrease in flow rate. This may cause the

Figure 1

Figure 2

Figure 3

Figure 4

This same method canbe used to determinethe proper selection

and operation of any number of

parallel pumps.

TAB Journal

Occasionally, AABCreceives short casestudy type technicalpapers from our members. These papersusually focus on observations made byAABC members work-ing on a project in thefield, in which theyexplain a certain prob-lem they have encoun-tered, and whatcorrective actions theyinstigated to overcomethat problem. Each ofthese papers presentscertain problems orchallenges to the testand balance profes-sional, and providesinsight into how thesesituations can beresolved.

These papers are rela-tively short but mayhold special appeal forothers involved with theeveryday experience oftesting and balancing.We therefore decided topublish these papers asa collection of articles inTech Tips, a technicalnewsletter inside TABJournal that can beremoved for yourconvenience.

In This Tech Tips:

Preliminary Testing Before Remodeling

Static Pressure Set Pointsfor VAV Systems

A Comparison of SMACNANew Duct Leakage TestCriteria

Circuit Setter Problems

A N e w s l e t t e r F r o m T h e A s s o c i a t e d A i r B a l a n c e C o u n c i l

11

Preliminary Testing Before RemodelingBob Severin, Kahoe Air Balance

If you can, imagine attempting to add a second floorto a building that was not structurally able to main-tain the additional load and the problems that mayresult from lack of preliminary inspections andresearch. Attempting such a foolhardy undertaking islike opening the door to problems and almost certaindisaster. Likewise, problems can also occur withexisting heating, air conditioning, and ventilating sys-tems that may not be able to handle the additionalload put upon them as the result of a remodeling job.

The need for obtaining preliminary systems test andanalysis is something that should be of concern toevery design professional considering adding to, ormodifying existing building environmental systems.Original design numbers, extracted from existingprints and old contract documents may no longer bevalid when the building is being considered forexpansion or remodeling.

We have seen many instances where new design per-formance levels for airflows and hydronic flows couldnot be obtained from existing systems because thesesystems were not performing up to the expectationsof the designers and owners. Reasons for this caninclude things such as the systems may not have everperformed up to the original design numbers, or theperformance levels may have deteriorated over timedue to equipment wear, dirt buildup and a multitudeof other problems that may not be readily evident.

If, during the planning stages and conceptual devel-opment, a system analysis and set of preliminarysystem performance levels are made, adjustments andallowances for equipment repair or replacement canbe included in the budget for the project.

The timing for preliminary testing and system inspec-tions should coincide with the beginning of thedesign phase for the proposed project. Far too often,the preliminary tests are only called for prior to dem-olition. Unfortunately, the general contractor, in hiszest to begin work, does not usually concern himselfwith such a trivial matter such as measuring the per-formance levels of the existing HVAC equipment. Asa test and balancing subcontractor, we often find our-selves prodding the mechanical contractor for a timewhen preliminary readings can be taken. If the pre-liminary reading can not take place prior to, orduring, the design phase, then they need to beaddressed in several areas of the specifications. Thegeneral contractor and the mechanical contractorshould both be made aware of the preliminary testsof the existing HVAC in their respective sections ofthe specifications as part of the general requirements.

An accurate set of preliminary readings, and systemsanalysis, can certainly help to ensure that the finalproduct is one with which the owner is satisfied.As with all planning, having the correct infor-mation pertaining to performance levels ofexisting systems at the start of the designprocess will AIP in the successful outcomeof the project at time of completion.

Static Pressure Set Points for VAV SystemsDerek R. Shupe, R and S Balancing L.C.

At a recent ASHRAE meeting, an owners representa-tive presented his thoughts on building commission-ing. He directed his comments primarily towardsHVAC problems, and how he thought that the com-missioning of these systems would solve most systemfailures. Indeed, I believe, in some instances, thatHVAC system failures can be prevented by commis-sioning these systems. However, the failures that theowner representative had touched on during his pres-entation were items that we, as an AABC Test andBalance Agency, try to address as our standard proce-dure of Total System Balancing.

The main complaint that the owners representativehad was the fact that Variable Frequency Drives onseveral projects did not function properly, and that thecontrol contractor and the testing and balancing tech-nician did not actually set up the drives to controlproperly. Even when the VAV boxes were satisfied,the drive was still running at 100%.

As an AABC Testing and Balancing agency, we are notsimply satisfied with the tasks of just proportionallybalancing the airflow. Rather, we feel that the checkout of the control system, as well as other integralparts of the system, is an important part of the TotalSystem Balancing process. Granted, items such asfunctional tests and start-up procedures (that are sup-posed to be part of the commissioning process)should remedy many of the problems that arise (suchas VFDs that do not function), or, in the very least,bring them to the attention of the appropriate tradesas non-performing items.

As a Test and Balance professional, and sometimesCommissioning agent, the responsibility of getting theVFDs to work with the control system based upon aset point lies with the test and balance technician andthe control contractor.

The Process

At the risk of sounding redundant to fellow Test andBalance Engineers, I will (for the benefit of theArchitects, Building Owners, and Design andMechanical Engineers that read TAB Journal) explainthe process of arriving at, and verifying, the set pointfor system static pressure.

After the system has been proportionally balanced andthe VAV boxes have been set for their respective airflows, and while the system is in the full cooling mode

(or set up for the diversity factor) and has been bal-anced to the acceptable percentage, locate the staticpressure sensing unit that was installed by the controlcontractor and do the following:

1. Drill a hole in the duct adjacent to the sensing element and measure the static pressure with amanometer.

2. Compare the actual static pressure with the staticpressure reading that the control contractor isshowing on the display for the particular system.(Note: If the two readings do not match, some calibration on behalf of the control is in order.)

3. Once the controls are calibrated, the reading thatis recorded at the sensing location becomes theset point for that particular system and should beprogrammed as the set point static pressure in thecontrol program or into the controller.

4. Once the set point has been installed into the control system, activate the cooling capabilities ofthe system. Begin to set back the thermostats totheir respective set points. While doing this, verifythat the VFD is responding to the increase in pres-sure due to the boxes being satisfied (outputs to thedrive can be monitored through the control system).

5. After all the thermostats have been adjusted totheir respective set points, and the cooling systemhas been activated, the drive should be modu-lated back to a 50%-60% range once everything issatisfied, depending upon design criteria.

Here are some other things to keep in mindwhen verifying the operation of the VFD:

1. Systems with large diversity factors may requirethat the drive run in the upper ranges of the driveoutput to the motor (70% - 90%).

2. If the system has a return fan, the performance ofthe VFD will have to be verified when the systemis operating on 100% outside air.

3. Verify that there is enough air at the ends of eachbranch of the system. If a box at the far extremitiesof the system is short on air, the set point mayneed to be adjusted to accommodate the short fall.

4. The above steps to arrive at the system set pointwill not work with extreme duct leakage.

5. Sensor location may dictate the amount of pres-sure that is needed to operate the system prop-erly. Get the 1" out of your mind, its a place tostart, but is not the set point for every system.

6. The drive manufacturers representative or start-uppeople may need to be present in order to verifyfull operation of the drive based on input andoutput criteria.

12 TAB Journal

...we are not

simply satisfied

with the tasks of

just proportionally

balancing the

airflow.

Tech Tips arewritten for andby our readers,members of theAssociated AirBalance Council.We thank themfor sharingtheir valuableexperiences andproviding solutionsto problems inour industry.

13TAB Journal

Conclusion

Not all projects come complete with a commissioningagent to verify every single component and sequence.As a result, it would behoove the test and balancetechnician and his company to make sure that thesystem is functional before leaving the project. This isnot to say that it is the test and balance contractors jobto baby sit the control contractor, but it makes per-fect sense for the two contractors to work together toarrive at the set point.

On a final note, be sure to record the set points in thetest and balance report for future use. I have returned tojob sites long after the test and balance had been per-formed to troubleshoot problems of one sort or another,and have found that the set point static pressure hadbeen reset back to the proverbial 1" w.c. that seems tobe written down somewhere in the control mans bibleas the set point for all systems around the world. In theend though, proper set point pressure comes down tothree words...location, location, location.

A Comparison of SMACNA New Duct Leakage Test CriteriaLaszlo A. Lukacs, Sr., Aerodynamics Inspecting Company

In the course of the duct leakage testing of a largerectangular duct section (see duct schematic), wewere unable to achieve the allowable duct leakagerate under the old SMACNA guidelines. As a result,we used the new SMACNA standard.

However, the use of the new SMACNA standardraising the allowable duct leakage rate on mediumand high-pressure duct sections from 1% to 4% cre-ates a technical loophole. This is revealed with thefollowing illustration of the two testing standards:

Fan Performance14,000 CFM at 4.5" WG total pressure delivered inthe section of the main duct per our schematic.1. The old method of duct leakage calculation per

AABC and industry standard.Allowable leakage 1% = 140 CFMTesting pressure 150% of 4.5" WG = 6.75" WG

Note: Presently constructed duct is not made for6.75" WG.

Test results at various duct pressures:A. At 4.5" WG/duct leakage = 475 CFMB. At 1.8" WG/duct leakage = 300 CFMC. At .8" WG/duct leakage = 140 CFM

It is very clear the allowable leakage at the requiredtesting pressure is 290% higher. To pass the leakagerate of 1%, the testing pressure needs to be reducedto .80" WG.

2. SMACNAs 1985 first addition manual duct leak-age calculates on the same duct section and atthe same fan performance data.

Duct leakage class table 4.1 = 12System testing pressure = 3.0 WG

Allowable duct leakage = 350 CFM25 CFM100 S.F. x 1,400 Sq. Ft./area

Comparison of the two duct leakage testing methods indicates that:A. 4 to 1 allowable duct leakage increase by the new

SMACNA method.B. Calculation of duct section surface area per the

new SMACNA method is time consuming andallows for possible mistakes.

C. The old method is simple and quickly testing criteria can be established.

As a result, on medium pressure VAV duct systems,we are using system design, external pressure as abase for duct section testing pressure, and it isreceiving acceptance from the design engineers.

14 TAB Journal

Do you have a Tech Tip that youwould like to share with our readers? If so, please contact

AABC at:

Associated Air Balance Council

1518 K Street NW, Ste 503Washington, DC 20005 Fax 202.638.4833E-mail: aabchq@aol.comhttp://www.aabchq.com

Circuit Setter ProblemsRichard Miller, Systems Testing and Analysis

When TAB Technicians set circuit setters, there oftenseems to be a built in error in the process - namely inthe play, or slack, in the indicating dial. This move-ment is derived during the manufacturing process, andthrough wear of the indicator and adjustment controldial during usage. I propose the following procedurethat can be used to help minimize error, and maintainconsistency and repeatability for Technicians making,measuring, and adjusting water flows during thehydronic balancing of HVAC Systems.

First, always adjust from either the full open or fullclosed position. Either method works, but be consis-tent in choosing one or the other. Second, alwaysmove in the same direction when adjusting the circuitsetter and stop at the point using the leading edge ofthe adjustment. If the adjustment requires changing, donot arbitrarily move the dial to the new location if itrequires changing direction. Instead, move it in theopposite direction, then move past the dial locationand return to the set point, again being sure to movein the initial direction, and reset the indicator using theleading edge.

The following example illustrates the errors that canoccur during a coil measurement with the play in thecircuit setter dial:

On a one inch circuit setter, measure a 10 foot differen-tial at a 40 degree setting and a 5 degree play in thecircuit setter dial. You will find that there will be a dif-ference of actual flow of 2.65 GPM vs. 2.05 GPM. Thismay not seem very large at first, but if you multiply thisby the number of circuit setters in the project, you willnote a significant difference. This difference calculatesout to be roughly 30 percent on a single circuit setter,and a 30 percent variation in water flow throughoutthe entire project - an amount that exceeds AABCs 10 percent tolerance by a wide margin.

If the Technician uses the proposed method, heshould be able to repeat his measurements, be assuredof the readings and set points in his reports, and avoiderrors shown in the example.

For many years, circuit setters were the best instrumentavailable for setting hydronic flows. However, becauseof changing technology and times, this is no longer thecase. Luckily, there is a better solution available today.This solution is the venturi type circuit setter. Why?First, the venturi is the closest to using an orifice plate.This is the standard used by gas, oil, and waterpipeline companies for measuring flows through pipesfor metering purposes. Second, the venturi type has aCv that remains constant whether the valve before it isfull open, or 40 percent open, and the valve is not partof the measuring device. Though many companieshave graphs and the balance wheel or disc for setting,measuring, and balancing, many of their graphs andcharts are designed for only 100 percent open. As aresult, they are not accurate at 40 percent, due to achange in the Cv. The venturi type does not have thisproblem. For these reasons, I would like to suggestthat AABC recommend the use of venturi type circuitsetters by the designers of the hydronic systems forbalancing. What do you think? Send your comments toEditor, TAB Journal at AABCs headquarters.

15TAB Journal

Fresh air and why it may not be so good for youJ a m e s P . B r a g gEnvironmental Balance Corp.

he publics concern over indoor air qualityhas increased at a steady rate over the last six orseven years. These concerns are partly due to themedia coverage of high profile Sick Buildingsand the litigious society that we live in, but alsopartly due to increased studies in this area thatdo point to a need to address this issue. This hadled to increased funding to find more buildingfriendly materials as well as investigate methodsto control the air quality in the building.

The most common means used to improve theindoor air quality is to dilute it with fresh airwhich has lower concentrations (it is hoped) ofcontaminants. The volume of fresh (outside) airneeded to control the indoor contaminants isoften based on a per person quantity. The plusside of this is that the buildings designed occu-pation levels easily lead to a required outside airquantity. The downside is that it is difficult toknow just what quantity of fresh air is actuallyneeded. Back in the seventies, it was thoughtthat 5 CFM per person of fresh air overall wassufficient to keep the air in a building at a safeand acceptable level. At low levels such asthese, there was only a minor impact on theresulting extra cooling/heating loads and the out-side air was most commonly relieved via thetoilet exhaust.

In the nineties, we started to see buildings withhigher and higher levels of outside air require-ments, usually between 15-20 CFM per person.Although the added load due to the outside airwas added to the coil, we started to see a variety

T

I N D O O R A I R Q U A L I T Y

These concerns are partly due to

the media coverage of high profile

Sick Buildings and the litigious

society that we live in...

16 TAB Journal

of problems in the buildings, the mostcommon being:

Building over-pressurization Severe negative building

pressurization at times Actual increase in occupant

complaints of poor air quality Mold and high humidity in building

Based on a number of buildings that ourfirm has been involved in over the years,we found the following reasons for themore common indoor air quality problems:

Over Pressurization: The over pressur-ization was often due to inadequate relief.Often we would find barometric damperson the transfer duct from the return to theoutside of the building. Either the baro-metric damper was set for too high a pres-sure, or the duct was too small. Lately, weare seeing more forced ventilation viaexhaust fans, sometimes tied to the returnducts, and the problem here is that the fansmay not be sized properly or dont run.

Negative Pressurization: The oppositeproblem, negative pressurization, wefound most often in environments withcooler weather. Most of these systemswere variable volume with relief viaexhaust fans. In the cool weather and withthe system in a heating mode, the totalairflow being moved by the airhandlerswas at its lowest. The problem buildingshad only an air shaft connected from aroof intake to the airhandler return. As thetotal airflow dropped, the outside airflowlikewise dropped. Since the relief is aconstant volume exhaust fan(s), the build-ing becomes negatively pressurized.The occupant complaints in the buildings

with 15-20 CFM of designed outside airper person were varied in nature andcause. Not surprising, we found most ofthe complaints occurred during heatingseason when people are inside moreoften. Using a CO2 meter, we did findinstances of high (above 800 PPM) con-centrations of CO2 which indicates inad-equate ventilation. The most commontype of system we found was a variablevolume system, and the most commoncause was low outside rates at the timewhen the airhandler was moving a lowpercentage of its overall capability.

Mold and high humidity in buildings wefind much more often, the more so sinceour firm is located in the southeast.Again, the reasons are varied, but manytimes we found the system functioningas designed, in that the building was notnegatively pressurized and outside air-flow rates were at design. The chief cul-prit was often the very thing that wassupposed to help the building, the highoutside airflow rate. The problem is oftenthat a particular building may bedesigned for a certain occupancy leveland the HVAC system is designed tomaintain comfort levels for those occu-pants, but, the HVAC system is alsodesigned to do that on the hottest days inthe summer. As a result, the cooling coilcapacity may be quite large compared tothe actual interior load. This would notusually be a problem if not for the large

amount of outside air being introduced tothe system every minute. The coolingcoil doesnt need to work very hard tomaintain space temperatures with theless than full load conditions, but, sinceit is not staying on much, it does notdehumidify well. This caused the humid-ity in the space to rise, which oftenresulted in the occupants feeling clammywho, in turn, had the thermostats low-ered to combat it. As this continues,there comes a certain point where moldbegins to grow.

To combat the various problems associ-ated with trying to maintain adequateindoor air quality, we have seen a varietyof implementations, but they all involveaddressing the main problem directly.That is, the fresh air is treated before it isintroduced into the buildings HVACsystem rather than assume that theHVAC system can take care of it in thecourse of maintaining the indoor spacerequirements. This often involves a dedi-cated system to temper and de-humidifythe fresh air and can include an enthalpywheel or some other form of heat recovery to lessen the cost of doing it.Alternately, reducing the amount of freshair needed by monitoring the CO2 levelsand only adding fresh air when it isneeded has shown to be a viable option,particularly for larger systems. Fresh airmay be good, but there may be timeswhen it is better left outside.

Fresh air may be

good, but there may

be times when it is

better left outside.

17TAB Journal

D U C T S Y S T E M S

Proportional Balancing Air Handling SystemsM i k e N i xDelta-T, Inc.

Proportional balancings virtue is thatsystems can be balanced at less thanactual design flow rates. With proportionalbalancing, part of a large system can beset up before other parts are completed,and once a damper is fixed, it need neverbe altered.

The basic assumption is that a duct systemwith terminals can be regarded as a fixedresistance network, each section alwayspassing the same proportion of the totalinflow to the network regardless of theabsolute value of the inflow. Thus, a ter-minal handling 10 percent of the total net-work flow passes 100 CFM at 1000 CFMtotal, 200 CFM at 2000 CFM total, etc.However, in a duct system, resistancedoes not change in the same ratio in allsections with change in flow rate. Theeffect is complicated by the influence ofthe Reynolds number on the frictionfactor. Tests showed this effect to be suffi-ciently small to have no practical signifi-cance on proportional balancing. The testswere performed by selecting a systemmade up entirely of straight duct losses(subject to greatest error from theReynolds number/friction factor) andsecond, a typical system with half frictionlosses and half dynamic losses in fittings.In the first case, the error in estimatedflow rate did not exceed 3 percent over aflow range of +140 to -60 percent, and theerror was less than 5 percent at -80 per-cent flow rate. Errors likely in practice,

Proportional Balancing Air Handling: Systems can be balanced at less than actual design flow rates.

reflected by the assumptions in the secondcase, were 1.3 percent for the same rangeand only 3 percent for a flow reduction of-80 percent. Therefore, this factor isinsignificant.

Dynamic effects can sometimes upset dis-tribution so that it does not flow the equalresistance change assumption. For exam-ple, a badly designed bend preceding abranch take-off can cause preferential flow

in one junction at certain air speeds; asplitter damper can produce the sameeffect. Systems with splitter dampers willnot stay in balance when fan or zone CFMis increased or decreased due to changes inair flow characteristics. To perform propor-tional balance, volume dampers must beinstalled. Balance is achieved by increasingor decreasing static pressure. Once bal-anced, any change in air flow is propor-tionally distributed to the system.

How it Works in Practice

The concept of percent-of-design, whichexpresses the ratio between the measuredflow rate at a terminal and the requireddesign rate, is an important one in propor-tional balancing. Terminals and branchesare dampered to work at the same percent-of-design. Attention to accuracy whenmeasuring air flow from terminals isimportant, and readings must be consistent-that is, repeatable. To achieve this, thesame instrument is used; the same methodof working out the air flow is employed;and, preferably, the same operator con-ducts the test so that personnel equationerrors are avoided.

Terminal Balance

Given the simplest case (a branch ductwith a row of terminals), the balance pro-cedures start in the normal way by open-ing all terminal connection dampers.However, instead of throttling the dampersnearest to the fan working toward the endof the duct, proportional balancing beginswith the low or last terminal. This terminal

18 TAB Journal

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becomes the index terminal. Leaving theindex terminal fully open, the terminalimmediately next to it upstream is throttleduntil its percent-of-design is the same asthe terminal. The technician then worksprogressively along the branch, comparingeach terminal in turn with the index termi-nal and adjusting the dampers so thatfinally every terminal works with the samepercent-of-design. The end or low terminalis always used as the index, and its damperremains fully open. An additional advan-tage of this method is the minimum throt-tling required to balance the system. (Seefigure number 1 for diagram.)

Since each terminal will be delivering thesame percent-of-design air flow, any subse-quent change in the total volume flow intothe branch (caused, for example, by throt-tling nearby branches) will still produce thesame proportional distribution of air intothe terminals.

Branch Balance

Exactly the same principle applies to thebalancing of branches by throttling theinduct dampers. Unlike the traditional

method, proportional balancing does notrequire direct air flow measurement in theduct itself. The percent-of-design at theindex terminal of the duct being tested iscompared with the index terminal on thelast branch duct in the system, referred toas the index branch. Again, the procedureis to work back toward the fan. Thus, thefirst branch damper to be adjusted will bethe one immediately next to the indexbranch; and, as in terminal adjustment, thedamper on the index branch is left fullyopen. If, for example, the recorded percent-of-design on the index terminal of theindex branch was 80 percent, and themeasured percentage at the index terminalof the branch being adjusted was 90 per-cent, the duct damper would be closeduntil both index terminals showed the samefigure. More air would be forced into theindex branch during this operation, and thepercent-of-design would rise at the indexbranch terminals. This would not mattersince both branches would then be runningat the same percent-of-design. Also, pleasenote that every terminal must be remea-sured after any changes are made to thesystem and that drastic adjustments willnullify any proportional balancing done.

19TAB Journal

With this method of branch balancing,using terminals only, the terminal flowmust be within +/- 10 percent of the totalbranch flow. On branches with many terminals, it may be necessary to averageflow at several terminals to obtain a representative sample.

To reduce the possibility of error magnification when working with airflowrates other than design rates (for both ter-minal and branch balance), brancheshaving about +/- 30 to 50 percent of thedesign air flow rate are usually selected,and other branches that may be outsidethese limits are dealt with later in the bal-ancing program.

Zone Balance: (Multi-zoneSystem)

The same method is used to perform zoneto zone balance. Proportional balancebegins with the low zone and the damperset fully open. The next low zone is throt-tled until the index terminal on each zoneequals the same percentage of design.Proceed to the next low zone until allzones are balanced. On zones with manyterminals, it is necessary to average theCFM and select a terminal that representsthe average flow as the index terminal.

Tolerances

Tolerances allowable in proportional bal-ancing must be realistic, take into accountinstrument repeatability, operator error, andselected so that a minimum number ofsteps are needed for damper setting. Moretolerance is allowable if all terminals dis-charge into a single space than if a numberof unconnected spaces are involved.

Balancing between terminals on the samebranch is not as critical as the degree ofbalance needed between one branch andanother.

Fan Volume

Since the balancing operation starts withfully open dampers, there is the possibilitythat certain fan types will take too muchpower and overload the motor. Therefore,the motor overload must be corrected byreducing the RPM or throttling fan vortexor discharge dampers before balancing.

Unless check measurements are necessaryin branch ducts where there is some doubtabout the consistency of terminal readings,or where the terminals are of mixed types,

the only in-duct traverse required is at theend of the balancing process. After systemproportional balance is complete, the mainduct is traversed to determine total CFMbeing delivered by the fan. The fan RPM isset to deliver total fan CFM within speci-fied tolerances.

And finally, bear in mind that the propor-tionally balanced terminal total CFM doesnot equal fan CFM. The reason for this isbecause it does not allow for duct leakage.This can be a major factor in long ductruns with unsealed ducts, which can leak at25 percent or more. The differencebetween balanced terminal and fan traverseCFM is an indication of duct leakage, butthis is not an accurate gauge and shouldnot be used.

Tolerances allowable inproportional balancingmust be realistic, take

into account instrumentrepeatability and

operator error

Figure1: Note- the changes in CFM at the Index terminal on this diagram are used to explain pro-portional balancing. When performing the balancing the technician may throttle several terminalsbefore the CFM changes at the index terminal.

Balanced 5 , 4 Read 160=1.20%Balanced 6 , 4 Read 170=1.25%Balanced B4 , 4 Read 200=100%

STEP 6, 4 Read 230=1.15%

Balanced 2 , 1 Read 160=.80%Balanced 3 , 1 Read 170=.85%Balanced B4 , 1 Read 200=100%

STEP 1, 1 Read 150=.75%

Balanced 2 , 1 Read 170=.85%Balanced 3 , 1 Read 200=100%

STEP 2, 2 Read 170=.85%STEP 3, 2 Balance 160=.80%

Balanced 6 , 5 Read 250= 1.25%Balanced B4 , 5 Read 200=100%

STEP 7, 5 Read 260=1.30%STEP 8, 5 Read 240=1.20%

Balanced B4 , 3 Read 200=100%

STEP 4, 3 Read 210=1.05%STEP 5, 3 Balance 170=.85%

Balanced B4 , 6 Read 200=100%

STEP 9, 6 Read 270=1.35%STEP 10, 6 Balance 250=1.25%

Balanced B4 , 1 Read 200=100%

STEP 12, A1 Read 170=.85%

Balanced B4 , 4 Read 200=100%

STEP 11,B4 Read 250=1.25%STEP 13, Balance B4 to A1

Shaded areas are terminal readingafter balancing of the next terminal in sequence.

Terminal Design CFM 200% = Percent of Design

Balance as Described

Branch ABalance 2 to 1,3 to 1

Branch BBranch 5 to 4, 6 to 4

Alternative ProcedureUsing this method, attention to accuracy when

measuring air flow terminals is important to insureterminals are balanced within ten percent.

Branch ABalance 2 to 1,3 to 2

Branch BBranch 5 to 4, 6 to 5

Branch BalanceBalance Branch B Index to Branch A Index

Calculate Target CFMTarget CFM is the estimated CFM of the next terminalto the balanced.Target CFM is required when balancing terminal withdifferent CFM values.

Actual Index CFM

DesignIndex CFM

DesignCFM

TargetCFM1 1 2 2=x

150 200 x 200 = 150

Terminal 1 & 2 Proportional Balance to160 CFM (80% of Design).

21TAB Journal

D U C T P R E S S U R E

Smoke Dampers - The Pressure DropDilemma A l b e r t L . E n g l e h a r tTBE, Mechanical Testing, Inc.

On a project our firm completed recently, we ended up with asystem that had almost twice the static pressure and brake horse-power than the original design. Investigation by the building owner,engineer, and our technician found that the static pressure dropsacross the smoke dampers were the major source of the problem.

In this particular system, the exhaust fan was located in the mechani-cal room with ductwork going to two floors. The longest duct run onthis system had four smoke dampers installed before reaching the endexhaust registers. One of these smoke dampers was in a 24" x 12"duct with a design of 1,985 CFM and a velocity of 990 FPM.Referencing the manufacturers pressure drop data (Figure 1), thisshould give us a pressure drop for the damper of between 0.08" and0.12" w.g. depending on the manufacturer. However, when we meas-ured the pressure drop across the smoke damper, it was 0.51" w.g.This is over five times what it should have been reading.

Next, we checked the duct work and found that the actual opening forthis damper was 213/4" x 6", or a free area of only 0.91 square feet. Toobtain the design of 1,985 CFM, we would now need a velocity of 2,180 FPM. If we again look at the manufacturers pressure dropdata (Figure 1), we can see that, at this velocity, we do indeed have apressure drop of between 0.40" and 0.60" w.g.

Consequently, when there are multiple smoke dampers in the sameduct line, we need to anticipate that the actual system may be signifi-cantly different than its original design. Once we discovered this prob-lem, we were careful in our plan and specification reviews to checkother manufacturers data. None of the data that we have reviewed todate reference anything other than 100% open duct free area.

We believe that the HVAC designer needs to be aware of this condition and be extremely careful when using smoke dampers in small ducts.

Figure 1: AMCA STD.210 Laboratory method of testingfans for rating

23TAB Journal

A A B C V I D E O

New AABC Training Videos Soon to be ReleasedAABC

The need for trained and qualified TAB technicians is greater thanever. Additionally, because of demanding schedules, it is becomingincreasingly difficult for company owners to find the time to conductformal training sessions. That is why AABC has developed a seriesof training videos for use by AABC member companies.

The first set of AABC Technician Training Videos should be availableto members sometime this summer. The first set consists of an intro-duction to AABC with three training modules for AABC technicians.The modules include: How to take a Pitot Tube Traverse; Balancing aSingle-Zone System; and Balancing a Multi-Zone System.

While the first series of training videos begins with some of the verybasics of testing and balancing, members who had an opportunity toview portions of the videos during the recent Zone Meetings, wereimpressed with the results. Many indicated an interest in purchasingthe videos as soon as possible for new as well as more seasonedtechnicians.

While the videos are not intended to replace formal training pro-grams sponsored by the company, members quickly acknowledgedthat using the videos would save a lot of time and effort in helping tobring new technicians up to speed on the fundamentals of testing andbalancing. They further agreed that because of the professionalgraphic presentation, the videos would serve to enhance employeesmorale and help technicians learn more easily and quickly about theworld of testing and balancing.

While final changes are being made to the first series of videos, plansare underway for the development and production of 4 more trainingvideos at a more advanced level. Subjects will include VAV systems;Pneumatic, Electrical and DDC Controls; Hydronic Systems; andSpecial Systems, such as fume hood exhausts, cleanrooms, etc.

Currently efforts are underway to duplicate, package, price and distribute the first series of videos. A distribution date of August 1st is projected.

Package design for theforthcomingTechnicianTraining video series

24 TAB Journal

Department of Energy Report Shows CommercialHVAC Auxiliary Equipment NOT Energy Efficient

corrections & commentsThank you for the addressing the sub-ject of Steve Youngs article Ode to aMultizone (Winter 2000 TABJournal). However, Steves article didnot mention the rooftop multizonemanufactured by the Carrier Company(48MA and 50ME). The design fea-tures a pre-cooling coil, individualzone gas or electric heating, and indi-vidual cooling coils. The big advantageis NO ZONE DAMPERS! The zonecontrol is heat-bypass cooling, cool-bypass heating, what could be moresimple? We just do not understand whyCarrier does such a poor job promotingtheir product.

Sal CosentinoEnergy Applications

Correction: In the Winter 2000 issue ofTAB Journal, figure #1 on page 17 is labeled Balometer application. Itshould read Tachometer application.

TAB Journal welcomes submissions for publi-cation. TAB Journal is published quarterly by

the Associated Air Balance Council. Send letters or articles to:

Editor TAB Journal1518 K Street, NW, Suite 503 Washington, DC 20005

Have an opinion?An interesting case study?A new method?Tell us about it.

A R E C E N T DE PA RT M E N T O F EN E R G Yreport reveals that energy consumption by auxil-iary equipment in commercial building heating,ventilating and air conditioning (HVAC) systemsis much larger than previously thought, about 1.5quadrillion British Thermal Unit (Btu), equiva-lent to the energy used by 23 million automobileseach year. This energy is used mainly to operateequipment, such as fans and pumps that support the primaryHVAC equipment, distribute heating and cooling, and pro-vide ventilation for buildings and offices.

The findings reveals that the energy used by auxiliaryHVAC equipment is comparable to about 10 percent of allenergy used by commercial buildings in this country. Indoorair quality concerns, added filtration, and increasing ventila-

tion rates have increased the energy used for airmovement in buildings. Energy consumption byauxiliary HVAC equipment appears to be increas-ing, while primary equipment - largely chillers - isbecoming much more energy efficient. Updateddesign trends, such as greater use of series fanboxes and possible reduced duct cross sectionsalso tend to increase energy use.

This report will assist DOE in identifying and implementingmore energy saving technologies for its building equipmentresearch and development programs. The full report, EnergyConsumption Characteristics of Commercial Building HVACSystems, Volume II: Thermal Distribution, AuxiliaryEquipment; and Ventilation, is available as a pdf file atwww.eren.doe.gov/buildings/documents.

A A B C N A T I O N A L M E M B E R S H I P

ALABAMAAlabama International Test andBalance (A.I.T.B.), Inc.

Columbiana, Alabama (205) 669-7834

Southwest Test and BalanceCleveland, Alabama(202) 559-7151

Systems Analysis, Inc. Birmingham, Alabama (205) 802-7850

ARIZONAArizona Air Balance Company

Tempe, Arizona (480) 966-2001

General Air Control, Inc. Tucson, Arizona (520) 887-8850

General Air Control, Inc. Mesa, Arizona (480) 964-0187

Penn Air Control, Inc. Tempe, Arizona (602) 438-2664

Precisionaire of Arizona, Inc. Phoenix, Arizona (602) 944-4644

Systems Commissioning & Testing, Inc.

Tucson, Arizona (520) 884-4792

Technical Air Balance, Inc. Phoenix, Arizona (623) 492-0831

CALIFORNIA(ABCO) Air Balance Company, Inc.

Fullerton, California (714) 773-4777

American Air Balance Co., Inc. Anaheim, California (714) 693-3700

American Air Balance Co., Inc. Canoga Park, California (818) 703-0907

American Air Balance Co., Inc. Poway, California (760) 737-0190

Carter Air Balance, Inc. Napa, California (707) 252-4859

Circo System Balance, Inc. Sacramento, California (916) 387-5100

National Air Balance Co., Inc. Fremont, California (510) 623-7000

Penn Air Control, Inc. Cypress, California (714) 220-9091

Penn Air Control, Inc. Petaluma, California (707) 763-7155

Penn Air Control, Inc. San Marcos, California (760) 744-2951

Precision Air Balance Co., Inc. Anaheim, California (714) 630-3796

RS Analysis, Inc. Folsom, California (916) 351-9842

San Diego Air Balance Escondido, California (760) 741-5401

San Diego Air Balance Fullerton, California (714) 870-0457

Technical Air Balance, Inc. San Diego, California (619) 737-6817

Winaire, Inc. Huntington Beach, California (714) 901-2747

CONNECTICUTJames E. Brennan Company, Inc.

Wallingford, Connecticut (203) 269-1454

CFM Test & Balance Corporation Bethel, Connecticut (203) 778-1900

FLORIDAAir Balance Associates, Inc. Altamonte Springs, Florida (407) 834-2627

Bay to Bay Balancing, Inc. Lutz, Florida (813) 971-4545

Bay to Bay Balancing, Inc. Orlando, Florida (407) 599-9151

Bernie Moltz Inc. Rockledge, Florida (321) 631-6411

Environmental Balance Corporation Jacksonville, Florida (904) 724-7881

Perfect Balance, Inc. Jupiter, Florida (561) 575-4919

The Phoenix Agency, Inc. Lutz, Florida (813) 908-7701

Southern Balance, Inc. Pensacola, Florida (904) 433-8933

Southern Independent Testing Agency, Inc. Lutz, Florida (813) 949-1999

Test and Balance Corporation Tampa, Florida (813) 933-4171

Test & Balance Corporation of Orlando Orlando, Florida (407) 894-8181

GEORGIAHydro-Air Associates, Inc. Atlanta, Georgia (770) 997-1116

Tab Services, Inc. Atlanta, Georgia (404) 872-1861

Test and Balance Corporation Atlanta, Georgia (404) 255-8295

HAWAIITest and Balance Corporation of the Pacific

Honolulu, Hawaii (808) 593-1924

ILLINOISUnited Test and Balance Service, Inc.

Glen Ellyn, Illinois (630) 543-1210

INDIANAFluid Dynamics, Inc.

Fort Wayne, Indiana (219) 482-1326

IOWASystems Management & Balancing, Inc.

Des Moines, Iowa (515) 270-8755

KENTUCKYThermal Balance, Inc.

Lexington, Kentucky (606) 277-6158

Thermal Balance, Inc. Paducah, Kentucky (270) 744-9723

LOUISIANACoastal Air Balance, Inc.

Metairie, Louisiana (504) 834-4537

Tech Test Inc. of Louisiana Baton Rouge, Louisiana (225) 752-1664

MARYLANDAmerican Testing Inc.

Ellicott City, Maryland (800) 535-5594

Baltimore Air Balance Company Baltimore, Maryland (410) 661-2515

Baltimore Air Balance Company Annapolis, Maryland (410) 266-5854

Chesapeake Testing & Balancing Easton, Maryland (410) 820-9791

Environmental Balancing Corp. Clinton, Maryland (301) 868-6334

Test & Balancing, Inc. Laurel, Maryland (301) 953-0120

Weisman, Inc. Towson, Maryland (410) 296-9070

MASSACHUSETTSThomas-Young Associates, Inc.

Marion, Massachusetts(508) 748-0204

MICHIGANAerodynamics Inspecting Company

Dearborn, Michigan (313) 584-7450

Airflow Testing, Inc. Lincoln Park, Michigan (313) 382-TEST

MINNESOTAMechanical Data Corporation

Wayzata, Minnesota (612) 473-1176

Mechanical Test and Balance Corp.Maple Plain, Minnesota(612) 479-6300

Systems Management & Balancing, of Minnesota, Inc.

Coon Rapids, Minnesota (612) 717-1965

MISSISSIPPICoastal Air Balance of Mississippi, Inc.

Terry, Mississippi(601) 878-6701

MISSOURIEnvirosystem Analysis, Inc.

St. Charles, Missouri (314) 947-6324

Miller & Associates Testing & Balancing St. Louis, Missouri (314) 961-4018

Miller Certified Air St. Louis, Missouri (314) 352-8981

Precisionaire of the Midwest Grain Valley, Missouri (816) 228-3271

Senco Services Corp. St. Louis, Missouri (314) 432-5100

Systems Testing and AnalysisCreve Coeur, Missouri (314) 567-6011

NEVADAAmerican Air Balance Co., Inc.

Boulder City, Nevada (702) 255-7331

Land Air Balance Technology-LABTECH Las Vegas, Nevada (702) 385-5227

Penn Air Control, Inc. Las Vegas, Nevada (702) 221-9877

Raglen System Balance, Inc. Reno, Nevada (775) 747-0100

Technical Air Balance, Inc. Las Vegas, Nevada (702) 736-3374

Winaire, Inc. Las Vegas, Nevada(702) 262-9606

NEW YORKAir Conditioning Test & Balance

Great Neck, New York (516) 487-6724

Enercon Testing & Balancing Corp.New York, New York (212) 696-0760

Mechanical Testing, Inc.Schenectady, New York (518) 374-9440

Precision Testing & Balancing, Inc.Bronx, New York (718) 994-2300

NORTH CAROLINAAir Balance Corporation

Greensboro, North Carolina (336) 275-6678

The Phoenix Agency of North Carolina, Inc.Winston-Salem, North Carolina(336) 744-1998

Test and Balance CorporationWinston-Salem, North Carolina(336) 759-8378

OHIOAirdronics Inc.

Swanton, Ohio(419) 825-1437

R.H. Cochran and Associates, Inc.Euclid, Ohio (261) 731-0163

Heat Transfarr, Inc.Columbus, Ohio (614) 848-4303

Kahoe Air Balance CompanyEastlake, Ohio (440) 946-4300

Kahoe Air Balance Milford, Ohio (513) 248-4141

Kahoe Air Balance Lewis Center, Ohio (740) 548-7411

Kahoe Air Balance Dayton, Ohio (937) 433-8866

PBC, Inc. (Professional Balance Company)

Willoughby, Ohio (440) 975-9494

OKLAHOMAEagle Test & Balance Company

Cushing, Oklahoma (918) 225-1668

PENNSYLVANIAButler Balancing Company

Thorndale, Pennsylvania (610) 383-5104

Flood & SterlingHarrisburg, Pennsylvania (717) 232-0529

Kahoe Air Balance (Pittsburgh)McMurray, Pennsylvania (724) 941-3335

WAE Balancing, Inc.Mercer, Pennsylvania (724) 662-5743

PUERTO RICOPenn Air Control, Inc.

Humacao, Puerto Rico(787) 850-1866

SOUTH CAROLINAHall Technology, Inc.

Leesville, South Carolina(803) 532-1132

Palmetto Air and Water Balance Greenville, South Carolina(864) 877-6832

TENNESSEEEnvironmental Test & Balance Company

Memphis, Tennessee (901) 373-9946

Systems Analysis, Inc.Hermitage, Tennessee (615) 883-9199

United Testing & Balancing, Inc.Nashville, Tennessee (615) 331-1294

United Testing & Balancing, Inc.Knoxville, Tennessee(423) 922-5754

TEXASAerodynamics Inspecting Company

San Antonio, Texas (210) 349-2391

AIR Engineering and Testing, Inc. Dallas, Texas (972) 386-0144

Austin Air Balancing CorporationAustin, Texas (512) 477-7247

Delta-T, Inc.Dallas, Texas(214) 348-7430

Engineered Air Balance Co., Inc.Dallas, Texas (972) 239-4800

Engineered Air Balance Co., Inc.Houston, Texas (281) 873-7084

Engineered Air Balance Co., Inc.San Antonio, Texas(210) 736-9494

Precisionaire of TexasHouston, Texas

(281) 449-0961Professional Engineering Balancing Services, Inc.

Dallas, Texas 75355(214) 349-4644

UTAHR. and S. Balancing Company

Salt Lake City, Utah (801) 485-1411

VIRGINIAArian Tab Services, Inc.

Vienna, Virginia (703) 319-1000

TESCO, Inc.Chesterfield, Virginia (804) 739-6155

WASHINGTONEagle Test & Balance Company

Bellevue, Washington (425) 747-9256

Penn Air Control, Inc.Auburn, Washington (253) 939-4293

WISCONSINProfessional System Analysis, Inc.

Germantown, Wisconsin (262) 253-4146

AABC CANADIAN CHAPTERA.H.S. Testing and Balancing Ltd.

Winnipeg, Manitoba (204) 224-1416

Accu-Air Balance Co. (1991) Inc. Windsor, Ontario (519) 256-4543

Air Movement Services, Ltd. Winnipeg, Manitoba (204) 233-7456

AIRDRONICS, Inc. Winnipeg, Manitoba (204) 255-8449

Airwaso, Ltd.London, Ontario (519) 471-6678

Caltab Air Balance Inc. Essex, Ontario (519) 259-1581

Controlled Air Management Ltd. Moncton, New Brunswick (506) 852-3529

D.F.C. Mechanical Testing & Balancing, Ltd. Winnipeg, Manitoba (204) 694-4901

Designtest & Balance Co. Ltd. Richmond Hill, Ontario (905) 886-6513

Dynamic Flow Balancing Ltd. Oakville, Ontario (905) 338-0808

Equilibration dair Danco, Inc. Sherbrooke, Quebec (819) 823-2092

Equilibration dair Danco Quebec, Inc. Quebec, Quebec (418) 847-6049

Kanata Air Balancing & Engineering Services Kanata, Ontario (613) 839-2163

Pro-Air Testing, Ltd. Toronto, Ontario (416) 233-2700

Scan Air Balance 1998 Ltd. Moncton, New Brunswick (506) 857-9100

Scotia Air Balance 1996 limitedAntigonish Co., Nova Scotia (902) 232-2491

Systems Balance Limited Eastern Passage, Nova Scotia(902) 465-5554

Thermo Mechanical Services Ltd. Edmonton, Alberta (780) 451-4762

VPG Associates Limited Etobicoke, Ontario (416) 674-0644

INTERNATIONAL CHAPTEREnergy 2000 Technical Engineering

Seoul, Korea 82-2-408-2114

Penn Air Control, Inc.South, Korea3-493-7983

SUMMER 2000CONTENTSUnderstanding Temperature and Altitude CorrectionsAir Basics - For Design IntentParallel Pumping System AnalysisPreliminary Testing Before RemodelingStatic Pressure Set Points for VAV SystemsA Comparison of SMACNA New Duct Leakage Test CriteriaCircuit Setter ProblemsFresh Air and Why it May Not be so Good for YouProportional Balancing Air Handling SystemsSmoke Dampers - The Pressure Drop Dilemma