Nyb Select Desing

ENGINEERING LETTERSThe New York Blower Company ●7660 Quincy Street, Willowbrook, Illinois 60521 -5530

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SUBJECT

System CalculationFan Laws and System CurvesUnderstanding Fan Performance CurvesTemperature and Altitude Affect Fan SelectionFan Performance - The System EffectIncreasing Fan PerformanceField Testing of Fan SystemsProper Selection of Pressure BlowersPneumatic ConveyingFans and Blowers for Combustion ProcessSelection Criteria for Fan DampersFan AcousticsFan Balance and VibrationStainless Steel Specifications for Fan EquipmentPractical Limits of Spark-Resistant ConstructionCorrosion-Resistant Coatings for Fan EquipmentCoating Surface Preparation SpecificationsCorrosion Resistance of FRP FansDesign and Construction of nyb FRP FansAccessories and Construction Modifications for FRP FansSurface Veil for FRP FansIntegral Motors for Centrifugal FansElectric Motor Codes and StandardsFundamentals of SteamIndustrial Steam Heating SystemsMiscellaneous Engineering DataEngineering Letter Glossary

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ENGINEERING LETTER 1The New York Blower Company ● 7660 Quincy Street, Willowbrook, Illinois 60527-5530

S Y S T E M CAL CUL AT I O NINTRODUCTION

A fan system is any combination of ductwork, hoods,filters,louvers, collectors, etc., that relies upon a fan to produceairflow.When the air moves past each of these components,resistance is created which must be considered in systemcalculations. It is also important to remember that fans are ratedindependently of a system and that fan performance will varydepending upon the accuracy of the system calculations. ThisEngineering Letter will explain some of the basic fundamentalsof system design and calculation.

SYSTEM DESIGN

The purpose of the system will dictate the design criteria to beused. Generally they will fall into one of the following twocatagories:

Velocity is typically the primary consideration in dustcollection, dilute pneumatic conveying, fume removal, andcontaminant applications. In these applications, a capturevelocity is required to redirect the flow of airborne materialsinto the duct system. In addition, a minimum conveying velocityis necessary to maintain the flow of the materials within thesystem.

Given these velocity requirements, system components can beselected to maintain the appropriate air volume andrequiredvelocity through the system.

Air Mass is the primary consideration in many drying,combustion process, and ventilating applications. Theseapplications generally require a certain amount of air mass,usually measured in pounds of air, to support the application.Because fan manufacturers publish fan capacities in actual cubicfeet per minute (ACFM), the mass of air required must beconverted from standard cubic feet per minute (SCFM) toACFM.

The velocity through a system can be determined once theACFM is known. The relationship between velocity and airflowis defined by the equation:

Q = VAwhere: Q = ACFM

V = velocity in lineal feet per minuteA = cross-sectional area in square feet

To determine the airflow requirement, the cross-sectional area ismultiplied by the required velocity.

System design is really a matter of defining the required work interms of volume or velocity and then sizing and selecting thenecessary system components to accomplish that work. Ofcourse, this must be done within the economic and spaceconstraints of the installation.

DETERMINING SYSTEM RESISTANCE

System resistance is the sum of the resistance through eachcomponent within the system. The system depicted in Figure 1may appear complex, but dealing with each componentseparately provides an orderly process for determining theoverall resistance.

HOOD LOSS

To determine hood or entrance losses, resistance calculationsmust be made for both the acceleration loss and the entry loss.Since the air or atmosphere surrounding the hood must beaccelerated from a state of rest, energy will be required to setthe air in motion. This energy is equal to the velocity pressure atthe entrance to the duct. Assuming the hood in this exampleempties into a 7" diameter duct, the required 1165 ACFMresults in a velocity of 4363 FPM:

V = Q ÷ A where: Q = 1165 CFM

Figure 1 – Typical System

(3.5 in. radius)2 x 3.1416A = 144 in.2 /ft.2 = .267 ft.2

therefore: V = 1165 CFM ÷ .267 ft.2 = 4363 FPMThe velocity pressure (VP) at 4363 FPM is calculated by:

Velocity 2VP = ( 4005 )

4363 2therefore: Acceleration Loss = ( 4005 ) = 1.19" W.G.

The same result can be obtained by interpolating from the datain Figure 2.The entry loss of a hood is a function of its efficiency. Theefficiencies of several common entry conditions are shown inFigure 3. The relative efficiencies are expressed as losses inpercentage of the duct velocity pressure. Consequently, thelowest percentage is actually the most efficient.

OutletVelocity

VelocityPressure

OutletVelocity

VelocityPressure

OutletVelocity

VelocityPressure

800 .040 2800 .489 4600 1.321000 .063 3000 .560 4800 1.441200 .090 3200 .638 5000 1.561400 .122 3400 .721 5200 1.691600 .160 3600 .808 5400 1.821800 .202 3800 .900 5600 1.952000 .250 4000 .998 5800 2.102200 .302 4200 1.10 6000 2.242400 .360 4400 1.21 6200 2.402600 .422

Figure 2 – Acceleration Loss

Figure 3 - Entrance Loss Percentage

The hood in this example is most similar to item 2 in Figure 3.Therefore, the entry loss from atmosphere into the hood is .90times the entering air velocity pressure at 1000 feet per minuteor:

1000 2Entry Loss = .90 x (4005 ) = .06" W.G.

This loss could have been reduced to .5 VP by simply adding aflange to the bottom edge of the hood as indicated by item 3 inFigure 3.

The total hood loss in the example is the acceleration loss addedto the entry loss:

Hood loss = .06" + 1.19" = 1.25" W.G.

PRIMARY BRANCH

The duct loss from the hood to the branch junction can bedetermined by using the equivalent length method. This run ofduct includes 62' of 7" diameter duct and one 4 piece 90° elbowof R/D = 2. According to Figure 4, the elbow has a loss equal to12 diameters of 7" duct, or 7'. Thus, the total equivalent lengthof straight duct is 69'.

Figure 4 - Loss in 90° elbows of round cross-section

Chart I on page 4 indicates a 4.0" loss for every 100' of 7"diameter duct handling 1165 CFM. The loss for this run can bedetermined as:

69Duct Loss = ( 100 ) x 4.0 = 2.76" W.G.

Therefore, the total resistance of the hood branch to the junctionis:Branch Loss = 1.25" + 2.76" = 4.01" W.G.

SECONDARY BRANCH

A secondary branch is calculated in the same manner as themain branch. For example, a grinder hood handling 880 CFMthrough a 6" pipe results in a velocity of 4500 FPM, which has a1.26" VP.

According to item 1 in Figure 3, a grinder hood has a .6 VP loss,so the total hood loss will be:

Hood Loss = 1.26" + (.60 x 1.26") = 2.02" W.G.

Page 2

The duct branch from the grinder hood to the junction consistsof 27' of 6" pipe and (2) 4 piece 90° elbows of R/D = 2. With anequivalent length of 39' (27' + 6' + 6') the duct loss for this runis:

39Duct Loss = ( 100 ) x 5.2 = 2.03" W.G.

See Chart I on page 4, which indicates a 5.2" loss for every 100'of 6" diameter duct handling 880 CFM.

The total resistance of the grinder branch to the junction is:

Branch Loss = 2.02" + 2.03" = 4.05" W.G.

Note that the resistance in both branches is nearly equal. This isbecause the pressures in converging branches must be equalduring operation or the system will automatically equalize byadjusting the flow different than the design point. If thevariation in resistance between any two converging branchesexceeds 5%, further design is required to balance the loss inboth branches. Where necessary, balancing can be accomplishedby altering duct lengths, duct diameters, or air volumes.

MAIN DUCT

The main duct resistance calculations begin with the selection ofthe appropriate duct diameter. Assuming a minimum conveyingvelocity of 4500 FPM and an airflow requirement of 2045ACFM (880 + 1165) in the main, a 9" diameter duct will sufficewith a resulting velocity of 4630 FPM.

The junction itself represents a loss due to the mixing effect ofthe converging branches. The ratio of the CFM in the branch(1165 ÷ 880 = 1.3) can be used to determine the loss in percentof VP in the main. Interpolating from the data in Figure 5 resultsin:

4630 2Junction Loss = .19 ( 4005 ) = .25" W.G.

LOSS IN MAIN AT JUNCTION WITH BRANCH. (BASEDON 45° TEE & EQUAL MAIN & BRANCH VELOCITIES.)

CFM in UpstreamMain ÷ CFM in Branch

Loss in Mainin % of Main V.P.

1 .202 .173 .154 .145 .136 .127 .118 .109 .10

10 .10CORRECTION FACTORS FOR OTHER THAN 45° TEE.

Tee Angle 45° Loss X Factor0 0

15 0.130 0.545 1.060 1.775 2.590 3.4

Figure 5

Chart II on page 4 indicates a resistance of 3.3" for every 100' of9" diameter duct handling 2045 CFM. According to Figure 4 thetwo elbows are equal to another 18' of duct, so the totalequivalent length is 68' between the junction and the fan.

39Duct Loss = ( 100 ) x 3.3 = 2.24" W.G.

Note that all the losses to this point, up to the fan inlet, areexpressed as negative pressure. Also that only the branch withthe greatest loss is used in determining the total.

Therefore:

SP inlet = (-4.05") + (-.25") + (-2.24") = -6.54"W.G.

Assuming the same size duct from the fan to the collector, the30’ of duct and the one elbow will have a loss equivalent to thefollowing:

39Duct Loss = ( 100 ) x 3.3 + 1.29" W.G.

The pressure drop across the dust collector, like coils or filters,must be obtained from the manufacturer of the device.Assuming a 2.0" loss for this example, the resistance at the fanoutlet is:

SP outlet = 1.29" + 2.0" = 3.29" W.G.

FAN SELECTION

At this point enough information is known about the system tobegin fan selection. Because fans are rated independent of asystem, their ratings include one VP to account for acceleration.Since the system resistance calculations also consideracceleration, fan static pressure can be accurately determined asfollows:

Fan SP = SP outlet - SP inlet - VP inlet

In this example with 4630 FPM at the fan inlet, and a 1.33" VP

Fan SP = 3.29" - (-6.54") - 1.33" = 8.5" W.G.

For this example, a fan should be selected for 2045 ACFM at8.5" SP and have an outlet velocity of at least 4500 FPM toprevent material settling. This presumes a standard airstreamdensity of .075 lbs./ft.3 If the density were other than standard,the system-resistance calculations would have been the same butthe resulting fan SP would have been corrected. Refer toEngineering Letter 4 for density correction procedures.

This example also assumes that the fan inlet and outletconnections are aerodynamically designed. Fans are sensitive toabrupt changes in airflow directly adjacent to the fan inlet oroutlet. The effects of abrupt changes and other “system effect”problems are discussed in Engineering Letter 5.

CONCLUSION

It is the responsibility of the system designer to ensure that thereare adequate air flows and velocities in the system and that theselection of duct components and fan equipment has beenoptimized. While computer programs do the bulk of systemcalculations today, this Engineering Letter should help toprovide a common set of methods and terminology to assist inthat effort.

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ENGINEERING LETTER 2The New York Blower Company ●7660 Quincy Street, Willowbrook, Illinois 60521-5530

F A N L A W S A N D S Y S T E M C U R V E S

INTRODUCTION

The purpose of this Engineering Letter is to explain the basisand application of the rules used to predict fan performance in agiven system. With a basic understanding of these rules, theperformance of a fan can be quickly calculated for variousconditions.

SYSTEM REQUIREMENTS

The three fundamental rules governing fan performance arecommonly called the “fan laws.” These rules are only validwithin a fixed system with no change in the aerodynamics orairflow characteristics of the system. For the purpose of thisdiscussion, a system is the combination of ductwork, hoods,filters, grills, collectors, etc., through which air is distributed.Therefore, these rules can also be referred to as “system laws.”

VOLUME AND PRESSURE

The motion of any mass causes friction with its surroundings.The movement of air through a system causes friction betweenthe air molecules and their surroundings (duct walls, filtermedia, etc.) and any other air molecules. Energy is required toovercome this friction, or resistance. The faster the air movesthe greater the resistance to flow and the more energy is requiredto push or pull the air through the system.

This energy is stated in terms of pressure. The portion of thepressure that results in air velocity is described as velocitypressure (VP). The portion necessary to overcome friction in theair and in the system is described as static pressure (SP). Thesum of the two is described as total pressure (TP).

The law of physics, for motion, is expressed algebraically as:

V = √2gh or V2 = 2gh

where V = velocity of flowg = force of gravityh = pressure causing flow

As can be seen from the equation, the pressure necessary tocause flow is proportional to the square of the velocity. In asystem, this means that SP will vary as the square of the changein velocity or volume expressed in cubic feet per minute (CFM).This makes it possible to predict all possible combinations of SPat the corresponding CFM given any one such calculatedrelationship of SP and CFM for a fixed system.For example, a system is calculated to require a static pressureequal to 2" water gauge at an airflow rate of 1000 CFM. If it isdesired to increase the flow to 1500 CFM without any physicalchange in the system, the required SP would be:

(1500 ÷ 1000)2 x 2” = 4.5” SP

CFM new SP new( CFM old ) 2= SP old

Figure 1 - System Curve

The same calculation using any number of varying CFM ratingswould result in a plotted curve as shown in Figure 1.Regardless of fan type, fan size, or volume of flow through asystem, the relationship of CFM to SP will not change unless thesystem itself is altered in some way. SP always varies as thesquare of the change in CFM. The only exception to this rule isfound in a laminar flow characteristic where VP is of far greaterimportance than SP. Such circumstances are not typical of fansystems.

FAN LAWS

In air movement systems, it is the fan wheel that does the work.In a sense, the fan wheel acts like a shovel. As it revolves, itdischarges the same volume of air with each revolution.Working within a fixed system, a fan will discharge the samevolume of air regardless of air density, (disregarding the effectsof compression at high pressures).If the fan RPM is increased, the fan will discharge a greatervolume of air in exact proportion to the change in speed. This isthe first “fan law.”

1. CFM varies in direct proportion to change in RPM

RPM (new)CFM (new) =RPM (old)

x CFM (old)

Figure 2 - A fan wheel is a constant volume device.

As shown earlier, in a system, the SP varies as the square ofthe change in CFM. Since CFM varies directly with RPM,RPM can be substituted for CFM in the system equation.Therefore, SP varies as the square of the change in RPM. Thisis the second “fan law.”

2. SP varies in proportion to the change in (RPM)2

RPM (new)SP (new) = ( RPM (old) ) 2x SP (old)

The efficiency of a fan is a function of its aerodynamic designand point of operation on its SP/CFM curve (see Figure 3). Asthe fan speed changes, this relative point of operation remainsunchanged as long as the system remains unchanged. Thus, thefan brake horsepower varies proportionally as the cube of thechange in RPM. This is the third “fan law.”

3. BHP varies in proportion to the change in (RPM)3

RPM (new)BHP (new) = ( RPM (old) ) 3x BHP (old)

It is important to remember that each of these “fan law”relationships takes place simultaneously and cannot beconsidered independently.

FAN CURVE AND SYSTEM CURVE

As stated previously, a system curve can be plotted to show allpossible combinations of SP and CFM for a given fixedsystem. Any fan used on that system must operate somewhereon that system curve.

Fan performance is determined by laboratory testing and ispresented graphically in the form of fan curves. Unless it isphysically altered in some way, a fan must operate somewhereon its SP/CFM curve. The relative shape of that curve will notchange, regardless of fan speed.

Because the fan and system can each only operate somewhereon their own respective curves, a fan used on a fixed systemcan only have one point of operation. The point of operation, asshown in Figure 3, is the intersection of the system curve andthe fan SP CFM curve.

If the fan speed is increased or decreased, the point ofoperation will move up or down the existing system curve.This is shown in Figure 4.The following are examples of how the fan curve can be used tocalculate changes to flow and pressure requirements.

Example 1: A fan has been selected to deliver 35,530 CFM at8" SP. The fan runs at 1230 RPM and requires 61.0 BHP.

After installation, it is desired to increase the output 20%. Atwhat RPM must the fan run? What SP will be developed?What BHP is required?

1. CFM varies as RPM(1230) (1.20) = 1476 RPM

2. SP varies as (RPM)2

(1476/1230)2 (8) = 11.52" SP

3. BHP varies as (RPM)3

(1476/1230)3 (61.0) = 105.4 BHP

Example 2: A fan was originally installed to deliver 10,300 CFM at2 1/4" SP and to run at 877 RPM, requiring 5.20 BHP.

After installation, it is found that the system only delivers9,150 CFM at 2 1/2" SP and uses 4.70 BHP. This indicates theoriginal calculations were in error, or that the system was notinstalled according to plan. What fan RPM and BHP will benecessary to develop the desired 10,300 CFM? What SPshould have been figured?

1. CFM varies as RPM(10,300/9,150) (877) = 987 RPM


(987/877)2 (2.50) = 3.17" SP


(987/877)3 (4.70) = 6.70 BHP

CONCLUSION

Use of the “fan laws” is based on a fixed system and a non-modified fan. Adding or deleting system components such asdampers, or incurring density changes, will create completelynew system curves. Changing fan accessories such as inletboxes, evases, or inlet dampers will alter the fan’s performancecurve from standard. These variables must be considered beforethe fan laws can be applied.

During the process of system design, the fan laws can behelpful in determining alternate performance criteria or indeveloping a minimum/maximum range. If “safety factors” areapplied to system calculations, it should be recognized that a10% factor on volume will result in an increase in horsepowerof 33% according to the third fan law. An evaluation should bemade weighing the necessity of the safety factor versus thecost penalty incurred.

Form 607 GAW

Figure 3

Figure 4

ENGINEERING LETTER 3The New York Blower Company ●7660 Quincy Street, Willowbrook, Illinois 60521 -5530

U N D E R S T A N D I N G F A N P E R F O R M A N C E C U R V E SINTRODUCTION

One of the most important documents customers request fromfan manufacturers is performance curves. In addition tographically depicting the basic fan performance data of CFM,RPM, and SP (on the static pressue curve) and BHP (on thebrake horsepower curve), these curves also illustrate theperformance characteristics of various fan types, like areas ofinstability, or the rate of change between flow and pressure.With some basic knowledge of performance curves, decisionscan be made concerning fan selection, fan and system alterations,or the advisability of using a fan in a modulating system, forexample.

Except for very large fans, performance curve information isgenerated by connecting the fan to a laboratory test chamber.Very specific test procedures are followed as prescribed in theAir Movement and Control Association’s Standard 210 to assureuniform and accurate readings. Data points are collected at agiven RPM while the flow is slowly modulated from full closedto full open. The information gathered is then used to developcomputer selection programs and published capacity tables foruse by system designers and end users.

STATIC PRESSURE CURVE

The static pressure curve provides the basis for all flow andpressure calculations. This curve is constructed by plotting aseries of static pressure points versus specific flow rates at agiven test speed. While the static pressure curve depicts a fan’sperformance at a given speed, it can be used to determine thefan’s pressure capability at any volume.

In addition, it is also possible to approximate the fan’s performanceat other speeds by applying the following fan laws:

1. CFM varies as RPM



To locate a fan’s point of operation, first locate the requiredstatic pressure on the SP scale at the left of the curve. Then drawa horizontal line to the right, to the point of intersection with theSP curve. Next, draw a vertical line from the point of operationto the CFM scale on the bottom to determine the fan’s flowcapability for that SP at the given speed.

As shown in Figure 1, the performance for this fan is 8750 CFMand 12" SP at 1750 RPM.

Figure 1 - Static Pressure Curve

Assuming this same fan was intended to operate at 1200 RPM, thefan laws can be applied to obtain performance at this lower speed.

1. CFM varies as RPM

CFM (new) RPM (new)CFM (old) = RPM (old)

Therefore:

1200CFM (new) = 1750 (8750) = 6000 CFM


SP (new) RPM (new)SP (old)

= ( RPM (old) ) 2

Therefore:

1200SP (new) = ( 1750 ) 2 (12) = 5.6” SP

BRAKE HORSEPOWER CURVE

Once the CFM and SP have been determined, a BHP rating canbe established. An accurate BHP rating is necessary to properlysize the motor or to determine the operating efficiency of one fanas compared to another. Performance curves contain a BHPcurve from which the BHP rating can be determined for specificcapacities. To determine BHP at a specific point of operation, ahorizontal line is drawn to the right from the point ofintersection of the vertical CFM line and the BHP curve.

Figure 2 - Performance Curve

As shown in Figure 2, the fan operating at 8750 CFM and 12" SPat 1750 RPM is rated at 30 BHP. By employing the third fan law,the BHP rating can be determined for operation at 1200 RPM.


BHP (new) RPM (new)BHP (old)

= ( RPM (old) ) 3

Therefore:

1200 3BHP (new) = ( 1750 ) (30) = 9.67 BHP

SYSTEM LINES

Since fans are tested and rated independently from any type ofsystem, a means of determining the fan’s capabilities within agiven system must be provided. The fan laws apply equally toany system; therefore, CFM and SP variations within the systemare predictable. This enables system lines to be superimposed onperformance curves to simplify performance calculations. Thesystem line is nothing more than the sum of all possible CFMand SP combinations within the given system. Any combinationof fan and system must operate somewhere along that system line.

Because a fan must operate somewhere along its SP curve andsince the system has a known system line, their intersection isthe point of operation. If the fan speed is changed, the point ofoperation must move up or down the system line. If the system ischanged in some way, the point of operation must move up ordown the SP curve. In practice, these principles can be used tocheck the accuracy of fan performance and system design.

USING PERFORMANCE CURVES

Figure 3 illustrates the point of operation of a fan selected for8750 CFM and 12" SP operating at 1750 RPM. Included inFigure 3 are a number of different system lines. If the systemdoes not operate properly upon start-up, measurements can betaken and compared against the available performance curve.

Figure 3 - Performance Curve with System Lines

Let’s assume that a tachometer reading indicates the fan isrunning at 1200 RPM instead of 1750 RPM, because of mistakesin motor speed or drive selection, and an airflow check indicatesonly 6000 CFM. These readings confirm that the system wascalculated correctly and that the fan speed must be corrected to1750 RPM as originally specified to achieve the desired 8750CFM. If the tachometer reading indicates the proper speed butthe airflow reading is down, additional system resistance beyondthat originally calculated is indicated. This additional resistancecould be caused by partially closed louvers/dampers, changes induct sizing from the original design, system effect losses, or justan error in the system-resistance calculations. The deficiency canusually be corrected by either altering the system or increasingthe fan speed.

Often, performance curves for one speed must be used todetermine the performance of a fan for use on systems requiringmore air or higher pressures. A performance curve such asFigure 4 can be used to determine fan performance beyond the SPscale shown by using the fan laws to obtain a reference point ofoperation on the system line. This can be accomplished byapplying some suitable factor to the required CFM and thesquare of that factor to the required SP.

For example, the performance curve shown in Figure 4 can beused to determine fan performance requirements for a systemcalculated at 15,000 CFM and 23.5" SP, even though that pointis beyond the curve. By determining a suitable referencecapacity using the fan laws, that falls within the curve data, fanperformance requirements can be obtained at the curve speedand then projected up to the system requirements using the fanlaws once again.

The required 15,000 CFM and 23.5" SP is on the same systemline as 10,000 CFM at 10.4" SP which intersects the fan’s SPcurve drawn for 1750 RPM and has a corresponding BHP of33.0 at 1750 RPM. Therefore:

15000RPM (new) = 10000 (1750) = 2625 RPM

15000BHP (new) = ( 10000 ) 3 (33.0) = 111 BHP

Page 2

FAN PERFORMANCE CHARACTERISTICS

The performance characteristics of a fan can be determinedfrom the performance curve at a glance. These characteristicsinclude such things as stability, increasing or non-overloadingBHP, and acceptable points of operation.

Fan performance is based on certain flow characteristics as theair passes over the fan wheel blades. These flow characteristicsare different for each generic fan or wheel type, (i.e. radial,forward-curved, backwardly-inclined, radial-tip, and axial).Thus, the performance characteristics will be different for each ofthese general fan types. Further, these performance characteristicsmay vary from one manufacturer to the next depending uponthe particular design. The characteristics described in this letterare based on nyb fan equipment.

The performance curves presented in Figures 1 through 4 aretypical of fans with radial-blade wheels. The SP curve issmooth and stable from wide open to closed off. The BHPcurve clearly indicates that the BHP increases steadily with thevolume of air being handled as shown in Figure 4.

Fans with forward-curved wheels, such as shown in Figure 5,also have a BHP curve that increases with the volume of airbeing handled. The SP curve differs significantly from theradial since it exhibits a sharp “dip” to the left of the staticpressure peak. This sharp dip (shaded area) indicatesunpredictable flow characteristics. Though not technicallyaccurate, this region is often referred to as the the “stall” region.It indicates that at these combinations of pressure and relativelylow volumes, the airflow characteristics across the wheel bladeschange or break away so that the fan performance point is nolonger stable. Any fan with this characteristic SP curve shouldnot be selected for operation in the unstable area.

As shown in Figure 6, the SP curve for a backwardly-inclinedfan has a sharp dip to the left of the static pressure peak. Thisindicates an area of instability. However, the backwardly-inclined SP curve is generally steeper than that of the forward-curved wheel indicating its desirability for use in higherpressure systems. Therefore, variations in system resistancewill result in smaller changes in volume for the BI Fan whencompared to the FC Fan.

Even though New York Blower centrifugal fans withAcoustaFoil® wheels are stable in the area left of the peak, themajority of fans with backwardly-inclined wheels exhibit an SPcurve similar in appearance to that of the forward-curved fan. TheSP curve shown (in Figure 7) for fans using AcoustaFoil (air-foil, backwardly-inclined) wheels exhibits a much smootherdepression to the left of the static pressure peak. This indicatesthat the overall fan design is such that internal flowcharacteristics remain desirable or predictable even in theregion left of peak and that performance in this region is stable.

AcoustaFoil ® is a trademark of The New York Blower Company.

StaticPressure

CFM in 1,000’s

Figure 4 – Typical Radial-Blade Fan Performance Curve

Figure 5 – Typical FC Fan Performance Curve

CFM

CFM

Figure 6 – Typical BI Fan Performance Curve

StaticPressure

StaticPressure

BrakeHorsepower

BrakeHorsepower

BrakeHorsepower

The BHP curve for all backwardly-inclined fans is the majordifference between them and all other fan types. As shown inFigures 6 and 7, the BHP curve for backwardly-inclined fansreaches a peak and then drops off as the fan’s volume increases.With this “non-overloading” BHP characteristic, it is possible toestablish a maximum BHP for a given fan speed and select amotor that can not be overloaded despite any changes or errors insystem design. Because BHP varies as (RPM)3 , this non-overloading characteristic does not apply to increases in fanspeed, but it is very useful for motor protection against errors orchanges in system calculations and installation.

Figures 5 and 6 indicate certain unacceptable selection areas onthe SP curve. Although stability or performance may not be aproblem, a point of operation down to the far right on the SPcurve should be avoided. Selecting a fan that operates far downto the right, eliminates the flexibility to compensate for futuresystem changes. Also, the fan is less efficient in this area ascompared to a larger fan operating at a slower speed. Figure 7shows the best selection area on the SP curve and the area in whichthe majority of capacity tables are published.

As is evident in Figure 8, the radial-tip fan design combines thebackwardly-inclined SP curve characteristics with the radial fan’sBHP curve. The radial tip is often more efficient than radial fansand typically best applied in high-pressure applications. As a resultof its efficiency and dust-handling capabilities, the radial-tip fancan also be applied to lower pressure material conveyingsystems.

The term axial fan is used to describe various propeller,vaneaxial, tubeaxial, and duct fans. The performance curves ofthese fans are characterized by the ability to deliver largevolumes of air in relatively low pressure applications. As can beseen in Figure 9, the axial flow fan is distinguished by adrooping BHP curve that has maximum horsepower at no flow orclosed-off conditions. The axial fan SP curve exhibits an area ofextreme instability to the left of the “hump” in the middle of thecurve. Depending upon the severity, axial fans are normally onlyselected to the right of this area.

CONCLUSION

A good working knowledge of performance curves is necessaryto understand the performance characteristics and capabilities ofdifferent fan equipment. Use of performance curves in theselection of fan types and sizing will assure stable and efficientoperation as well as future flexibility.

Form 607 GAW

CFM

Figure 7 – Typical AcoustaFoil Fan Performance Curve

CFM

Figure 8 – Typical Radial-Tip Fan Performance Curve

CFMFigure 9 – Typical Axial Fan Performance Curve


TEMPERATURE AND ALTITUDE AFFECT FAN SELECTIONINTRODUCTION

Fan performance changes with the density of the gas beinghandled. Therefore, all fans are cataloged at a standard conditiondefined as: 70°F. air, at sea level, with a gas density of .075lb./ft.3 at a barometric pressure of 29.92" Hg. At any othercondition, the fan’s horsepower requirement and its ability todevelop pressure will vary. Therefore, when the density of thegas stream is other than the standard .075 lb./ft.3, correctionfactors must be applied to the catalog ratings in order to selectthe correct fan, motor, and drive.

In addition, the maximum safe speed of a wheel, shaft, orbearing can change due to an alloy becoming too brittle or toopliable at temperatures other than 70°F. Temperature deratefactors must be applied to the fan’s catalog maximum safespeed to ensure against overspeed situations.

HOW TO CALCULATE ACTUAL FAN PERFORMANCEAT OTHER THAN 70 DEGREES FAHRENHEIT

As illustrated in Figure 1, a fan wheel is similar to a shovel. Ina given system, it will move the same volume of air regardlessof the air’s weight. If a fan moves 1000 CFM at 70°F., it willalso move 1000 CFM at 600°F.

However, air at 600°F. weighs half as much as it does at 70°F.Therefore, the fan requires just half the horsepower. (SeeFigure 2.) Likewise, since the air weighs half as much, it willcreate only half the static and velocity pressures. The reductionin static pressure is proportional to the reduction in horsepower,thus the overall fan efficiency will remain unchanged.

CFM x Total PressureTotal Efficiency = 6356 x Brake Horsepower

Example 1. A fan handling standard density, 70°F. air, delivers12,400 CFM against 6" SP (static pressure) requiring 14.6BHP (brake horsepower). If the system and fan RPM are notchanged, but the inlet airstream temperature is increased to600°F., how will the fan perform?

The fan will still deliver 12,400 CFM, but since the air at600°F. weighs half as much as the air at 70°F., static pressureand horsepower will be cut in half. The fan will generate only 3"SP and require only 7.3 BHP.

A typical fan specification based on hot operating conditions isillustrated in Example 2.

Example 2. Required: 11,000 CFM and 6" SP at 600°F. (Thismeans the actual, measurable static pressure of the fan at600°F. will be 6 inches of water.)

The fan’s catalog performance tables are based on 70°F. air at.075 density. The specified SP must be corrected by the ratioof the standard density to operating density. Since densities areinversely proportional to absolute temperature (degrees F. +460):

460 + 600 10606” ( 460 + 70 ) = 6” ( 530 ) = 12”

The fan must be selected from the rating tables for 11,000CFM at 12" SP. The BHP obtained from the table should bemultiplied by the ratio of operating density to standard densityin order to obtain the BHP at 600°F. If the rating table showed30.0 BHP, the operating BHP would be 30.0 (530/1060) = 15.0BHP.

In most “hot” systems, the fan is required to handle cold airuntil the system reaches temperature. A good example is inoven exhaust systems.

Figure 1 - A fan wheel is like a shovel.

Figure 2 - With hot gas, there is less weight to shovel.

If Example 2 were such a case, the fan would require 30.0 BHPwhen operating at 70°F., and 15.0 BHP when the oven hadwarmed to 600°F. Very often a damper is furnished with the fanso that, during the warming-up period, the fan can be damperedto reduce the horsepower. Without the damper, a 30 HP motorwould be needed.

Confusion can be avoided if the SP is specified at the temperatureit was calculated. In Example 2, the specifications should readeither:

11,000 CFM and 6" SP at 600°F., or11,000 CFM for operation at 600°F. and 12" SP at 70°F.

Table 1 gives correction factors used to convert from a non-standard density to a standard density of 70°F. air. These factorsare merely the ratios of absolute temperatures. Multiply theactual static pressure by the specific temperature/altitude factorso standard catalog rating tables can be used. Divide the brakehorsepower from the catalog rating table by thetemperature/altitude factor to get BHP at conditions.

Table 1 - Corrections for Temperature

AirTemperature

°F.Factor

AirTemperature

°F.Factor

-50 0.77 275 1.39-25 0.82 300 1.43

0 0.87 325 1.48+20 0.91 350 1.53

40 0.94 375 1.5860 0.98 400 1.6270 1.00 450 1.7280 1.02 500 1.81

100 1.06 550 1.91120 1.09 600 2.00140 1.13 650 2.09160 1.17 700 2.19180 1.21 750 2.28200 1.25 800 2.38225 1.29 900 2.56250 1.34 1000 2.76

Table 2 - Corrections for Altitude

AltitudeFeet AboveSea Level

FactorAltitude

Feet AboveSea Level

Factor

0 1.00 5000 1.20500 1.02 5500 1.22

1000 1.04 6000 1.251500 1.06 6500 1.272000 1.08 7000 1.302500 1.10 7500 1.323000 1.12 8000 1.353500 1.14 8500 1.374000 1.16 9000 1.404500 1.18 10000 1.45

HOW TO CALCULATE ACTUAL FAN PERFORMANCEAT OTHER THAN SEA LEVEL

A fan operating at an altitude above sea level is similar to a fanoperating at air temperatures higher than 70°F.; it handles airless dense than standard. Table 2 gives the ratio of standard airdensity at sea level to densities of 70°F. air at other altitudes.

Example 3. Required: 5800 CFM at 6" SP at 5000 ft. altitude.70°F. air at sea level weighs 1.20 times as much as 70°F. air at5000 Ft. Therefore, at sea level, the SP is 1.2 x 6 = 7.20" SP.The fan would need to be selected for 5800 CFM at 7.2" SP at70°F. .075 density.

When both heat and altitude are combined, the density of the airis modified by each, independently, so that the correction factorscan be multiplied together.

Example 4. Required: 5800 CFM at 6" SP at 5000 ft. altitude at600°F. Air at 70°F. at sea level weighs 2.00 x 1.20 = 2.40 timesas much as 600°F. air at 5000 ft. altitude. At sea level and 70°F.,SP = 2.40 x 6 = 14.4" SP. Select a fan for 5800 CFM at 14.4"SP. Divide the brake horsepower in the rating table by 2.40 toobtain horsepower at 600°F. and 5000 ft. If the fan is to startcold, it will still be at 5000 ft. altitude. Therefore, to get the“cold” horsepower requirement, divide by 1.20, the altitudefactor only.

DENSITY CHANGES FROM OTHER THAN HEAT ANDALTITUDEFan densities may vary from standard for other reasons than heatand altitude. Moisture, gas, or mixtures of gases (other than air)are a few possibilities. In these cases, it is necessary to obtainthe actual density of the airstream gas by some other referencematerial. A similar factor, as shown in Table 1, is then createdusing the standard density of air .075 lb. per cubic foot divided bythe new density.

.075 lb./ft.3

Factor = special gas density

ACFM-SCFM DEFINITION

The terms ACFM and SCFM are often used in design work andcannot be used interchangeably.

SCFM is Standard Cubic Feet per Minute corrected to standarddensity conditions. To determine the SCFM of the volume usedin Example 2, which was 11,000 CFM at 600°F., we wouldmultiply the CFM by the density ratios.

.03711000 x .075 = 5500 SCFM

This indicates that if the weight of air at 600°F. were corrected tostandard conditions its volume would be reduced to 5500 CFM

ACFM stands for Actual Cubic Feet per Minute. It is the volumeof gas flowing through a system and is not dependent upondensity.

The terms ACFM and SCFM are often used in system designwork where both quantities need to be known. It should beremembered, however, that since a fan handles the same volumeof air at any density, ACFM should be used when specifyingand selecting a fan.

FAN SAFE SPEED AND TEMPERATURE

Whenever a fan is used to move air at temperatures substantiallyabove or below 70°F., care must be taken to ensure that the safespeeds of wheel and shaft are not exceeded, and that bearingtemperature and lubrication are satisfactory.

The maximum safe speed of a particular fan must be determinedby calculations or actual tests. Safe speed depends entirely uponthe wheel and shaft assembly’s ability to withstand thecentrifugal forces created by its own weight. Highertemperatures can affect the wheel and shaft assembly’s ability towithstand these forces and therefore must be considered.

Most metals become weaker at higher temperatures. Thisweakness is measurable in terms of yield and creep strength. Itcan be translated into formulas that accurately determine the safespeed of a wheel and shaft assembly in relation to its testedmaximum speed at standard conditions. Manufacturers providesafe speed reductions in their catalogs based on the alloy thatwas used to manufacture the wheel and/or shaft.

Some metals withstand heat better than others. Certain grades ofstainless steel can be substituted to increase temperature limits.On the other hand, fan wheels constructed of aluminum shouldnever be operated above 200°F.

For information regarding fiberglass reinforced plastic fanequipment, consult the appropriate product bulletin.

Table 3 gives an indication of the speed derate factors for severaldifferent alloys. These are listed for reference purposes only.For a specific fan, consult the appropriate product bulletin.

Table 3 - RPM Derate Factors By Material

Stainless SteelTemperature°F.

Mild Steel Aluminum304L 316L 347

70 1.0 1.0 1.0 1.0 1.0200 .97 .97 .88 .95 .95300 .95 -- .82 .92 .93400 .94 -- .78 .89 .90500 .93 -- .75 .86 .90600 .92 -- .73 .84 .90800 .80 -- -- .79 .86

1000 -- -- -- .75 .83The limiting temperature on any fan is the highest temperaturethat any component of the fan assembly will reach during anyoperating cycle. A fan in a process oven application may handleair several hundred degrees above the highest temperature theoven reaches, especially during start-up. On such applications, atemperature indicator should be located in the fan inlet tocontrol the heat source and to keep the fan within its maximumsafe temperature. This is particularly true where burners arelocated on the inlet side of the fan. In all cases, the fan shouldremain in operation until the air is cooled to 180°F. or less toprevent “heat soaking” of the fan shaft which could cause sagging.

Bearings must be kept cool; otherwise standard lubricants losetheir effectiveness and bearing failures are likely. For axial fans,where the bearings are located in the airstream, care must betaken to ensure proper lubrication. Special fan and bearingdesigns, as well as high temperature lubricants, are available toextend the operating range to higher temperatures.

Arrangement 4 centrifugal fans, where the fan wheel is mountedon the motor shaft, should not be used above 180°F., unlessspecial provisions are made (i.e., a shaft cooler or heat shield) tokeep heat radiated from the housing from increasing motorbearing and winding temperatures.

When fan bearings are located outside of the airstream, as inArrangement 1, 8, and 9 centrifugal fans, higher airstreamtemperatures are possible. Table 4 lists some typical maximumrecommended operating temperatures for fans using ball orroller bearings.

A conventional fan using standard bearings and standard lubricantcan normally be operated to a maximum of approximately300°F. With the addition of a shaft cooler (Figure 3), thistemperature limitation can be extended to 650°F. The shaftcooler has the effect of absorbing and dissipating heat from theshaft while circulating air over the inboard bearing.

Table 4 - Maximum Fan Inlet Temperatures

Arrangement 1 and 8 (Overhung Wheel)

Standard ConstructionWith Shaft CoolerWith Shaft Cooler and Heat GapWith Shaft Cooler, Heat Gap,

Stainless Wheel, and Alloy Shaft

300°F.650°F.800°F.

1000°F.

Arrangement 3 (Wheel Suspended Between Bearings)Standard Construction 200°F.

Arrangement 4 (Wheel on Motor Shaft)

Standard Construction 180°F.

Enclosed Bearing Fans (Axial Fans)Arrangement 4 105°F.Arrangement 9 120°F.

With Special V-Belts with 2.0 S.F. 200°F.Arrangement 9 Duct Fan

With Heat-Fan Construction 600°F.Plenum Fans

Arrangement 3 105°F.Arrangement 4 105°F.

Figure 3 – Shaft Color

With the addition of a heat gap (Figure 4) the temperaturelimitation can be extended to 800°F. since the fan pedestal isisolated from the hot fan housing. For specific applications,consult the appropriate product bulletin. Also, recognize thatthese limitations apply only to bearings and that wheel andshaft limitations must be treated independently.

All of the foregoing is based on the use of standard lubricants.When high-temperature lubricants are required, the type oflubricant and the frequency of relubrication are normally muchmore critical.

When the fan shaft is heated to the point that it expands morethan the structure to which it is attached, one expansion bearingand one fixed bearing should be furnished. The fixed bearing islocated on the drive end of the fan while the floating bearing islocated next to the fan. This arrangement, however, is notcritical and may vary by manufacturer.

When the fan is handling air below 70°F., there is the possibilityof other problems. Below -30 to -50°F., ordinary steel is toobrittle. Aluminum wheels or wheels of steel containing at least5% nickel must be used, and shafts must be made of nickel-

bearing steel. In addition, lubricants become stiff, or even solid inthese low-temperature applications. Exact operating conditionsshould be given to the fan manufacturer to relay to the bearingsupplier for proper selection.

CALCULATING “HOT” RESISTANCE FOR SYSTEMS

Figure 5 shows a system that operates at the same temperaturethroughout. If the inlet temperature is known, the fan may beselected from the fan capacity tables and the rated horsepower andstatic pressure corrected by the temperature correction factorfrom Table 1. However, what happens to the system that the fanwas operating against? If a fixed system, which originally wascalculated for standard air, was subjected to the same temperatureincrease as the fan, then system static pressure will change and beidentical to the fan static pressure change. The result is that if afan and system operate together the flow will remain unchanged.(See Figure 6.) Unfortunately, this example assumes that theentire system is being subjected to the same temperature change,which is not always the case.

Figure 4 – “Heat Gap” between fan andbearing.

Figure 6 – Fan-system curve relationshipwith fan and system at the same temperature.

Figure 5 – A system with the same temperature throughout.

Figure 4 – “Heat Gap” betweenfan and bearing.

Figure 7 shows a system in which different temperatures areinvolved. The fan will not handle the same volume of air whenoperating hot as it does when cold. If the burner is on, the fanwill handle 1430 ACFM against an actual static pressure of 1.2inches. This is arrived at by adding the filter, burner, and nozzleresistance, neglecting for the sake of simplicity any externalresistance from additional ductwork. The fan would be selectedfrom the capacity tables on the basis of 1430 CFM at 1.72inches static pressure (300°F. correction factor times 1.2inches).

If the burner is turned off while the fan continues to operate atthe same RPM, it is necessary to determine the systemcharacteristic curve and plot its intersection with the fan todetermine how much air the fan would move and at what staticpressure. To accomplish this we must assume an arbitrarycapacity, such as 1000 CFM at 70°F. The filter louver resistancewould be the same, cold or hot, at .3 inches 70°F. The burnerresistance would remain unchanged with temperature since it must

be assumed that air expansion takes place after the highvelocity section of the burner. The nozzles will vary inresistance directly as the density changes and inversely as thesquare of the flow. The nozzle would then have a resistance coldat 1000 CFM of:

1000.5” x ( 1430 )2

x 1.43 = .35”

Summing these resistances yields the cold resistance at 1000CFM of 1 .05"SP. This new system point and correspondingcurve are then plotted against a fan curve at standard conditionssuch that the resulting intersection will be the final operatingpoint of the cold system. Using an actual fan as an example,the resulting flow would be 1220 CFM at 1.5 inches staticpressure. (See Figure 8.)

Figure 7 - A system with different temperatures.

Figure 8 - Fan-system curve relationship withfan at different temperatures.

FAN LOCATION IN HOT PROCESS SYSTEMS

Figure 9 shows how a fan may be located more economically inone part of a system, as contrasted to another. Suppose 10,000CFM is to be heated from 70°F. to 600°F. Obviously, the heaterwill require the same 3-inch pressure differential whether thefan is to push the air into, or pull the air out of, the heater.

A fan pushing air into the heater would be specified to handle10,000 CFM at 70°F. against 3 inches of static pressure at70°F. One possible selection is a fan with a 27-inch wheeldiameter, Class I design utilizing a 71 / 2 HP motor.

The alternative fan, pulling air from the heater, would bespecified to handle 20,000 ACFM at 600°F. against 3" SP at600°F. It would be selected from the capacity tables for 20,000CFM at 6" SP. One suitable choice is a fan with a 3 61 / 2-inchwheel diameter, Class II design utilizing a 15 HP motor. (Note:26 HP, from the tables, at 70°F., divided by temperaturecorrection factor, is 13 HP at 600°F.) This example illustrateswhy it is usually more economical to locate the fan at thecoolest part of the system. In this case, the “push” fan mightcost half as much as the “pull” fan.

Figure 9 - The importance of fan location.

Form 607 GAW


F A N P E R F O R M A N C E - T H E S Y S T E M E F F E C TINTRODUCTION

Fans are typically tested and rated in prescribed test configurationsdefined by the Air Movement and Control Association. This isdone to ensure standardized procedures and ratings so thatsystem designers can make realistic choices among variousmanufacturers. Beyond the routine system resistance calculations,the location of some common components and their proximity tothe fan inlet or outlet can create additional immeasurable lossescommonly called System Efect. These losses, if not eliminatedor minimized, will necessitate fan speed and horsepowerincreases to compensate for the performance deficiencies. ThisLetter will outline some of the common causes for thesedeficiencies and provide useful guidelines for more efficient andpredictable air-handling systems.

SYSTEM DESIGNThe term system refers to the path through which air is pushedand/or pulled. Since it can be any combination of ducts, coils,filters, etc., through which air flows, a system can range incomplexity. The system can be as simple as exhausting airthrough an opening in the wall of a building, or as involved as amulti-zoned system with varying flows and densities. Thecalculations for determining the performance requirements arediscussed in Engineering Letter 1. The effects of the systemdesign on the actual performance capability of a fan representseparate and equally important considerations.

In the typical process of system design, the performancerequirements are calculated and then used to select theappropriate fan. However, in many cases the effects of therelationship between the system components and the fan are notconsidered in the calculation or selection process. For example,the resistance of a given size elbow at a given flow can be easilydetermined using the equivalent length calculation method.However, if that elbow is located at the fan inlet or outlet, furtherimmeasurable losses will be imposed in addition to the simpleloss through the elbow itself. Most importantly, these lossescannot be measured or even detected with field instrumentsbecause they are, in fact, a destruction of the fan performancecharacteristics.

Standardized testing and rating methods for fans have beenestablished by the Air Movement and Control Association,(AMCA). The test methods are described in AMCA Standard210, titled Test Code for Air Moving Devices. Specifying fanequipment tested and rated in strict accordance with AMCAStandard 210 is the best way to ensure accurate fan performance.However, the system effects that alter or limit the ultimateperformance remain the most frequent causes of fieldperformance problems.

The four most common causes of system-induced performancedeficiencies:

1. Eccentric flow into the fan inlet.2. Spinning flow into the fan inlet.3. Improper ductwork at the fan outlet.4. Obstructions at the fan inlet or outlet.

ECCENTRIC FLOW

Fans perform correctly when air flows straight into the inlet. Airshould be drawn into the fan inlet with an evenly distributedvelocity profile. As shown in Figure 1, this allows all portions ofthe fan wheel to handle an equal air load.

If the air is not drawn into the fan inlet evenly, performancedeficiencies will result from the combined effects of turbulenceand uneven air distribution. This is illustrated in Figure 2, wherean elbow is installed directly on the fan inlet.

Figure 1 – Even Air Loading

Figure 2 – Uneven Air Loading

When the system attempts to change the direction of flow, theair hugs the outside of the inlet elbow entering the fan. Thiscauses uneven, turbulent airflow into the fan. Another commoncause of non-uniform flow into the fan inlet is a poorly designedinlet box, such as the one shown in Figure 3. It is important toremember that air has mass.

SPINNING FLOW

Unintentionally spinning air into the fan inlet can have the sameeffect on performance as the intentional pre-spin produced by avortex-type inlet damper.

The direction air is flowing when it enters the fan wheel is veryimportant. In order to produce its rated capacity, the fan workson the air by changing its direction and accelerating its velocity. Ifthe air is spinning in the same direction as the wheel rotation, thefan capacity will be diminished. If the air is spinning in theopposite direction of the wheel rotation, the brake horsepowerand noise of the fan will increase. The static pressure of the fanmay also increase slightly, but far less than indicated by theincreased power consumption.The evaluation and control of pre-spinning flow is more difficultthan eccentric flow because of the variety of system connectionsor components that can contribute to pre-spin. Also, spinningoften occurs in combination with eccentric flow such as the casewith the inlet box shown in Figure 4.

Pre-spinning flow can result from any number of commonsituations. Two elbows in close proximity to one another canforce the air to make consecutive turns in perpendicular planesto form a corkscrew effect. As shown in Figure 5, air convergingtangentially into the main duct or plenum can create an obviousspinning effect.

Pre-spinning flow can also be induced by such common aircleaning devices as a venturi scrubber or a cyclone as seen inFigure 6. In these cases, it is often the very function of the aircleaning device to create a spinning effect.

Figure 6 - Fan/Cyclone System

Figure 3 – Poorly Designed Inlet Box

Figure 4 – Eccentric Flow with Pre-Spin

Figure 5 – Spinning Effect

Page 2

CORRECTING BAD INLET CONNECTIONSThe ideal fan inlet connection creates neither eccentric norspinning flow. Where an inlet duct is required, the bestconnection is a long straight duct with straightening vanes.However, it is usually necessary to adapt the system to theavailable space. When space becomes the limiting factor, twochoices are available:

1. Install corrective devices in the duct.2. Increase fan speed to compensate.

The first choice is preferable, but the second is often necessary. Inmany cases, the corrective devices themselves will representsome resistance to flow. A combination of both choices could benecessary to correct extreme field performance problems.

If the fan and system are properly matched, their common point ofoperation should fall within the recommended range on the fanstatic-pressure curve. Figure 7 illustrates the recommended rangefor backwardly-inclined fans. A deleterious system effect couldmove the point of operation to the left on the pressure curve.This would force the fan to operate at an unstable point. The samesituation can occur with any of the basic fan types that exhibitunstable flow characteristics as discussed in Engineering Letter3. When this happens there are three options: alter the system toallow greater flow without increasing resistance significantly,replace the fan with a smaller one, or replace the fan with onethat has a stable curve.

Simple or complex turning vanes, such as those shown in Figure8, can be used to minimize the effects of both eccentric and/orspinning flow. The egg-crate straightener, such as the one shownin Figure 6, can be used in the available space to stop pre-spinand improve fan inlet conditions.

Most of the inlet connections illustrated, with or withoutcorrective devices, can produce losses in performance. Theselosses would be difficult, if not impossible, to predict. Even the inletbox shown in Figure 8, with all the turning vanes installed, couldstill easily represent losses of 10% to 15% of the required flow.

To overcome these losses, the fan speed must be increased to thespeed shown in the fan’s rating table at the required volume and apressure 21% greater than originally calculated:

(110% ÷ 100%)2 = 1.21

Of course the fan’s speed should never be increased beyond thecataloged maximum safe speed!

It is important to note that the increased resistance will not beobserved on the system. The pressure increase is only for thepurpose of selecting the fan to compensate for the lossesassociated with the particular system effect.

The fan laws cannot be applied selectively, only simultaneously.According to the fan laws, if the fan speed is increased 10% for agiven system, the flow through the system will increase 10%, thesystem resistance will increase 21%, and the fan BHP willincrease 33%. This represents an obvious waste of energy due toan often avoidable system-related deficiency. In most cases, sucha change would require the purchase of a larger motor as well as anew drive. If the fan is a direct-connected arrangement, limited toone fixed motor speed, the solution becomes even moreexpensive. These considerations and horsepower penalties apply toall the major causes of system-induced performancedeficiencies.

If the available space dictates the need for a turn into the faninlet, a standardized inlet-box design, with predictable losses,should be used whenever possible.

DISCHARGE DUCTWORK

The connection made to a fan outlet can affect fan performance.An outlet duct ranging in length from 21/2 to 6 fan wheeldiameters, depending on velocity, is necessary to allow the fan todevelop its full rated pressure. If the outlet duct is omittedcompletely, a static pressure loss equal to one half the outletvelocity pressure will result. The system resistance calculationshould include this loss as additional required static pressure.

StaticPressure

BrakeHorsepower

CFM

Figure 7 - Static Pressure Curve forBackwardly-Inclined Fan

Figure 8 – Turning Vanes

Page 3

Air is not discharged from a fan with a uniform velocity profile.The main reason for this is the fact that air has weight and isthrown to the outside of the scroll. Figure 9 shows a typicalvelocity profile.

In a duct with a uniform cross-section, the average velocity willbe the same at all points along the duct. However, where velocitydistribution changes (such as the duct adjacent to the fan outlet)the velocities are not typically the same.

Since velocity pressure is proportional to velocity squared, theaverage velocity pressure at the fan outlet will be higher than theaverage downstream. Since total pressure will be virtually thesame, the static pressure cannot be fully developed until somepoint 21/2 to 6 duct diameters downstream.

Although duct turns directly at the fan outlet should be avoided,there are times when they cannot. In such cases, the turns shouldfollow the same direction as the wheel rotation. Turns made inthe opposite direction of wheel rotation (such as those shown inFigure 10) can have a pressure drop beyond normal systemcalculations. Usually the drop is between .5 to 1.5 fan outletvelocity pressures.

INLET OR OUTLET OBSTRUCTIONS

System obstructions can be as obvious as the cone-shaped stackcap which can have a pressure drop as high as one velocitypressure, or as subtle as the installation of a large fan sheavedirectly in front of the inlet on an Arrangement 3, double-width,double-inlet fan.

One of the most common situations is to place a fan inside aplenum or near some obstruction and fail to account for theeffects on the airflow to the fan inlet. The installation shown inFigure 11 is typical of the sort of non-uniform flow that couldresult in additional losses beyond the normal system calculation.These losses will increase as the velocity increases or as thedistance between the obstruction and the fan inlet decreases.

CONCLUSION

AMCA Publication 201 - Fans and Systems, presents an in-depthdiscussion of system effect and provides methods for estimatinglosses associated with many common situations.

If system effect situations cannot be avoided, their impact onperformance should be estimated and added to the calculatedsystem resistance prior to sizing or selecting the fan. Ignoringthe system effect could lead to difficult field performanceproblems later. It could be that the installed fan does not havethe necessary speed reserve, or the motor is not of sufficientbrake horsepower. The cost of correcting such a fieldperformance problem could escalate rapidly.

System designers need to carefully consider the system effectvalues presented in AMCA Publication 201. By accuratelydefining the true performance requirements of fans in installedsystems, field performance problems can be reduced significantly.

Form 607 GAW

Figure 10 - Poor Fan Outlet Connections

Figure 11 - Plenum System

Figure 9 - Velocity Profile at Fan Outlet


I N C R E A S I N G F A N P E R F O R M A N C EINTRODUCTIONIndustrial processes and plant-ventilation systems often needmore air than originally designed. Increased productionrequirements, process changes, and facility renovations are afew of the major reasons. Additionally, the lack of adequatemaintenance over time can negatively impact system airflows.This letter discusses several procedures that can increase airflow.CHECK THE FAN’S MECHANICAL CONDITION

Often airflow can be increased by adhering to proper fanmaintenance procedures as outlined in fan installation andmaintenance literature.

Properly aligned and tightened V-belt drives. See Figure 1.Fan speed can decrease by as much as 10% to 20% when beltsare too loose, with a corresponding loss of airflow.

Clean airstream surfaces. A fan cannot perform as designed ifthe air flow surfaces are distorted by contaminants. Even inlarge fans, a sixteenth of an inch of build up can reduceperformance.

Check fan rotation. See Figure 2. Centrifugal fans will movesome air even when running backwards. While some typeswould use so much horsepower they would trip circuit breakers,other designs could run for years without being detected.

Check wheel and inlet cone alignment. See Figure 3.Components may be out of position due to routine cleaning orpainting or the wheel could have shifted during shipment. Forbackward inclined fans, the relation of wheel to inlet cone isvery critical. Even a quarter of an inch can have a majorimpact. The fan’s installation and maintenance literature showsthe proper positioning of the wheel to the inlet cone (“A”dimension) or inlet plate.

INSPECT THE SYSTEMThe design and maintenance of the system plays a large role inachieving the overall desired performance. Visual inspectionsoften reveal some easily rectified problems that cansignificantly impair performance.

Check for clogged filters or coils. If the system has not beenproperly maintained, clogged filters or obstructed coils willreduce airflow. The greater the obstruction, the greater the lossin airflow.

Eliminate System leaks. Any leaks in the ductwork willcontribute to reduced performance, especially leaks aroundplenum bulkheads that can lead to recirculation of air. Wornflexible connectors are a common source of leaks and shouldbe inspected regularly.

Verify that dampers are installed correctly and operatingproperly. If the damper linkage is out of adjustment, thedamper may not be opening completely, thereby reducingperformance. If inlet dampers are used, make sure they areinstalled so that the air is pre-spun in the same direction aswheel rotation. See Figure 4. If the air distribution systememploys balancing dampers, make sure they are set properly.

Figure 1 - Poor Drive Alignment and Belt Tension

Figure 2 – Incorrect Wheel Rotation

Figure 3 – Wheel to Cone Alignment

For all dampers, make sure there is sufficient clearance for theblades to open and close completely without hitting theductwork or other system components. Last, for systems witheither pneumatic or electric controls, make sure damperactuators are operating properly.

Look for system effect. Sharp changes in the direction of air-flow at either the fan inlet or outlet will disrupt the flow throughthe fan and impair performance. If it is impossible to straightenthe ductwork entering and leaving the fan, the use of inlet boxesand turning vanes can minimize performance losses as shown inFigure 5. For a more detailed explanation, refer to EngineeringLetter 5, Fan Performance - The System Effect.

INCREASE THE FAN SPEED

One of the easiest solutions to low airflow problems is speedingup the fan. While airflow is increased by speeding up the fan, sotoo are static pressure, noise, and power requirements. Figure 6presents this graphically. Therefore, while increasing the fan’sspeed is an easy procedure with low first cost, the additionaloperating expense over time makes it the most costly solution.See Engineering Letter 2 - Fan Laws and System Curves, foradditional information.

When increasing fan speed, it is necessary to check themaximum safe speed of the fan and make sure the motor iscapable of the horsepower required to run the fan at the newspeed. Never run a fan beyond its maximum safe speed.

ADD OR REPLACE FAN EQUIPMENT

On a first-cost basis, adding or replacing fan equipment is themost costly alternative. However, on a life-cycle-cost basis,considering operating and maintenance expense, it can be theleast expensive, as compared to increasing the speed of anexisting fan.

Sometimes a second fan may be added, either in series orparallel with the original, although it may be more cost effectiveto simply upgrade the system with a new fan capable of therequired airflow and pressure.

Adding another fan in series will increase the airflow because ofthe additional pressure. The operating point of the new systemmoves further out/up the system curve. Where duct size isadequate to handle the desired amount of air but the existing fandoesn’t provide sufficient pressure, a second fan in series maybe the best solution. However, make sure the ductwork canhandle the increase in pressure.

Adding another fan in parallel with the first will increase airflowdue to the combined capacities. Because capacities are beingcombined instead of pressures, a greater increase in airflow willresult for a given system. However, system pressures will alsoincrease and caution is required to avoid the unstable operatingarea of the combined fan system.

CONCLUSION

When more air is required it is important to investigate thesystem on a step-by-step basis, considering the least expensivepossibilities first. For existing systems that seem to have lostperformance, fan and system maintenance is the place to start.Often, simply improving the efficiency of existing componentswill suffice. For systems that require greater airflow and/orpressure, increased fan speed is generally the first alternative.However, when large increases in performance are required,there may be no alternative but to purchase a larger fan.

Form 607 GAW

Figure 4 – Inlet Damper/Fan Wheel Rotation

Figure 5 – Fan Inlet Connections

Figure 6 – The effects on brake horsepower, static pressure andloudness when fan speed is increased.

FIELD TESTING OF FAN SYSTEMS

INTRODUCTION

A fan system may require field testing when the system isthought to be malfunctioning, needs modification or requiresbalancing of its volume and pressure characteristics.

When it has been determined that a field test is required, the testcan provide a complete check on fan performance. This includesdetermination of air volume, fan static pressure and fan brakehorsepower.

This Engineering Letter details the steps involved in performinga field air test. A field test sheet, which simplifies the recordingof test data and the calculation of test results, is provided. A listof safety precautions to be observed while conducting the testis also included.

INSTRUMENTS REQUIRED

1. The best method of measuring both air velocity and staticpressure in the field is with a Pitot tube and manometer.The absence of moving parts, combined with fundamentalsimplicity, make this set of instruments accurate and nearlyfoolproof. Both instruments may be used in nearly anyatmosphere and require no adjustments except for zeroingthe manometer prior to testing. Figure 1 shows a Pitot tubecross-section. Figure 2 demonstrates how it is connectedto the manometer to indicate pressures by measuring thedifference in heights of water columns in the “U” tubes.

Most manometers, such as shown in Figure 3, read directlyin inches of water column. Some manometers may havevelocity graduations marked directly in feet per minutefor use where barometric pressure and temperaturecorrections are normal (i.e., test conditions assumed to be70�F. and 29.92 inches of mercury).

For greater convenience, a more compact Magnehelicpressure gauge may be used with a Pitot tube as a substitutefor the manometer mentioned earlier. These gauges,illustrated in Figure 4, are available in a variety of pressureranges.

2. A clip-on ammeter/voltmeter is used to obtain a reasonableestimate of fan motor horsepower.

3. A calibrated hand tachometer is used to determine the fanRPM.

4. An accurate temperature probe is used to measuretemperature at each test location where volume or staticpressure readings are taken.

Sometimes there are no accessible test duct locations suitablefor use with the Pitot tube. In this case, the air volume can bedetermined at the system entrance or exit, or through a grille orcoil by using an anemometer or velometer. This method, however, is not as accurate and readings should only be taken byexperienced service personnel familiar with this type of testing.

PERFORMING A PITOT TUBE/MANOMETER TEST:

1. Make a sketch of the system as a record and as a guide forselecting locations for taking test readings. Often this willcall attention to poor system-design features. Includedimensions, such as duct diameters or areas, duct length,motor size, motor speed and sheave diameters on beltdrive fans.

Figure 1 – Pilot Tube Cross-Section Figure 2 – Pilot Tube Connection Figure 3 – Pilot Tube/Manometer Test Kit

60527-5530

17

J10543_7 10/16/07 10:50 PM Page 1

2. Determine the best possible location for obtaining the airvolume readings via a Pitot tube traverse (set of readings).The traverse location should not be directly after any turns,transitions or junctions. The traverse should be after aminimum of 2-1/2 duct diameters of straight duct. To obtainthe correct air volume, the Pitot tube and manometer orgauge should be connected to display velocity pressures,not velocities (see Figure 5). The location of the testpoints within each traverse is shown on the field test sheetincluded with this letter.

3. Take static pressure readings several duct diameters fromthe fan inlet and outlet to avoid turbulence (see Figure 6).If the fan has either an open inlet or outlet, assume thestatic pressure to be zero at the opening. Record theairstream temperatures at each static pressure location.

4. Record the fan speed after measuring it with the tachometer. If a tachometer is unavailable, make sure you record themotor nameplate RPM and sheave diameters from whichthe fan speed can be calculated.

5. Read the voltage and amperes supplied to the motor andrecord the values for calculation of fan motor horsepower.

6. Measure the barometric pressure at the fan site with aportable barometer or obtain the pressure from the nearestweather station or airport. Be sure the barometric pressureis correct for your altitude and that it has not been correctedto sea level reference.

7. Determine whether the air being handled contains quantitiesof moisture, particulates and/or gases other than clean air.If so, obtain the concentrations and densities of the gasesor mixture for use in making density corrections.

The attached test sheet is used to calculate flow through afan. For additional information on conducting field tests offan systems, AMCA Publication 203, Field PerformanceMeasurements of Fan Systems, is recommended.

Figure 5 – Air Flow Pressure

Figure 4 – Magnehelic Gauge

Figure 6 – Static Pressure Readings

J10543_7 10/16/07 10:50 PM Page 2

CALCULATING FAN PERFORMANCE

The following steps explain how to calculate density, CFM, SP,and BHP using the acquired test data.

1. Determine the density of the airflow through the fan duringthe test by using the dry-bulb temperature at the fan inletand the barometric pressure. Density in pounds per cubicfoot is determined by:

Densityinlet = 0.075 ( 530 ) (Barometric Pressure)460 + �F. 29.92

2. Determine the density of the airflow at the CFM test location (if different from inlet density) by:

DensityCFM = 0.075 ( 530 ) (Barometric Pressure)460 + �F. 29.92

3. Calculate fan inlet air volume in CFM as measured with thePitot tube and manometer/gauge as follows: First, take thesquare roots of the individual velocity pressures and compute the average of the square roots. Then:

CFMinlet = [ 1096 x test duct area (ft2) ] x

(Avg. of Sum of �VP’s) x ( Density CFM )�Density CFM test Density Inlet

The above calculation gives air volume in actual cubic feet per minute (ACFM) which is the conventional catalograting unit for fans. If standard cubic feet per minute isdesired, it may be calculated as follows:

SCFM = ACFM x ( Actual Inlet Density )Standard Density

4. Determine the fan static pressure (SP) by the followingformula:

SP fan

= SP outlet

- SP inlet

- VP inlet

Where: VP inlet = ( CFM inlet )2x Density inlet

1096 x inlet area in sq. ft.

NOTE: Correct inlet and outlet static pressure to standard values by the following formula before summing.

SP standard = SP actual ( Actual Density )Standard Density

5. Fan motor horsepower may be determined in several ways.The best is to read the volts and amperes supplied to themotor and apply the formula:

For single phase motors:

Fan BHP = Volts x Amps x Power Factor x Motor Eff.746

For three phase motors:

Fan BHP = Volts x Amps x Power Factor x Motor Eff. x �3746

Page 3

This method requires power factor and motor efficiency data,which may be difficult to obtain.

Another method is to draw an amps versus horsepowercurve, (see Figure 7). This is done by plotting a rough horsepower versus amps curve for the motor as follows:

a. Establish no-load amps by running the motor disconnected from the fan (point a).

b. Draw a dotted line through one-half no-load amps, atzero HP, and nameplate amps at nameplate HP (points b).

c. At one-half nameplate HP, mark a point on this line(point c).

d. Draw a smooth curve through the three points (a, c, b).

e. Determine running HP by plotting running amps.

Multiply fan horsepower by the “K” density correction factor to determine HP at standard conditions.

6. Locate volume, static pressure and horsepower on a performance curve drawn at the fan RPM. Curves can begenerated using manufacturer’s fan-selection software atspecific densities, temperature and altitude.

The test plot values will probably not fall exactly on the curve. If the fan system has been designed and installed properly, the difference should be small, reflecting test accuracy. If the difference is great, the system should be analyzed as described in the next section. Figure 8 shows a typical fan curve and field test points which fall on the curve.

Figure 7 – Amperes versus Horsepower

Figure 8 – Typical Fan Curve and Field Test Points

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POOR PERFORMANCE TEST RESULTS

If the test results indicate poor fan performance, a number ofsimple steps can be taken that could improve performance.

Be sure that any dampers at the fan inlet or outlet are set to thecorrect position and that no other system dampers such as firedampers, smoke dampers or balancing dampers have beeninadvertently closed.

A frequent cause of poor fan performance is the presence of poorinlet connections. Sharp elbows, inlet boxes without turningvanes and duct configurations causing the air to spin uponentering the fan, are examples of undesirable inlet connections.

Fan performance is also impacted by poor outlet conditions.Examine the outlet connection, keeping in mind that sharpelbows, rapid expansions, reductions or the absence of an outletconnection all together can reduce fan performance.

By connecting the Pitot tube and manometer/gauge to readvelocity pressure and inserting the Pitot tube through a hole atthe inlet connection (as illustrated in Figure 9), pre-spin can be

determined. Once inserted, slowly twist the tube. The angle atwhich air is entering the fan can be determined by observingthe angle of the tube generating the highest gauge reading. Ifthe angle deviates noticeably from being parallel to the fanshaft, the air entering the fan inlet may be spinning andtherefore reducing fan performance.

Another reason for poor performance could be stratification ofthe air entering the fan. By taking four temperature readingsninety degrees apart in the inlet duct near the fan, the possibilityof stratification can be determined. A temperature difference of10 degrees or more in the readings indicates stratificationexists. An illustration of stratification is shown in Figure 10.

Refer to Engineering Letters 5 and 6 for more detailedexplanations of system effect and improving fan performance.

SAFETY PRECAUTIONS

The included list of safety precautions should be observedwhenever testing or servicing fan equipment.

Form 1007

Figure 9 – Testing Fan Inlet for Spinning Airflow Figure 10 – Condition Causing Stratification

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P R O P E R S E L E C T I O N O F P R E S S U R E B L O W E R SINTRODUCTIONIn general terms, a pressure blower provides relatively highpressure at low volume when compared to other types ofcentrifugal fans. For purposes of this letter, fans with volumesto 10,000 CFM with pressures to 80" WG are consideredpressure blowers. Typical applications require constant pressurethroughout the system’s operating range. A fan outlet damper orsystem damper is usually used to control air volume.Consequently, a requirement of pressure blowers is that theyprovide stable performance from full-closed to full-open.

Most pressure blowers employ a radial-blade wheel design.New York Blower’s research has resulted in a unique wheeldesign that is not truly radial. The blades are slightly cantedbackward and dual tapered from the hub to the blade tip. SeeFigure 1. This design provides better efficiencies and, as aresult, significantly lower noise levels. The volume-pressurecharacteristics remain the same as radial-blade wheels.

POINT OF OPERATIONSince typical pressure-blower applications require a constantpressure, selections are normally near the flat peak of the staticpressure curve. See Figure 2. Because of the flat nature of thepressure-blower curve, a typical question is, “what keeps thefan’s performance from fluctuating between different points onthe fan curve?” The answer lies in the relationship between thefan’s performance curve and the system curve.

At a given RPM, the fan can only operate on its performancecurve. The only way to alter this curve is to either increase ordecrease the fan’s speed. Conversely, the system can onlyoperate along one system curve. The only way to change thissystem curve is to increase or decrease the resistance throughthe system. Since the two curves can only intersect at onepoint, the actual performance of the fan can occur only at theintersection of the fan curve and the system curve. This isdepicted in Figure 3.

Figure 2 – Typical Pressure Blower Performance CurvesNote: Broken lines denote typical system curves.

Figure 1 – Dual-Tampered Pressure Blower Wheel

Figure 3 – Typical Pressure Blower and System Curves

Considering that pressure blowers are often selected near thepeak of their pressure curve, dampering usually results in anoperation left of the pressure peak. One benefit of radial-bladewheel design is that it delivers stable performance left of peak.

Radial wheels bring other advantages to pressure blowers. Theradial design delivers greater pressures at a specific RPM thanboth the radial-tip and backwardly-inclined designs. Theinherent strength of the radial wheel allows for the relativelyhigh wheel tip speeds required for the development of highpressures. Remember, pressure is approximately proportional tothe square of the change in wheel tip speed. Therefore, a 2 PSIpressure blower must be capable of speeds 1.414 times as fast asa 1 PSI unit.

1.4142= 2

SINGLE-STAGE VS. MULTI-STAGE

Single-stage pressure blowers are the most common and leastexpensive of the two designs for the range of flows and pressuresnoted in the introduction. A single-stage pressure blower consistsof a single wheel in a volute-shaped housing design, such asshown in Figure 4.

Single-stage units are usually far more economical in applicationsup to about 3 PSI. They are also less complex and easier tomaintain than multi-stage pressure blowers. Power consumptionis also less because the single-stage blowers are more efficient.

It is possible to place two, and sometimes more, single-stagepressure blowers in series to develop pressure as high as 5 PSIand still represent an economical alternative when compared tothe multi-stage units for the same performance. There is theadded reliability factor of being able to “limp along” with oneunit while the other unit is down for maintenance. When amulti-stage unit is down, the entire system is down. Consult themanufacturer for proper selection and application informationwhen designing pressure blowers for series operation.

SELECTION PROCEDURES

Selecting pressure blowers or any other type of fan forapplications involving relatively high pressure requires somespecial considerations. Pressure blowers are generally used withthe pressure entirely on the inlet or entirely on the outlet. Air iscompressed as it passes through the fan, lowering the volumeand raising the density. In negative pressure systems, air israrefied to become less dense. The extent to which the effects ofcompression and rarefication must be considered dependslargely on the degree of accuracy employed in the actual systemdesign and calculation process.

During compression there is also a temperature rise associatedwith the energy expended to overcome the system resistanceand fan inefficiency. The rule of thumb is to allow 1°F.temperature rise for every 2" static pressure differential. Forexample: a supply fan with 40" SP at the outlet will develop a20°F. temperature rise at the fan outlet, as compared to the airtemperature at the fan inlet. To determine the proper air volumefor selection purposes, the effect on density of both compressionand temperature must be considered.

One notable exception to these rules for performance correctionsis the combustion-air-supply application. Burner manufacturersuse SCFM ratings to arrive at lbs./hr. of air. The air will becompressed through the fan to a proportional lower volume, yethigher density so that the total weight of air in lbs./hr. remainsconstant and is sufficient for the combustion process.

PERFORMANCE CORRECTIONS

Fan performance is based on a standard density of .075 lbs./ft.3

Density corrections for positive or negative pressure are basedon changes in absolute pressure.

A. Standard absolute pressure is 408" WG at sea level.

B. Compressed density for + 40" SP at the fan outlet is:

408 + 40”( 408” ) x .075 = .082 lbs./ft.3

C. Rarefied density for - 40" SP at the fan inlet is:

408 - 40”( 408” ) x .075 = .0676 lbs./ft.3

Density corrections for temperature changes are based onabsolute temperature in degrees Rankin (°R).

A. Standard absolute temperature is 530°R., 70°F.(0°F. = 460°R.)

B. A 20° temperature rise over a fan inlet temperature of70°F. gives the following density:

460° + 70°( 460° + 70° + 20° ) x .075 = .072 lbs./ft.3

Also refer to the following sample selections

Figure 4 - Single-Stage Pressure Blower

SAMPLE SELECTIONS

Example 1: No performance correction due to compression.

What actually happens in the system?A. 2300 ACFM at 70°F. at 408" atmospheric pressure

enters the pressure blower inlet (A).B. The pressure reading at (B) is 34.6" gage pressure or

408" + 34.6" = 442.6" absolute. The temperature hasincreased to 87°F.

34.6( 2 + 70°)C. Density ratio is:

442.6 460° + 70°408 x 460° + 87° = 1.05

D. Air density at the burner (B) will be:.075 x 1.05 = .0788 lbs./ft.3

E. ACFM at (B) will be:2300 ÷ 1.05 = 2190 ACFM

F. The SCFM equivalent at (B) will be:2190 x .0788 = 172.6 lbs./minute2300 x .075 = 172.5 lbs./minute

Note: The changes in volume and density can be ignored in thiscase because the proper amount of air by weight will still beavailable at the burner (B). Select the pressure blower for 2300CFM at 34.6" WG pressure at .075 lbs./ft.3 density.

Example 2: Performance correction required due to compression.

Given: injector conveying system, as illustrated.

Required: 2300 CFM for the velocity required at (B).Resistance is 20 oz. or 34.6" WG.

What actually happens in the system?

A. Air enters at 70°F. at 408" atmospheric pressure at thepressure blower inlet (A).

B. The pressure reading at (B) is 34.6" gage pressure or408" + 34.6" = 442.6" absolute. The temperature hasincreased to 87°F.

34.6( 2 + 70°)

C. Density ratio is:442.6 460° + 70°408 x 460° + 87° = 1.05

D. Air density at the burner (B) will be:075 x 1.05 = .079 lbs./ft.3

E. ACFM at (B) will be:2300 ÷ 1.05 = 2190 ACFM

F. To get 2300 ACFM at (B), the volume of air enteringat (A) must be increased by the density ratio:2300 x 1.05 = 2415 ACFM

Select the pressure blower for 2415 CFM at 34.6" WG pressure at.075 lbs./ft.3 density.

Example 3: Performance correction due to negative pressure.

Given: draw-thru pneumatic conveying, as illustrated.

Required: 4800 SCFM at - 34" WG.

What actually happens in the system?

A. Air enters at 70°F. at 408" atmospheric pressure at thesystem inlet (A).

B. The resistance at the pressure blower inlet (D) is - 34"gage pressure or 408" 34" = 374" absolute.

C. Density ratio is:374( 408 ) = .92

D. Air density at (D) will be:.075 x .92 = .069 lbs./ft.3

E. To get - 34" at (D) at .069 lbs./ft.3 density, the pressuremust be increased by the density ratio for proper fanselection: -34" ÷ .92 = - 37" WG.

F. Capacity = 4800 ÷ .92 = 5217

G. Select the pressure blower for 5217 ACFM at 37" WG.

H. Operating horsepower would be:.92 x rated BHP, corrected for the lower density.

Note: The actual air volume at the fan outlet will be less thanthe volume at (A) by the density ratio, but the actual air volume atthe fan outlet is not important in this system.

NOISE ATTENUATIONA rising concern in many of today’s industrial applications isOSHA’s criteria for noise levels. To meet these requirements,many pressure blowers require sound attenuation. Thebackward-canted and dual-tapered wheel design can result in an8-10 db noise reduction over the traditional straight bladedesign. In some cases, this may eliminate the need for asilencer.

If attenuation is required, silencers are readily selected based ontheir connection to either the inlet or outlet of the pressure

blower. The most common connection is directly on the blower,flange to flange. See Figure 5. Silencers are rated in dynamicinsertion loss (DIL) in decibels. These values are subtractedfrom the pressure blower sound power level’s eight octavebands.

The pressure drop through the silencer must be added to thesystem requirements, but generally the values are less than 0.2"and are insignificant.

Figure 5 – Pressure Blower Silencer

Form 607 GAW


PNEUMATIC CONVEYINGINTRODUCTIONA well designed pneumatic conveying system is often a morepractical and economical method of transporting materials fromone point to another than alternative manual or mechanicalsystems. This Engineering Letter outlines some of thefundamental principals of pneumatic conveying systems andexplains various special considerations for fan selection.

TYPES OF PNEUMATIC CONVEYING

Pneumatic conveying encompasses numerous different systemdesigns, technologies, and pressure ranges; however, there areonly three basic methods for moving material with air. Thesecan be categorized into the following system types:Dilute-phase conveying is the process of pushing or pullingair-suspended materials from one location to another bymaintaining a sufficient airstream velocity to capture andconvey the suspended particles.

Dense-phase conveying relies on a pulse of air to force a slugof material from one location to another. This form of conveyingusually requires positive displacement blowers or compressorsto generate the necessary pressure of 1.5 to 30 psig or more.

Air-film or air-float conveying is a means of moving productalong a conveyor on a cushion of air.The use of fans for pneumatic conveying is generally limited todilute phase conveying and air film conveying.

DILUTE-PHASE CONVEYINGIn this method of conveying, material is suspended in theairstream. Suction or vacuum are not factors in this type ofsystem and fan static pressures are no greater than 60" WG. Ifthe system uses a fan on the exhaust end and the material iscollected or separated from the airstream before it reaches thefan, the fan itself can be of a more efficient type such asbackwardly inclined. If the system is designed so that thecombined material and air mixture passes through the fan,selection is limited to the more rugged but less efficient fantypes intended for material laden airstreams. A number ofradial-blade wheel designs are available to handle variousconcentrations, sizes, and types of airborne particles. Radial-tipwheel designs are tolerant of airborne contaminants, but radial-tip fans are not generally thought of as bulk material handlingdesigns. In all cases, the fan manufacturer should be consultedto determine the most appropriate fan type available to handlethe specific material quantity and type, but it must be understoodthat the fan manufacturer can neither control the variables inpneumatic conveying systems nor provide any guarantee of theservice life of the fan itself.

Applications requiring fans for dilute-phase pneumatic conveyingfall into one of three basic categories: dust collection, fumeremoval, or material conveying.

DUST COLLECTION AND FUME REMOVALDust collection, fume removal, and material conveying systemseach have unique characteristics, but all three are similar in theirdependence upon proper air velocities.

Dust collection and fume removal are generally thought of as“housekeeping” systems that usually incorporate a hood at thesystem entry point. There are many types and styles of hoods incommon use, and hood design is a subject in itself. Some stateand local codes offer hood design criteria, and there are severalreference texts, such as Industrial Ventilation - A Manual OfRecommended Practices, that can assist in the selection anddesign of hoods. In all cases the hood design should minimizeturbulence and offer the lowest possible entrance losses.

Determining the minimum velocity for dust collection or fumeremoval is often a matter of practical trial-and-error judgment.State and local codes may dictate minimum velocities for certainmaterials. Where no codes apply, the velocities shown in Figure1 can be used as conservative estimates. Since these velocitiesare conservative, it is often possible to reduce them throughexperimentation. Reducing the velocity to near the settling pointwill generate the lowest overall operating cost but raises the riskof system plugging, increased maintenance costs, and lostproduction.

Dust Collecting and Fume RemovalDuct Velocities

Material Velocityin FPM

Material Velocityin FPM

1. Grinding Dust 50002. Foundry Dust 4500 20. Jute Dust 35003. Sand Blast Dust 4000 21. Grain Dust 30004. Wood Flour 2000 22. Shoe Dust 40005. Sander Dust 2000 23. Rubber Dust 35006. Shavings, Dry 3000 24. Rubber Buffings 45007. Shavings, Wet 4000 25. Bakelite Moulding8. Sawdust, Dry 3000 Powder 35009. Sawdust, Wet 4000 26. Bakelite Moulding

10. Wood Blocks 4500 Dust 250011. Hog Waste 4500 27. Oven Hood 200012. Buffing Lint, Dry 3000 28. Tail Pipe Exhaust 300013. Buffing Lint, Wet 4000 29. Melting Pot and14. Metal Turnings 5000 Furnace 200015. Lead Dust 5000 30. Metallizing Booth 350016. Cotton 3000 31. Soldering Fumes 200017. Cotton Lint 2000 32. Paint Spray 200018. Wool 4000 33. Carbon Black 350019. Jute Lint 3000 34. Paper 3500

Figure 1

MATERIAL CONVEYING

Although the differences between dilute-phase materialconveying systems and dust collection or fume removal systemsmight appear to be minimal, there are certain distinctions thatare critical to the successful operation of material-conveyingsystems. These differences include the method of introducingthe material to the hood, the velocity requirements, the ductconfiguration, and the fan type.

The introduction of material into a material conveying systemcan be difficult. The most important criterion is to feed thematerial into the airstream evenly. This can be accomplished bymeans of gravity or by a mechanical device.

A hood or hopper can be used as a gravity feeder. Use of thesecomponents is limited to dry, free-flowing materials. It isimportant to remember that it is the velocity moving around andpast the material that induces it to flow. If the entry becomesplugged with material, the required velocity cannot bemaintained, significantly impeding air and material flow.

A venturi feeder can be used to introduce material into theairstream. Like the hood, it has no moving parts so there isvirtually no maintenance. However, the design of the venturimust be tailored to each application and even the best ones can berather easily blocked if system conditions vary. Typical throatvelocities are 2 to 3 times the velocity in the main duct . . . seeFigure 2.

Rotary valves and screw-type (auger) feeders (see Figure 3) arethe most common mechanical devices used to introducematerial into the airstream. Both types offer a controllable flowrate and are readily available in a number of standard designs tohandle pressures common to dilute phase conveying. However,there are some precautions. Both are typically more expensivethan gravity-feed alternatives. Rotary valves can experienceinternal air recirculation which causes a reduction in materialthrough-put. The screw-type feeder is a relatively highmaintenance device. In either case, the manufacturer of thespecific feeder should be consulted for selection, equipmentrecommendations, and system limitations.

Since the purpose of a conveying system is to move quantitiesof material suspended in air, the ratio of material to air (byweight) is critical. The most common form of reference is tostate the ratio according to the combined weight in pounds perhour. A conservative design approach is to keep the ratio ofmatter-to-air below a 1:2 proportion. However, successfulsystems have been designed using material loadings of 1:1 ormore when the system components are well-designed andeliminate sharp turns, abrupt junctions, or other potential pointsof binding, clogging, or drop-out and the material beingconveyed is well-defined and consistent.

Certain minimum conveying velocities must be maintained tokeep the material in suspension and flowing. To some extentthese velocities are dictated by, or at least related to, the mater-ial-to-air ratio. For example, conveying sawdust at a rate of1800 lbs./hr. through a 6" pipe with a material loading ratio of1:2 will result in an air velocity of 4073 FPM.

1800 lbs./hr. material = 30 lbs./min.

60 lbs./min. air ÷ .075 lbs./ft.3 std. density = 800 CFM.

6" pipe = .1964 ft.2 area inside.

800 CFM ÷ .1964 ft.2 = 4073 FPM.

Figure 4 provides conservative minimum conveying velocitiesto be used for some common materials. The velocity shown forsawdust is 4000 FPM. If the same 1800 lbs./hr. of sawdust hadbeen introduced to a system with a 1:1 design ratio and therewere no other changes to the system, the resulting velocity wouldonly be half and the material would probably settle and clog. Tocompensate for the lower ratio, the pipe size could be reduced to4", but this might introduce new problems in feeding thematerial to the pipe or transitioning to the fan. In this example,the 1:2 ratio would seem to be ideal.

Figure 2 – Typical Venturi Feeder

Figure 3 – Typical Rotary Valve Feeder

Material ConveyingDuct Velocities

Material Velocityin FPM Material Velocity

in FPM

1. Wood Chips 4500 12. Cotton 40002. Rags 4500 13. Wool 45003. Ground Feed 5000 14. Jute 45004. Powdered Coal 4000 15. Hemp 45005. Sand 7500 16. Vegetable Pulp,6. Wood Flour 4000 Dry 45007. Sawdust 4000 17. Paper 50008. Hog Waste 4500 18. Flour 35009. Pulp Chips 4500 19. Salt 6000

10. Wood Blocks 5000 20. Grain 500011. Cement 6000 21. Coffee Beans 3500

22. Sugar 6000

Figure 4

Sufficient velocities must be maintained throughout theconveying system to avoid material settling. All airbornematerials, except the finest of dusts or fumes, can settle in asystem or even in the fan itself. When settling occurs in thehorizontal plane, it is known as salt ation. When settling occursin the vertical plane, it is called choking.

Saltation is probably the most difficult to avoid because eventhe smallest ridge or duct seam can begin the process.Whenever possible, it is advantageous to employ the aid ofgravity to minimize potential build-up by designing the pipingor ductwork with a downward slope. This is particularly truewith fine granular materials.

Choking in downward movement often occurs in the verticalline as a direct result of saltation in the adjacent horizontalline. Upward movement is often easier to control because allthat is needed is sufficient momentum (velocity) to keep thematerial in suspension. All falling materials simply drop backinto the airstream. However, choking in the upward flowdirectly above the fan discharge poses additional problems. Ifenough material is forced back into the fan where itrecirculates, the fan will exhibit premature wear due toexcessive loading.

To minimize the potential for saltation or choking, it isrecommended that some provision be included in the systemfor bleeding in excess air through adjustable vents or dampers.See Figure 3. This excess air will effectively increasevelocities in the system to assist material transportation. It isimportant to remember that the fan selection must account forthe maximum potential excess air, and that handling more airthen the minimum system requirements will result in increasedpower consumption.

FAN SELECTION

Just as designing around a velocity that is too low will impedethe material conveying capability of the system, unnecessarilyhigh velocities can also be detrimental. System resistanceincreases as the square of the increase in velocity. Therefore,additional energy is required to overcome that resistance.Also, the abrasive or erosive characteristics of the materialbeing conveyed will increase with an increase in velocity,shortening the service life of all system components.

Only the air volume is considered in determining the velocity.The material volume is ignored to compensate for the periodsof inconsistent material loading that occur during start-up andshut-down. However, the material content of the overallairstream mixture cannot be ignored when calculating systemresistance or when sizing the fan.

Fans are constant volume machines that discharge a fixedvolume of air at a fixed speed. If a fan is required to handle agiven volume of air and a given volume of material, it shouldbe sized to handle the combined volume. Using the previousexample, 1800 lbs./hr. of sawdust at an average bulk density of11 lbs./ft.3 results in 164 ft.3/hr. or nearly 3 CFM. The fanshould be selected to handle 803 CFM (800 + 3). In thisexample the 3 CFM is negligible. However, in situationswhere greater material volumes are being handled or when thebulk material density is much lighter, the volume cannot beignored.

The effects of the material on system resistance must beconsidered. Since most materials usually exhibit a lowercoefficient of friction than air, a simple density correctionbased on the combined weight and volume of the air/materialmixture would result in an unnecessarily high correction. Nodependable methods of determining the flow resistance ofair/material mixtures have been proven, so only reasonableestimates are available. Some researchers have theorized thatthe bulk material content merely acts to reduce the effectivearea of the pipe or duct and so ignore the density effect bycalculating air resistance through the resulting smaller pipediameter. The best method for determining the resistance ofthe air/material mixture is through pilot-plant testing orexperimentation. Figure 5 provides correction factors that canbe used as reasonable starting points for estimating resistance.

FRICTIONMULTIPLIER

CUBIC FEET OF AIR PER MINUTE PER POUND OF MATERIAL

Figure 5 – Resistance Factors

MULTIPLY FRICTIONFOR CLEAN AIRBY MULTIPLIER

Even though the air/material mixture does not follow thetraditional laws of fluid flow as they apply to friction orresistance, it is suggested that the fan brake horsepower (BHP)will increase according to the bulk density of the mixture. Thecombined weight and total volume can be used to determinethe maximum airstream density for selecting a motor that willhandle the fan BHP at the bulk density.

Where,

1800 lbs./hr. material + 3600 lbs./hr. air =5400 lbs./hr.5400 ÷ 60 = 90 lbs./min.90 ÷ 803 CFM = .112 lbs./ft.3 bulk density

To determine the approximate BHP for this example, multiplythe rated BHP at standard density of .075 lbs./ft.3 by 1.5.

(.112 ÷ .075) = 1.5

It is sometimes thought that a larger fan is naturally better than asmaller one. This is far from correct since material is just asliable to settle in a fan as in a duct. If the inlet and outletvelocities of a fan are at least as high as the minimum conveyingvelocity, no settling should occur in the fan. This is true forboth dust collection and conveying.

AIR-FILM CONVEYING

This method of pneumatic conveying uses a film or cushion ofair to move items such as cans, boxes, or plastic containersthrough a plant. Used primarily in the packaging industry, airfilm conveying usually requires fan static pressures of no morethan 8" WG. In most cases, the system utilizes several smallerfans as opposed to one large fan. Because the air is clean,various fan types can be used in these systems, includingbackwardly inclined and radial-bladed designs. Selection isbased on pressure and flow, but configuration is equallyimportant.

Either positive pressure or vacuum can be used to move thecontainers. In a pressurized system, air is directed through adrilled or slotted surface, where the air is discharged at a slightangle in the direction of flow. The greater the discharge angle,the higher the velocity from one station to the next. Vacuumelevators are used to raise or lower containers to differentlevels in the system by holding them to a moving, perforatedbelt. Vacuum transfer devises allow fallen or damaged productto drop out of the system, thereby reducing downtime andmaintaining efficient high-speed processing. Both techniquesmay be employed in different portions of complex conveyingsystems.

The benefits of air film conveying over conventional mechanicalconveying include:

Increased process speed. Lower maintenance costs (fewer moving parts). Reduced energy consumption. Reduced noise and safety hazards. Reduced downtime from jamming. Gentler handling of the product.

Many companies in the packaging industry use a combination ofair and mechanical conveying systems in their manufacturingprocesses.

CONCLUSION

Pneumatic conveying systems have limitations, and alternatemanual or mechanical means cannot be ruled out. However,pneumatic conveying systems usually require less plant space,can be easily installed in the available or wasted space, can beeasily automated, can usually be easily altered for futurechange, and usually carry a lower capital cost. Beyond theseeconomic advantages, pneumatic conveying systems can alsobe useful in controlling or minimizing product loss, improvingdust control, and thus improving overall plant conditions.

Form 607 GAW


F AN S AN D B L O W E R S F O R C O M B U S T I O N P R O C E S SINTRODUCTION

The burning of gas, oil, coal, or other combustible materialrequires air. When the end result of the burning is to be anefficient combustion process, in compliance with Federal andState Clean Air Act requirements, the volume of supply airmust be reliably controlled. Insufficient air volume will result inwasted fuel and excessive particulate along with potentiallyexplosive gases in the exhaust system. Too much air increasesthe amount of heat carried up the stack by the excess draft.Either extreme increases the cost and difficulty of controllingexhaust emissions.

NATURAL AND MECHANICAL DRAFT

Air can be supplied to the combustion process by natural ormechanical draft.

Natural draft simply refers to the use of a chimney or stack toinduce an upward flow of air. The stack effect pulls air intothe combustion chamber, as shown in Figure 1. The amount ofairflow depends on the stack’s height and diameter, theprevailing wind velocity, and the resistance of the burnermechanism or fuel bed itself. The demand for air-cleaningapparatus on combustion systems, particularly oil and coal,has increased the overall system resistance to such an extentthat natural draft alone is seldom sufficient.

Mechanical draft refers to the use of fans or blowers to createairflow through the combustion area. When mechanical draftis incorporated, the chimney or stack is used primarily todirect the exhaust gases up and away where they will not be anuisance. Because wind velocity and direction are lessimportant, the combustion process can be much more carefullycontrolled.

Mechanical draft is accomplished in one of two ways: whenair is blown or forced into the combustion chamber it is knownas forced draft . . . see Figure 2; when the air is drawn throughthe combustion chamber it is called induced draft . . see Figure3. When both forced and induced draft are used, the system istermed a balanced-draft system.

Generally, the fans or blowers used for induced-draft applicationsare larger and more expensive than those used for similarforced-draft applications. The combustion process itselfcreates gases and elevated temperatures that expand theexhaust airstream, requiring fans with greater volumetriccapacity than would be required on the supply side of thecombustion process to supply clean, ambient air. Also, the hotexhaust serves to lower the density of the airstream, so densitycorrections must be applied to the fan static pressure (SP) toovercome the actual system resistance. The fact that theexhaust or flue gases are hot often requires induced-draft fans tobe of a construction suitable for higher temperatures.

Figure 1 – Natural Draft

Figure 2 – Forced Draft

Figure 3 – Induced Draft

The first induced-draft fans were applied to hand-fired, solid-fuelboilers where the combustion chamber had to be at a negativepressure to permit the operator to shovel in fuel.

When oil and gas became primary fuel sources, boiler designerswere able to seal the combustion chambers. As a result, forced-draft fans became popular. The advantages were lower fanpower and fans handling clean air (no corrosion or abrasion) atambient conditions. These factors encouraged the use of airfoilfans, further reducing power consumption. This led to thealmost universal use of pressurized firing in gas, oil, andpulverized coal boilers by the mid 1950s.

However, by the late 1 960s the combustion process industryhad learned through experience that it was impossible to maintainan airtight quality in large (approximately 150,000 lbs./hr. andlarger) industrial and power boilers. These units (some over 100feet high) simply had too much thermal expansion. Fly ash andnoxious and corrosive fumes were creating tremendousmaintenance and personnel problems. This led to the developmentof balanced-draft systems, in which both forced-draft andinduced-draft fans are used. (See Figure 4)

Forced-draft fans and blowers are common for cast iron firetubeand small water tube boilers. The fan or blower serves toprovide the air and the velocity necessary for the fuel-to-airmixture to enter the actual combustion chamber. When used inconjunction with induced draft, the forced-draft fan is oftencalled the primary air fan since it provides the primarycombustion supply air. The induced-draft fan provides theairflow necessary to overcome system resistance and exhaustthe flue gases.

Some combustion systems draw hot, perhaps dirty air fromother processes. Forced-draft fans for such systems are calledgas recircula tion fans, and must be selected for the rigorousconditions under which they will operate.

ALTERNATIVE FAN DESIGNS

Figure 5 shows a typical Pressure Blower for forced-draftapplication. This type of unit is normally direct connected to a3600 RPM motor and develops pressure sufficient to overcomethe total system resistance on small combustion systems. Tominimize motor bearing load and starting current, the wheel isnormally fabricated of high-strength aluminum. It is thereforelimited to handling clean, often filtered air. Pressure Blowersare commonly used on small firetube boilers.

Figure 6 shows a typical backwardly-inclined airfoil wheelcommonly used for forced-draft fans on balanced-draft boilers.This type of fan is preferable because it typically is highvolume, medium pressure and is usually the most efficient fanselection. Some airfoils, such as the nyb AcoustaFoil, arecapable of stable operation throughout a complete, damperedrange from wide-open to closed-off so the combustion rates canbe closely controlled with inlet or outlet dampers.

On boilers that use hot, dirty gases for combustion supply, thegas recirculation fan most frequently selected is a radial-bladetype. This type of wheel is considered to be the most “rugged”and will run at lower tip speeds. It is therefore less subject toabrasion than radial-tip or backwardly-inclined wheels. (SeeFigure 7)

Figure 4 - Balanced Draft - Forced and Induced

Figure 5 – Pressure Blower

Figure 6 – Backwardly-Inclined Airfoil Wheel

Figure 7 – General Industrial Wheel (Radial)

Radial-blade fans were at one time commonly used for induced-draft service. However, as pollution requirements have becomemore stringent and control devices have been added to reduceflue gas particulates (ahead of the induced-draft fan), radial-tipblade or even backwardly-inclined fans have become populardue to their higher efficiencies and higher volumetriccharacteristics. (See Figures 8 and 9) The exception to this iswhere high efficiency scrubbers are used and the pressurerequirements are increased to where the radial-bladed fans aremore suited.

The combustion of coal and most fuel oils will release sulfurfumes into the flue gas. If a wet scrubbing or cleaning apparatusis used, water vapor will combine with the sulfur to formsulfuric acid. This can place severe constraints on the fan typesavailable to handle this highly corrosive gas stream. For thisvery reason, flue-gas-desulfurization (FGD) equipment isdesigned into the pollution control systems of many combustionprocesses. Another alternative to reduce the potential for sulfuricacid in the exhaust system is to mix lime or crushed limestone ina fluidized bed combustion process so the lime will neutralize thesulfur and stabilize the pH of the exhaust gases.

FAN SELECTIONIdeally, the fan in any combustion process will supply justenough air to completely burn all the fuel, and no more. Thiswill help keep heated, but unused, air from going up the stack.Actually, this idea is approachable with gas burners butimpractical with wood- or coal-fired combustion. Thus, nearlyall air volume requirements for combustion processes arecalculated to include some margin of excess air.

As stated in the introduction, insufficient air volume will resultin wasted fuel and excessive particulate along with potentially

explosive gases in the exhaust system. From this viewpoint it isbetter to include excess air volume. Some typical excess airpercentages are shown in Figure 11 for reference only.

The amount of air required for theoretically perfect combustion isbased on the portion of the combustible substances carbon (C),hydrogen (H 2), oxygen (O2), and sulfur (S) contained in fuel.These are the only combustibles found in common fuels.

Air Required for CombustionCombustible

SubstanceLbs. of Air Per

Lb. of Combustible

CH2

O2

S

CarbonHydrogenOxygenSulfur

11.534.3

--4.3

Figure 10

The ratio of air required for perfect combustion to a pound ofeach element in the fuel is shown in Figure 10. This ratio shouldbe multiplied times the percentage of the element in the fuel,times the weight of the fuel to get the required weight of air, andthen excess air must be added and the result must be correctedto the corresponding air volume per minute.

For example: Assume a fuel oil with 86.1% C, 13.8% H2 , 0.1% S,and negligible free oxygen. The fuel oil weighs 6.8 lbs./gallonand will be consumed at a rate of 5 gallons each hour.

.861 C x 11.5 x 6.8 = 67.3 lbs. air.138H2 x 34.3 x 6.8 = 32.2 lbs. air

.001 S x 4.3 x 6.8 = .03 lbs. airTotal = 99.53 lbs. air

99.53 lbs. x 1.10 excess x 5 = 547.4 lbs./hr. air

547.4 ÷ 60 = 9.12 lbs./min. air

In a combustion supply application handling standard densityair, this equates to (9.12 lbs./min. ÷ .075 lbs./ft.3 = 121.6 CFM).

Although there are a number of accepted methods for determiningcombustion air requirements . . . some rules of thumb, someexact calculations . . . they all rely on the actual portion ofcombustion constituents found in the fuel in question. Figure 11lists some typical examples, but a full, accurate list would beimpractical as there are hundreds of unique coal grades. Inpractice, the combustion system designer should determine theactual air volume requirements and the excess air margin basedon an analysis of the fuel in question.

In addition to the fundamental volume specifications, combustionprocess fans, particularly larger fans, are often specified for twoconditions . . . actual and test block. The actual condition is thecalculated volume (including excess air) and pressurerequirements. The test block condition is a theoretical duty thatincludes some safety factor beyond the actual volume andpressure requirements. The fan selection for the applicationshould be capable of meeting both conditions with good efficiency,economy, and stability. Whenever possible, the actual conditionshould represent the most efficient point of operation for the fanselected for the application.

Figure 8 – Radial Tip Wheel

Figure 9 – Flat Blade Backwardly-Inclined Wheel

Typical Pe rce nta g e o f C o mb u st i b l e s (% Weight)Fuel

Typical ExcessAir Range

(% Volume) Carbon Hydrogen Oxygen Sulfur

Fuel Oil 5-20No. 1 86.3 13.7 -- 0.3No. 2 87.2 12.9 -- 0.5No. 4 87.9 11.8 -- 1.1No. 5 87.9 11.3 -- 1.8No. 6 88.4 10.8 -- 2.1

Natural Gas 5-15 70.6 22.7 1.4 0.3Wood, Pine 10-25 54.3 5.6 37.9 0.1

Coal (ref. only) 10-60 85.0 5.4 5.8 1.5Coke 10-30 80.0 0.3 0.5 0.6

Figure 11 - Typical fuel analysis of excess air requirements and amount of combustibles.

Typically, direct-drive fans are preferred to belt-drive fans.Direct-drive fan arrangements used are 4, 7, and 8.

To reduce volume and pressure to meet the actual design orreduced load conditions, inlet dampers or variable frequencydrives are used. Variable speed offers the most efficient meansof performance reduction, although the initial cost andequipment maintenance is greater than that of dampers. Thesemust be evaluated on an individual job basis to determinewhether the power savings will offset the greater initial pricedifferential and added maintenance costs.

Belt-drive fans, Arrangements 1, 3, 9, and 10, can be selectedfor the most efficient operation at the actual operating conditions.It is usually enough that belted units simply have sufficient speedreserve to meet the speeds necessary to fulfill the test blockcondition by means of a change in drive sheaves.

The criteria for selecting the fan motor is usually specified perjob. Often, the motor is sized to handle the hot test blockconditions so the fan can be dampered for low load periods suchas start-up or shut-down. This reduces the dampering, or turn-down, range required under actual conditions.

FAN CONSTRUCTION

Fans used in combustion processes, whether forced or induceddraft, should be capable of meeting the following minimumrequirements:

1. The fan pressure curves should be stable throughout theentire operating range of the system (actual and testblock). Certain fans, such as most radials, the nybPressure Blower, and the nyb AcoustaFoil, are stablefrom wide-open to completely closed-off to offer thebroadest possible control range.

2. The fan and all its components should be designed tomeet even the test block condition without passingthrough the first critical frequency of the rotating parts.A common specification calls for the fan shaft’s firstcritical speed to be 125% of the maximum operatingspeed.

3. The entire fan assembly should be rugged to withstandindustrial service. Catalogs or drawings should containcomplete material specifications.

4. Whenever possible the entire fan, motor, and driveassembly should be factory assembled, aligned, andtest run to ensure smooth operation. The fanmanufacturer should be capable of test runningcomplete assemblies.

Induced-draft fans have further special requirements:

5. Where fan airstream temperature exceeds 300°F., thefan should include a shaft cooler and the bearing baseshould be separated from the fan housing.

6. The fan should be selected to handle the maximumparticulate loading. nyb offers radial, radial-tip, andbackwardly-inclined designs in a variety of alloys tohandle a wide range of contaminated airstreams.

7. Fans handling particulate-laden airstreams should befurnished with shaft seals to protect the inboardbearings. Ceramic-felt shaft seals usually provide thebest protection in these applications.

8. Fans handling particulate-laden airstreams should befurnished with a cleanout door and a drain to facilitateperiodic cleaning. Various quick-opening, bolted, andraised, bolted cleanout doors and drain connections aregenerally available.

9. Blade liners, housing liners, and hard surfacing ofblades and/or inlet cones may be desirable, dependingon the particulate loading.

CONCLUSION

The proper specification and selection of fans for combustionprocesses require a careful communication between the systemdesigner and the fan manufacturer. Given a clear understandingof the specification, the fan manufacturer can offer theappropriate fan type and accessories for the application.

Form 607 GAW


SELECTION CRITERIA FOR FAN DAMPERSINTRODUCTIONDampers are the most common volume controldevice used in fan systems. Low in cost,dampers require little maintenance, easily adjustairflow during operation, and need little space.For these reasons, they are often selected overmore complex control systems such as variablefrequency drives.To select the best damper for a particularapplication, it is necessary to understand therequirements of the application as well as thecapabilities of different damper systems.Since dampers may be placed on either side ofthe fan, they are classified as either inlet oroutlet. Both reduce airflow in predictableamounts, but by different means.Outlet dampers control the air after it has passedthrough the fan by changing the resistance thefan is working against. Figure 1 shows theeffects of various outlet damper settings on abackwardly-inclined fan. It illustrates how thedamper controls CFM, static pressure, and itsimpact on fan BHP.As the outlet damper is closed, the point ofoperation moves to the left of the selection pointalong the fan’s static pressure curve. Addingresistance with the outlet damper also moves thefan horsepower to the left on its curve. Withradial-blade and forward curved-fans, thedampered horsepower will be less than the wideopen horsepower as the fan moves to the left onthe BHP curves. With backwardly inclined fans,the dampered horsepower may be less, the same,or more than its wide open horsepower,depending on the original point of operation. Formore information see Engineering Letter 3.Inlet dampers affect the air before it enters thefan. External, internal, or inlet box inlet damperscause the entering air to spin in the samedirection as the fan rotation. Because of this, thefan wheel can not develop full output. Thisresults in lower volume and reduced BHP.When a backwardly inclined fan has an inletdamper, it reacts as shown in Figure 2 as thedamper vane angle is changed. For each newdamper vane position, new SP and BHP curvesare generated. The new point of operation isdefined by the system in which the fan isinstalled. The end result is similar to the changethat occurs when slowing down an undamperedfan.

The horsepower and electrical power savings of this damper make it attractivefor systems required to operate at reduced flow rates for extended periods, suchas in variable-air-volume systems. While Figure 2 illustrates an inlet damper’seffects on a backwardly inclined fan, the same general results are achievedusing inlet dampers on any type of centrifugal fan.

Figure 1 - Static pressure and brake horsepower curves for backwardly-inclined fan with outletdamper. As the damper closes, the point of operation - brake horsepower and static pressure - movesto the left of the original fan selection point to the 90 -degrees (wide open) damper setting.

Figure 2 - Effect of applying inlet dampers to the fan in Figure 1. Separate SP and BHP curves aredeveloped for each vane setting. Fan operating points at these settings are determined by systemresistance (points where system curve intersects SP and BHP fan curves).

TYPES OF OUTLET DAMPERS

The parallel blade arrangement shown in Figure 3 is the simplest,most economical, and most popular type of outlet damper. Thecross-sectional area of a wide-open damper is not greatlyreduced until the blades have been moved to the 30 degree openposition. Consequently, the outlet damper control arm swingsthrough a relatively large arc to reduce fan capacity a smallamount. This makes the parallel-blade outlet damper particularlyuseful when installed on a continuous process system wheresensitive control of air volume between wide open and 70% or80% of wide-open is desired. The large control arm swing alsoallows predetermined settings of airflow to be repeatedaccurately. This damper, being the least expensive of thevarious designs, also makes it the usual selection for systemsthat require two position damper operation (either full-open orfull-closed). Another common application involves cold startson a “hot” system requiring a reduction in airflow to reduceBHP until the system reaches temperature.Opposed-blade outlet dampers, as pictured in Figure 4, are usedwhen a straight line relationship between fan volume and controlarm swing is desired. In this design, alternate blades turn inopposite directions. Therefore, the change in volume, with respectto the damper position, is proportional to control arm swing.

The opposed-blade damper is usually selected when it isnecessary to maintain an even distribution of air immediatelydownstream from the damper. Figure 5 illustrates thedownstream air pattern of an opposed-blade versus a parallel-blade damper. Opposed-blade dampers cost more than parallel-blade models of the same size due to the increased complexityof the linkage required to provide the opposed-blade motion.

TYPES OF INLET DAMPERS

Inlet dampers can provide a substantial horsepower savings forfans that are operated at reduced capacity for extended periodsof time. Concerns for energy conservation and reducedoperating expense make this feature desirable and oftenmandatory when designing a system.A good example of how inlet dampers are used to accomplishenergy savings can be seen in a typical variable volume heating-cooling ventilation system. In this application much less air isneeded for winter heating than for summer cooling. In addition,during summer operation, less air is needed for cooling duringthe nighttime hours than during the peak daytime hours. Yet, thefan system must be selected for the worst condition/highest airflow. The inlet damper offers the greatest long term savings inVAV applications due to reduced horsepower requirements atreduced volumes.

External inlet dampers, as shown in Figure 6, are mountedexternal of the fan structure. The configuration is circular withthe damper vanes connected to a central hub through pivotbearings. The control linkage is also circular and exposed foreasy inspection and maintenance.Generally, this is the most expensive damper configuration. It isalso capable of handling higher velocities and pressures than theinternal inlet damper.

Figure 3 - Parallel-Blade Outlet Damper

Figure 4 – Opposed - Blade Outlet Damper

Figure 5 - Airflow Patterns through Dampers

Figure 6 – External Inlet Damper

The internal inlet damper, pictured in Figure 7, is similar to theexternal inlet damper with respect to controlling fanperformance. The most significant difference is that the internaldamper is a self-contained unit furnished as an integral part ofthe fan inlet cone. This provides considerable space savings andeases installation. The internal inlet-damper design, however,may tend to create some resistance at wide-open, due to thecontrol vanes being in the high velocity region of the fan inlet.Therefore, appropriate airflow reduction factors, as listed in aseparate engineering supplement, must be used when sizing a fanwith this type of damper. In addition, the damper control linkageis in the airstream on the inside of the fan housing and must beserviced through a cleanout door in the housing.

Inlet-box dampers (Figure 8) are parallel-blade rectangulardampers mounted on an inlet box in such a way that the airflowfrom the damper produces a vortex at the fan inlet. Inlet-boxdampers are generally preferable on fans equipped with inletboxes and have the same general control requirements asstandard inlet dampers. Because the bearings are not in theairstream, inlet-box dampers are often used in airstreams thatcontain some particulate. Predicting the exact flow reductionwith damper angle varies with damper types and products.Normally this is not a requirement since flow should beestablished using manual reference or feedback from automaticcontrol systems. For all inlet-vane dampers, vane angle versusflow relationship will change when dampers are applied towheels that have been narrowed to establish specific capacitiesat direct drive speeds.

Inlet dampers typically improve the stability of most productsbecause they control the flow through the fan inlet. At extremedampering, about 30° open, inlet dampers can no longer create avortex and become essentially a blocking damper. This causes thefan to operate far to the left on its curve. When this happens, a fanis subject to the same problems of instability as if the point ofrating was established by an outlet damper or other systemchanges.

COMBINED INLET AND OUTLET DAMPERS

Occasionally it is desirable to save more power at reducedcapacity while maintaining very sensitive control. In this case,the fan may be equipped with both inlet and parallel-blade outletdampers. With the outlet damper set at wide-open, the inletdamper is set to give just slightly more air than needed. Exactair delivery is obtained by adjusting the outlet damper. Becausethe outlet damper vanes require a lot of movement to achieve aslight change in air delivery, sensitive control is achieved.

PERFORMANCE COMPARISON

Figure 9 shows the effects of damper settings on airflow andbrake horsepower for parallel and opposed-blade outletdampers, and inlet and inlet-box dampers. These plots representgeneralizations of damper effect on fan performance and can beused to compare one type to another.

Figure 9

Effect of vane setting on airflow and power for various damper types. When a parallel-blade outlet damper is set for 80 percent ofwide-open capacity, the damper setting is 40 degrees, and the fan operates at 85 percent of wide-open horsepower. However, with aninlet damper, operation at 80 percent of wide-open requires a 53 degree damper setting and 72 percent of wide-open horse-power.Note: These curves are representative, not precise. See text.

Figure 8 – Inlet - Box Damper Mounted To Inlet BoxFigure 7 – Internal Inlet Damper

Parallel-BladeOutlet Damper

Opposed-BladeOutlet Damper

External andInternal

Inlet Dampers

Inlet-BoxDamper

1. Cost Least costly. 1.1 to 1.2 times as muchas parallel blade.

Internal - 1.5 to 2.5 timesas much as parallel-blade.External - 3 to 4 times asmuch as parallel-blade.

1.3 to 1.4 times as muchas parallel-blade;combined with inlet box3 to 4 times as much asparallel-blade.

2. Control Best for full-open orclosed requirements orfor fine control between80% to 100% full-flow.

Best for systems whereair volume is changedover a wide range and astraight line relationshipof volume to control armswing is desired.

Same as opposed-bladeoutlet damper.

Used on fan inlet box.Can be used with someparticulate in airstream.

3. Horsepower Depends upon characteristic BHP curve;Backwardly inclined - same, more, or less thanwide-open, FC and Radial – less than wide-open.

Power consumption atreduced air volumes isless than with outletdampers.

Same as inlet damper

4. Air flow after fan Throws air to one side. Distributes air evenly. No effect. No effect.

Figure 10 - Comparison of Inlet and Outlet Dampers

SUMMARY

Each system has its own requirements with respect to thecontrol of air volume. System designers must be aware of notonly first cost considerations but, more importantly, of the longterm operating savings that can be achieved by a properlyengineered system. Each system also imposes limits on which

dampers can be used with respect to fumes, control sensitivity,and temperature. No one damper design is best for allapplications. Figure 10 provides a comparison to help thedesigner recognize some of the factors to be considered indamper selection.

Form 607 GAW


AN INTRODUCTION TO FAN ACOUSTICSINTRODUCTION

Fan Acoustics is an important consideration in the industrialenvironment and with commercial ventilation systems. Thesound generated by some fans can be a potential hazard topersonnel in close proximity to the fan, and the sound can betransmitted, via the ductwork connected to the fan, to all areasserviced by the fan. Because of these concerns, fanmanufacturers publish sound ratings for their products to serveas a guide for selecting fans to meet sound specifications, and toassist acoustical consultants in predicting the total noise levelsin specific environments. This Letter provides basic informationto help understand fan sound ratings and how to apply them.

WHY FANS MAKE NOISE

Like any mechanical device, fans generate sound, whichemanates naturally from the turbulence of moving air, themechanics of moving parts of the fan, and from vibration.

AIR TURBULENCE

Air turbulence within the fan increases the sound coming fromthe air movement. The noise resulting from air turbulence is amajor factor in the sound levels of a fan in a specificapplication. Further, duct work can transmit this turbulent noiseto all areas serviced by the fan.

Factors contributing to air turbulence include the resistance toflow, flow separation along fan surfaces, and shock related toabrupt changes in the direction of airflow, pressure, or velocity.The principal areas where such turbulence is encountered withina fan are shown in Figure 1.

A lower noise level can be achieved by reducing air turbulence.This can be done by considering several factors related to airmovement when selecting fans.

The first factor to consider is the fan’s blade pass frequency,which is a pure tone produced when the blades of the fan wheel(impeller) rotate past the housing cut-off sheet in centrifugalfans, or the turning vanes, in axial fans. The blade passfrequency is calculated by multiplying the number of bladestimes the rotating speed in revolutions per minute. If thisfrequency matches the natural frequency of the ductwork, it canexcite the ductwork, which can cause it to resonate, therebyincreasing the noise level.

Because of this possible increase in sound, and because certainpure tones are irritating to people, the sound output of the bladepass frequency should be investigated when sound reduction isdesired.

The next factor to consider is the fan design. Generally a fanoperating at peak mechanical efficiency will produce less noise,because high efficiencies result from minimal air turbulencewithin the fan.

There are four basic centrifugal fan wheel designs - forwardcurved, backwardly inclined, radial, and radial tip - and a varietyof axial flow wheel designs (see Figure 2). Each wheel designhas unique sound characteristics due to the way they handle air,and the efficiencies they can achieve.

Fan speed does not always determine which fan will be quieter.For example, centrifugal fans have higher amplitudes at lowerfrequencies, while axial fans exhibit higher amplitudes at thehigher frequencies.

The amplitude of the blade pass frequency on an axial fan ishigher and more pronounced than on backwardly-inclined fans,and commonly will have amplitude peaks at multiples of thisfrequency.

Figure 1 - Typical Areas of Turbulence

Of the four centrifugal designs, the backwardly-inclined fansare the most efficient, and therefore, the quietest. Those withairfoil-shaped blades offer the highest efficiencies, for cleanair environments, while those with single-thickness blades canbe used in applications where light dust or moisture is present,although the efficiencies are somewhat lower.

Certain types of axial fans offer the next highest efficiencies.An excellent example is the nyb Vaneaxial fan that uses airfoilshaped blades in an in-line flow design. This fan is used tohandle high volumes of clean air at low pressures, which is atypical ventilation application.

Radial fans are typically low efficiency, open designs forspecial purpose applications, such as bulk material handling,or exhausting/supplying lower volumes of air at higherpressures. An exception to this is the nyb DH design (Figure3), which has superior efficiencies for a radial wheel andrelatively low sound levels. A radial fan will be much louderthan a backwardly-inclined fan operating under the samevolume and pressure conditions.

Radial-tip fans, commonly used to handle larger volumes ofair that contains particles or material, exhibit soundcharacteristics similar to the radial fans.

The sound spectra of radial and radial-tip fans containamplitude spikes at various frequencies, and a noticeable spikeat the blade pass frequency.

The forward-curved fan design operates at speeds that aremuch slower than the other fan types, which results in lowernoise levels from mechanical operation and vibration.However, because of its modest efficiencies, a forward-curvedfan may be noisier than a backwardly-inclined fan whenoperating at comparable volume and pressure. The soundspectrum of the forward-curved fan shows a slower rate ofreduction in amplitudes than the other centrifugal types, andbecause of the large number of blades, the blade passfrequency occurs much later in the spectrum and is notpredominant.

MECHANICAL NOISEThe moving components of the fan - the motor, bearings, anddrive - produce sound. This too can be transmitted through thesystem via the fan structure or shaft, or when these

components are in the airstream. Motor sound will vary withspeed, enclosure, electrical characteristics, and even the manu-facturer. Antifriction bearings can be used to reduce bearingnoise, and proper drive selection will reduce the likelihood ofbelt hop, or slap. Of course, proper maintenance must beemployed to keep the moving parts running smoothly, andquietly.

VIBRATION

Excessive vibration can significantly add to the overall noiselevel of an installation. This will occur if the fan or any of itscomponents are not adequately balanced, if the fan is installedon an insufficient foundation, or if the fan is not properlyisolated from other system components. For example, it is notuncommon for the fan’s support structure or ductwork to have anatural frequency at the fan’s operating speed or blade passfrequency, either of which can cause the system to resonate atthat frequency, increasing the sound levels, and the possibilityof damaging the installation. These risks can be eliminated bychanging the speed of the fan, installing appropriate isolation,and/or detuning of the fan or affected system components.

NOISE MEASUREMENT

Overall noise levels can be measured at any installation using avariety of portable sound level meters, or more sophisticatedequipment like a frequency analyzer (Figure 4).

Figure 2 – Various Wheel Types

Figure 3 – DH Wheel

Sound Pressure (Lp), is an atmospheric pressure change that isaudible to the human ear, and is measured from a point inspace where the microphone or listening device is located. Thehuman ear can perceive a broad range of sound pressures,from the threshold of hearing (2 x 10 - 7 microbar) to thethreshold of pain (1 microbar). The threshold of pain is fivemillion times louder than the threshold of hearing. The decibelis used in acoustical work to indicate sound pressure levelsbecause it condenses this tremendous range of values to aworkable range of from 10 dB to 130 dB. A decibel (dB) is alogarithmic ratio of some measured value to some referencevalue. It is standard international practice to use the soundpressure at the threshold of hearing as the reference value forthe sound pressure level scale.

Figure 5 shows the relationship between the sound pressuremeasured in microbar, and the sound pressure levels measuredin decibels.

Often a single sound pressure value is used to represent thetotal sound spectrum. This is expressed as dBA, indicating thatthe sound pressure, in decibels, has been adjusted to reflect asingle number value for a sound pressure, weighted by the “A”scale. The “A” scale weighting reduces the effect of lowerfrequencies, with the intent to establish a value moreproportional to the human ear frequency response. dBA isused by OSHA to set maximum allowable noise levels, pre-scribing a maximum dBA limit for an 8 hour exposure. dBAcan be measured with a sound level meter, or calculated byapplying the weighted values to the eight octave bandsencompassing the range of hearing.

Better definition of sound pressure levels is gained by breakingthe sound spectrum into discreet ranges. The standard practiceis to divide the audio spectrum into eight octave bandsidentified by the center frequency of each band. Figure 6shows the octave bands of the audio spectrum as defined bythe American National Standards Institute (ANSI) Standard S1.6, series 2.

Series 2ANSI S1.6

From(Hz)

To(Hz)

CenterFrequency (Hz)

BandNumber

45 90 63 190 180 125 2

180 355 250 3355 710 500 4710 1400 1000 5

1400 2800 2000 62800 5600 4000 75600 11200 8000 8

Figure 6

Fan manufacturers generally test and rate fan noise according toAir Movement and Control Association (AMCA) Publication300 - Test Code for Sound Rating Air Moving Devices, andPublication 301 - Methods for Calculating Fan Sound Ratingsfrom Laboratory Test Data. This testing procedure requires areverberant or semi-reverberant room with a calibratedreference sound source to determine the room characteristics,and is known as the substitution method.

Sound data is acquired in the octave bands shown in Figure 6.

The measured sound pressure of a test fan is mathematicallyconverted to a sound power level using predeterminedmicrophone locations.

MEASURING FAN NOISE

dBA is a useful measurement for evaluating the overall noiselevel at a particular location, but this measurement takes intoaccount all of the sound sources affecting that particularlocation, which include the sounds from all equipment in thearea, natural sounds of the environment, and from otherenvironmental factors. Some of these factors are the currentphysical properties of the air such as temperature, humidity,and pressure, whether the location is outside or inside, the sizeand material of the room. All of these affect the sound

Figure 4 – Frequency Analyzer

Figure 5 – Sound Pressure Measurements

pressure experienced by the listener, and recorded by thesound level meter. Because of this, it is impossible for the fanmanufacturer to guarantee sound pressure levels or dBAvalues.

For several years fan manufacturer,s and other makers ofindustrial equipment, have used Sound Power (Lw) values totest and rate fans. Sound power has been chosen because it isindependent of the acoustical environment in which the fan isinstalled. It is the only value that is specific to the particularfan.

Sound power is the total energy emitted from a fan which is afunction of the fan’s speed and point of operation, and isindependent of the fan’s installation and surroundingenvironment. A sound power level is the acoustical powerexpressed in dB radiating from a sound source. It is defined as:

10 log (Watts)Sound Power (Lw) =(10 -12)

Sound power levels can be converted into predictable soundpressure levels once the acoustical environment surroundingthe fan is defined.

Sound pressure for a given fan changes with a change in airvolume, pressure, or efficiency. Because of this, fans must betested at several speeds and efficiency points. After a fan’ssound power level has been determined at different speeds andpoints of operation, it is important to remember that theselevels will always be the same unless the fan is physicallyaltered. If a fan line is geometrically proportional, the soundpower for other fan sizes can be accurately projected from thebase fan. AMCA Publication 301 defines methods for acquir-ing such data.

FAN SOUND RATINGS

The sample table shown in Figure 7 shows a listing of totalsound power for a particular fan size and type at several speedsin each octave band. Sound power ratings can also bepresented graphically.

Octave BandsFanRPM 1 2 3 4 5 6 7 81100 73 72 68 62 59 58 51 451300 79 74 75 67 63 62 57 501500 85 77 78 71 67 66 62 541700 90 80 81 75 70 69 66 581900 93 83 83 78 73 71 69 612200 96 86 86 82 76 74 72 662600 99 90 86 88 80 77 75 703000 101 93 88 90 83 79 78 743400 105 98 90 94 87 82 81 783800 107 102 94 96 90 85 83 814200 109 106 97 98 93 88 86 844600 111 109 99 99 97 91 88 865000 113 112 102 100 101 93 90 88

Figure 7

Total sound power can be broken up to inlet sound power andoutlet sound power. For all functional purposes, the soundpower that is radiated from the inlet and outlet of a fan is equalto each other. Because a fan manufacturer can present itssound information in the form of inlet, outlet, and total soundpower, it is important to clarify the identity of the rating beforeany comparisons and calculations are made.

In general, a fan manufacturers’ sound ratings are at peak pointof efficiency as shown in Figure 8. As stated earlier, fanefficiency and air turbulence contribute to changes in noiselevels. Consequently, if a fan is operating at a point ofoperation outside its maximum efficiency range, the user willhave to correct the manufacturers’ sound ratings as shown inthe table on page 5.

Figure 8 shows a fan's point of operation at the intersection ofthe static pressure and volume range on the curve. Since airvolume can be defined by a velocity or velocity pressurethrough the fan’s outlet area, the fan’s point of operation canbe defined by stating the ratio of velocity pressure to staticpressure, or VP/SP. By using a chart such as the one shown inFigure 9, the user can make the necessary sound correctionsfor fan operations outside the maximum efficiency range.

Published fan sound power ratings and corrections onlyreflect noise created by air turbulence within the fan.Because of the infinite variables, mechanical noise andvibration noise are impossible to accurately predict, andare not included in the rating.

Another rating method is described in AMCA Publication 302 -Application of Sone Ratings for Non-Ducted Air MovingDevices. A Sone is a ratio of loudness between two sounds.The Sone scale is linear, ranging from soft to loud. Unlike thedecibel, two Sones are twice as loud as one Sone. This methodwill produce reasonably accurate estimates of sound pressurein a free-field condition, and is used by manufacturers of roofventilators and other non-ducted commercial ventilationproducts, but is not suitable for analytical purposes.

Figure 8

Octave BandsFanSpeed

VP/SP Point ofFan Operation 1 2 3 4 5 6 7 8

0 to .03 Peak SP 5 3 0 1 1 0 -3 -1.03 to .10 Peak ME 0 0 0 0 0 0 0 0.10 to .30 1/2 Peak SP 4 2 0 0 2 1 2 2

up to2500

.30 and up Near Wide Open 2 2 2 1 3 3 2 30 to .03 Peak SP 3 4 5 4 0 0 0 -2

.03 to .10 Peak ME 0 0 0 0 0 0 0 0

.10 to .30 1/2 Peak SP 4 2 0 3 3 3 2 1over2500

.30 and up Near Wide Open 3 3 1 3 4 5 3 4Figure 9 - Typical dB Corrections for Point of Operation

APPLYING SOUND POWER

When the sound power for a fan has been calculated at a fixedspeed and known point of operation, the sound pressure can beestimated. It should be remembered that sound pressure or dBApredictions are only estimates based on certain knownconditions or assumptions regarding the location of the fan andthe physical installation.

The Short Form for Sound Calculations shown on page 8 is oneway to calculate sound pressure. This is a step-by-step methodfor estimating sound pressure levels or dBA for a specificinstallation.

The short form only applies to outdoor installations or to indoorinstallations where the listener is relatively close to the fan andthe room is relatively large. Such installations may be termed“free field.” Even given these assumptions, reflecting surfaces,inadequate support structures, high-loss ductwork, or flexibleduct connections could seriously alter the outcome.

For example, the fan corresponding to Figures 7 and 9 might berequired to operate at 1500 RPM:

Octave Band 1 2 3 4 5 6 7 8

Center Frequency 63 125 250 500 1000 2000 4000 8000

1. Fan Total SoundPower @1500 RPM 85 77 78 71 67 66 62 54

2. VP/SP Correction 5 3 0 1 1 0 -3 -1

3. Fan Sound Power(1) + (2) 90 80 78 72 68 66 59 53

4. Correction for Insta-llation (inlet or outlet) -3 -3 -3 -3 -3 -3 -3 -3

5. Corrected SoundPower at Fan (3) + (4) 87 77 75 69 65 63 56 50

6. End ReflectionValues 14.5 9.0 4.5 1.5 0 0 0 0

7. Corrected SoundPower (5) - (6) 72.5 68 70.5 67.5 65 63 56 50

8. Conversion forSound Pressure, Q=2 20 20 20 20 20 20 20 20

9. Sound Pressureat 15 feet 52.5 48 50.5 47.5 45 43 36 30

Line 1 - enter the published sound power for each octave bandcorresponding to the required speed.

Line 2 - enter the appropriate VP/SP correction. For thisexample, assume VP/SP = .025.

Line 3 - enter the algebraic sum of lines 1 and 2.

Line 4 - enter the appropriate correction for the type of faninstallation. If neither the inlet nor outlet are ducted,no correction is necessary. If either the inlet or outlet isducted away from the listening location deduct 3 dB.This 3 dB reduction accounts for the assumption thatthe amplitude of inlet and outlet noise is approximatelyequal and half the noise is ducted away. Figure 10provides a graphic depiction of the effects of adding orsubtracting noises of similar or like amplitude.

If the inlet and outlet are both ducted away from thelistening location, only the sound power radiatedthrough the fan housing will remain. The appropriatereduction will vary from one fan to another dependingupon the specific housing thickness and reinforcementsand their attendant transmission loss. Refer to themanufacturers’ rating tables for the appropriatereduction for a specific fan type.

For this example, assume only the outlet noise isducted to the listening location.

Line 5 - enter the algebraic sum of lines 3 and 4.Line 6 - End reflection is a phenomenon that takes place when a

sound wave reaches an abrupt expansion such as theend of an open duct. At this point some of the soundwaves are actually reflected back into the duct so thatthe resultant sound power level is reduced. The effectsare more pronounced in lower frequency ranges and insmaller duct diameters as shown in Chart III, page 8.For applications where noise level emitted from theinlet or outlet duct concerns the listening location, theduct diameter must be determined and the appropriatevalues subtracted from the fan sound power.

For this example, assume only outlet ducted noise isavailable at the listening location and the duct is 15" indiameter. (See Chart III on page 8.)

Line 7 - enter the difference between lines 5 and 6.Line 8 - enter the correction for directivity and distance.

Figure 10

As mentioned previously, the amplitude of a noiselevel will vary depending upon the installations andthe distance between the source and the listeninglocation. The number of reflecting surfaces alsodetermines the sound wave radiation pattern. Thesepatterns are known as directivity factors (Q) andindicate the type of radiation from the number ofreflecting surfaces.

AMCA Publication 303 - Application of SoundPower Level Ratings describes Q = 1 as havingspherical radiation with no reflecting surfaces. Anexample would be an axial fan located in a stack. Q =2 is used for hemispherical radiation where onereflecting surface is present such as a fan on the floorin the middle of a room. For each additionalreflecting surface, the directivity factor is doubled.For example, a fan mounted on the floor directlyadjacent to a wall would have a Q = 4 factor.

The appropriate directivity factor must be used inconjunction with the distance from the noise sourceto the listening location to obtain the reduction factor(Lw - Lp) to convert sound power to an estimatedsound pressure. Using Chart I on page 8, thelistening distance from the source must be plotted onthe bottom horizontal graph and a vertical line shouldbe drawn at that point. A horizontal line drawn fromthis vertical line at its intersection with theappropriate directivity line will indicate the (Lw - Lp)reduction.

These estimates apply to a listener’s position fromthe noise source and do not consider outsideinfluences from other machinery or unpredictableobstructions, but produce reasonably accurateestimates of sound pressure in a free field condition oroutside installations.

For this example, assume a Q = 2 directivity factor ata distance of 15 feet. (See Chart I on page 8.)

Line 9 - deduct line 8 from line 7 and enter the result.

The sound power levels represent the final estimatebased on all the stated conditions. The one remainingstep is to determine the proper dBA value.

The dBA value is the sound pressure level corrected tothe “A” weighting network. This is accomplished bydeducting the proper “A” weighting value from eachof the eight octave bands, then using the graph fromFigure 10 to combine the results to obtain a singlenumber dBA value that represents the fan and itsparticular installation. Because decibels arelogarithmic values, simple addition cannot be used.

A simpler method of approximating dBA values canbe found on Chart II on page 8. Using the scale onthe left hand side of the graph, plot the sound pres-sure levels from line 9 directly on to the graph foreach octave band. Then the maximum dBA can bederived by finding the band number (centerfrequency) that exceeds the highest octave band levelby the most decibels. In our example, band number 5(1000Hz) exceeded the octave band level 40 dBA by8 dB. This was greater than any other band number.Therefore, the dBA level for this fan would beapproximately 48 dBA at 15 feet based on a Q-2directivity.

Another method is to combine decibels such that alogarithmic addition can be employed in lieu of thetabular method shown in Chart II. Logarithmicaddition involves calculating the antilog of eachdecibel to be added, summing the antilogs, findingthe logarithmic sum, and multiplying by 10. Thismethod and the formula are given in AMCAPublication 303.

TROUBLESHOOTING

To avoid undesirable noise levels in the final installation, thesystem designer needs to consider many factors. First, anacceptable noise level criteria must be established, based onthe activity in the area, the nature of the noise, the relationshipof the listening location, noise-criterion curves, and the OSHApermissible noise exposure regulations.

Properly selecting a fan type and operating it at peakmechanical efficiency will assure the quietest possibleoperation. It is not always possible to select a fan that doesnot exhibit a predominant blade pass frequency, but anawareness of this will help in selecting acoustical attenuationwhen necessary.

OctaveBand

SoundPressure

FromLine 9

CorrectionFor“A”

WeightedNetwork

dBAValue

byOctaveBand

Diff.

FactorFromIII.#10

Factor+

HigherValue

Diff.

FactorFrom

III.#10

Factor+

HigherValue

Diff.

FactorFromIII.#10

SingleNumber

dBAValue

1 52.5 -25.5 27.52 48 -15.5 32.5

-5.0 1.2 33.7

3 50.5 -8.5 424 47.5 -3.0 44.5

-2.5 1.9 46.4-2.7 .2 46.6

5 45 0 456 43 +1.0 44

-1.0 2.5 47.5

7 36 +1.0 378 30 -1.0 29

-8.0 .7 37.7-7.8 .7 48.2

-1.6 2.3 50.5

Location of the fan with respect to the listener is veryimportant. The greater the distance, the lower the noise level.The use of absorptive and reflective materials as well asisolation usually control excessive noise.

If the final installation seems excessively noisy, an octaveband sound analyzer should be employed to measure the noiselevel. Because it analyzes the spectrum by octaves, it ishelpful in isolating components within the spectrum that aremajor contributors to the noise problem.

Often, the fan is not the major source of the noise; many timesit is nearby machinery or the surrounding environment that islouder than the fan. After identifying the noise source, itsreduction can be approached from two directions:

1. Reduce the noise at the source.2. Reduce the noise at the listening location.

The first approach is usually the most cost effective. Toreduce fan acoustical noise, a reduction in sound energy isimportant. Lining ductwork with sound absorbing material oradding duct silencers will reduce airborne noise within theduct system. Flexible connectors between the fan inlet, outlet,and connecting ductwork will aid in reducing both vibrationnoise and mechanical noise that may be transmitted throughthe entire system.

Fan noise produced by vibratory forces can be induced by anumber of components. Sometimes the source is easilydetected from experience and at other times measuringinstruments are required. The solution to vibratory noise willdepend on where it occurs. Reducing the amount of thevibration, eliminating it by substitution, isolating it, orchanging the frequency are all possible solutions.

For example, unbalance is a chief cause of vibratory noise.Consequently, balancing the rotor will reduce the vibrationcaused by imbalance. Replacing a noisy bearing or drivecomponent will eliminate the source. Installing rubber orspring isolators will prevent transmission of the noise to themounting structure. Detuning natural frequencies of a structureby changing the fan speed or the natural frequency mayeliminate this problem.

Using the second approach, the noise level at the listeninglocation can be reduced by increasing the distance of the soundpath. This can be accomplished by moving either the fan or thelistener or by rotating the fan so that the noise is directed awayfrom the listener. Changing the characteristics of the room byadding sound absorbing material will help reduce noiseHowever, the effectiveness of sound absorbing material dropsoff rapidly at frequencies below 250 Hz.; consequently, thisapproach is somewhat limited. Enclosing the fan in a soundabsorbing room, for example, will aid in reducing noisetransmitted from the fan structure but will do nothing aboutnoise within the duct system. Erecting sound barriers oremploying some type of ear protection are also alternativesolutions.

These troubleshooting tips only cover a few possiblealternatives. Volumes of reference material are available onthe subject, and acoustic consultants are available to assist in theareas of noise abatement and acoustical control. Fanmanufacturers can provide assistance in resolving noise issuesrelated to the specific fan but normally do not perform overallacoustical engineering consulting.

SHORT FORM FOR SOUND CALCULATIONS

This form is to be used for the approximate sound pressure level calculation of a fan, assuming that the listener’s position is in thedominant free field. In most cases this can be considered no more than 5 feet in an enclosed room, or an outside installation freefrom reflecting surfaces.

OCTAVE BANDS 1 2 3 4 5 6 7 8CENTER FREQUENCIES 63 125 250 500 1000 2000 4000 8000

1. Fan Sound Power Rating at __________RPM2. VP/SP Correction3. Fan Sound Power (1) + (2)4. Correction for Installation (Inlet, Outlet)5. Corrected Sound Power at Fan (3) + (4)6. End Reflection Value (Chart III)7. Corrected Sound Power (5) - (6)8. Conversion to Sound Pressure (Chart I)9. Sound Pressure at ___________ ft. (7) - (8)

The estimated dBA value is _______ at _______ ft. (Chart II)

CHART IDIRECTIVITY/DISTANCE REDUCTION

[Given directivity and distance, Sound Power is converted toSound Pressure.]Q-1 UNIFORM SPHERICAL RADIATION with no

reflecting surface. Example: Stack discharge.Q-2 UNIFORM HEMISPHERICAL RADIATION with

one reflecting surface. Example: Floor mounted fan.Q-4 UNIFORM RADIATION over 1/4 SPHERE with two

reflecting surfaces. Example: Fan mounted on floornear interior wall.

CHART IIIEND REFLECTION VALUES (Decibels)

Octave Band 1 2 3 4 5 6 7 8Hz 63 125 250 500 1000 2000 4000 8000

5 23.5 17.5 12.0 7.0 2.5 .5 -- --10 17.5 12.0 7.0 3.0 1.0 -- -- --15 14.5 9.0 4.5 1.5 -- -- -- --20 12.0 7.0 3.0 1.0 -- -- -- --30 9.0 4.5 1.5 .5 -- -- -- --

Duct

Diameter

Inches

40 6.5 2.5 1.0 -- -- -- -- --Form 607 GAW

CHART IISOUND PRESSURE TO DBA CONVERSION

ENGINEERING LETTER 13The New York Blower Company ●7660 Quincy Street, Willowbrook, Illinois 60527 - 5530

F A N B A L A N C E A N D V I B R A T I O NINTRODUCTION

Vibration always has been a good indicator of how well a pieceof equipment was designed, installed, and maintained. Withsophisticated, computerized, preventative maintenanceprograms, vibration can now also be used as a precursor offuture maintenance requirements.

Fans are subject to vibration because they have a high ratio ofrotating mass to total mass and operate at relatively highspeeds. Unlike most mechanical equipment, there are twomajor causes for vibration in fan equipment. These areaerodynamic, having to do with airflow, and mechanical,having to do with rotating components, fasteners, and structuralsupport. This Engineering Letter will discuss both causes ofvibration and provide guidelines for their reduction.

AERODYNAMIC VIBRATION

Aerodynamic vibration, also referred to as aerodynamicpulsation, is one cause of fan-system vibration. It occurs whena fan operates to the left of its peak static pressure point. Thevibration frequency, when checked with instruments, is at afrequency other than the wheel rotation speed.

This area of operation is illustrated in Figure 1. In this regionthe fan wheel does not move enough air to fill the blade passages.Aerodynamic vibration is most easily identified by increasingthe volume of air flowing through the fan, thereby moving thefan’s point of operation to the right. If the cause is aerodynamic,the vibration will usually disappear or be reduced significantly.Increasing the airflow is accomplished by opening dampers,cleaning filters and coils, or as a test, removing a section ofduct near the fan. These actions will reduce system pressureand, correspondingly, increase the airflow.

AcoustaFoil™ is a trademark of The New York Blower Company

Because of their inherent wheel geometry, some fans are moresusceptible to pulsation when operating to the left of the peakon their static pressure curve. Centrifugal fans utilizing forward-curved or flat, backwardly-inclined blades are particularlysubject to this phenomenon. However, fans with backwardly-inclined airfoil blades, such as the AcoustaFoil™ wheel, aredesigned to be stable left-of-peak. Figure 2 illustrates this areaof unstable operation in a typical fan performance table (cross-hatch area). These points of operation indicate fan instability.

Operation left-of-peak may be due to an error in systempressure calculations, less than optimal system installation, orpoor maintenance practice. The fan’s point of operation mayhave also changed because the process/system has beenmodified since installation. For example, a drying system mayhave initially been designed to pull air through a 2" bed ofmaterial. Subsequent system changes now require a 6" bed ofmaterial with a significantly higher pressure drop. This willcause the fan to operate at a different point on its curve whichmay be left-of-peak.

Refer to Engineering Letter 7 to better understand how to takesystem measurements to determine a fan’s point of operation.

If it is determined that the vibration is aerodynamic, there areseveral steps that can be taken to restore the fan to anacceptable operating point. If some type of blockage is causingthe problem, dampers can be opened, filters and coils cleaned,and the process can be restored to a configuration more closelyresembling the initial design. More expensive alternativesinclude increasing duct sizes, reducing duct lengths, andeliminating abrupt turns.

1” SP 2” SP 3” SP 4” SPCFM OV

RPM BHP RPM BHP RPM BHP RPM BHP1240 800 1207 0.17 1516 .032 -- -- -- --1550 1000 1355 0.23 1620 0.41 -- -- -- --1860 1200 1517 0.32 1757 0.52 2178 0.97 -- --

2170 1400 1690 0.42 1904 0.65 2288 1.15 2633 1.702480 1600 1867 0.56 2065 0.81 2415 1.35 2732 1.943100 2000 2239 0.91 2405 1.22 2708 1.87 2983 2.553720 2400 2620 1.40 2765 1.78 3032 2.54 3276 3.32

4340 2800 3007 2.06 3138 2.51 3378 3.40 3600 4.294960 3200 3401 2.92 3518 3.44 3736 4..45 3939 5.475580 3600 3797 4.01 3902 4.58 4104 5.74 4286 6.856200 4000 4196 5.35 4292 5.99 4476 7.26 4647 8.52

Figure 2 – Typical Fan Performance TableCross-Hatch Indicates Area of Instability

Figure 1 – Typical Fan Static Pressure CurveCross-Hatch Indicates Areas of Instability

Figure 3 – Induced Air Recirculation

If a redesign of the system is not practical but current airvolume is adequate and the fan in question is a centrifugal, itmay be possible to eliminate or reduce pulsation by adjustingthe fan wheel toward the inlet cone. As shown in Figure 3, byadjusting the wheel so the edge of the cone is inside the wheelfront plate, additional air will recirculate in the fan. The fanwheel will now receive a sufficient volume of air, allowing itto perform without pulsating; however, the efficiency of thefan will be reduced. In general, increasing the overlap by adistance equal to 2% of the wheel diameter will eliminatepulsation.

Aerodynamic vibration may also be caused by poor inletconnections to the fan. Inlet boxes and inlet elbows should bevaned to reduce losses. When air is forced to flow through asharp turn as it enters the fan, it tends to load just a portion ofthe fan wheel. The result is always decreased performance butmany times pulsation as well.

The same phenomenon can also develop, though generally to alesser degree, at the discharge of the fan. Fans do not dischargeair at an even velocity across their entire outlet. They generallyoperate best when the air is discharged into a long, straightduct, the minimum being three duct diameters beyond the outletof the fan.

MECHANICAL VIBRATION

Mechanical vibration is the most common type of fan vibration.It is caused by unbalanced wheels or other rotating fancomponents. Its negative impact is increased with loosefasteners and poor structural support. Two terms are importantin understanding mechanical vibration.

Balance primarily refers to the fan wheel or other rotatingcomponents. The procedure of balancing involves adding orremoving weight in an attempt to move the center of gravitytoward the axis of rotation.

Vibration primarily refers to the complete fan. Fan vibration ismeasured during a “run test” and is the vibration amplitude atthe fan bearings expressed in units of displacement or velocity.The vibration level for new fan equipment is a result of thedesign and construction by the fan manufacturer. For operatingfan equipment, the installation and subsequent maintenancepractices can have a major effect on fan vibration.

There are a number of causes for wheel unbalance:

Construction - in new fan wheels unbalance exists because ofthe nature of the fabrication and assembly process. Part andassembly tolerances, material density variations, and warpageduring welding all contribute to non-concentric wheelassembly. Balancing compensates for these factors.

Material build-up - even a thin layer of dirt can cause asurprising amount of wheel unbalance. Using solvent, wirebrushes, scrapers, etc., wheels can typically be cleaned andrestored to a balanced condition.

Abrasion/corrosion - in material conveying applications orapplications handling corrosive fumes, abrasion or corrosion ofthe wheel will cause unbalance. For safety reasons, thiscondition is more serious than simple vibration and the fanmanufacturer’s representative should be contacted for repairrecommendations, up to and including wheel replacement.

Drive components - sheaves, belts, couplings, and motors canhave their own unbalance resulting in fan vibration. Checkcomponents for alignment, examine the grooves of sheaves,and check the surfaces of belts. Replace worn components.Couplings can shift even a few thousandths of an inch inshipment, causing misalignment and vibration.

Several drive components can be easily checked to determine ifthey are the cause of vibration. Disconnect the drive orcoupling and run the motor with one sheave or half-coupling inplace. If this assembly runs rough, remove the sheave or half-coupling and run the motor alone.

It is much more difficult to determine if the fan wheel or thedriven sheave/coupling is causing the vibration withoutremoving it and sending it to a balancing facility. Sheaves andcouplings should have been dynamically balanced originally.Unless it is important to determine whether the wheel or drivecomponent is out of balance, it is probably best to balance thewheel, shaft, and drive component as an assembly.

Fasteners -wheel and drive component setscrews, bearingbolts, and the fan base mounting hardware are all subject toloosening, especially when some vibration is present. Withoutattention, loose components will add to the overall fanvibration magnitude.

Structural support - too frequently, fans are mounted onsupports that have a natural vibration frequency near that of thefan. At this frequency, the structure will tend to continue tovibrate once it has been set in motion. Under such conditions it isalmost impossible to balance all of the rotating componentsfinely enough to prevent an objectionable amount of vibration.Adding mass or stiffeners will move the structure’s naturalfrequency out of the range of the operating fan.

Optimum mounting structures include thick concrete slabs,steel bases supported by isolators, or heavy, all-welded steelstructures. Structures must have adequate sway bracing, with nolong, unsupported spans. They should be designed to beheavier than if they were designed merely to support a staticload. All vertical supports should be directly underneath the fanand the fan should not be located in the middle of beam spans.

Bent shaft - can cause significant vibration which usuallyresults in a vibration magnitude that is proportional to theamount by which the shaft is bent. Using a simple dial indicator,the shaft can be checked for trueness. It should not be out morethan one or two thousandths of an inch on a short shaft or twoor three thousandths on a longer shaft. If the shaft is bent, it canstraightened, replaced, or compensated for trueness bybalancing.

BALANCE CRITERIA

Major fan manufacturers balance fan wheels prior to assemblyon precision balancing machines (see Figure 4). The balancingprocedure involves detection of and compensation for ounce-inches of unbalance.

For most HVAC, agricultural, and industrial applications, anISO balance quality grade of G6.3 is adequate. Using thisbalance grade, the permissible residual unbalance is calculatedas follows:

6.01 x G x WUper = N

Where:

Uper = permissible unbalance per balance qualitygrade (oz.-in.)

W = wheel weight (lbs.)

N = wheel operating speed (RPM).

G = balance quality grade (6.3)

For example, using a Size 264 Series 20 DH wheel:

Where:

W = 78 lbs.

N = 2280 RPM

G = 6.3

6.01 x 6.3 x 78Uper = 2280

Uper = 1.3 oz. – in.

VIBRATION CRITERIA

After wheel installation, assembled fans are “trim balanced” as acomplete unit before shipment (see Figure 5). Manufacturershave some limitations on what fans can be run tested based onelectrical requirements, test speeds, and customer furnishedcomponents.

To perform a vibration run test, the fan is mounted on a rigidbase. The base may be more or less rigid than that which thecustomer will use. Because of this difference, vibration limitsdetermined from the factory vibration run test cannot be used asa guarantee of the minimum level of vibration once the fan isinstalled in the system. To account for this difference invibration sensitive applications, more and more fans are beingmounted on vibration absorption bases. These bases containsprings or rubber-in-shear isolation and may or may not befilled with concrete for additional mass. The purpose of thesebases is to allow the fan to vibrate without transmitting thevibration to the building structure.

Figure 4 – Fan Wheel Balance Figure 5 – Fan Vibration Run Test

Fan assembly vibration is typically measured in the horizontaldirection with “filter in”. Filter in refers to the vibration beingmeasured only at the frequency of interest. This methodprovides an accurate measure of wheel unbalance. Transducerorientation may vary by product and/or test stand configurationat the discretion of the manufacturer (see Figure 6).

Major fan manufacturers have seismic vibration standards aspart of their manufacturing/quality procedures. These limitswill vary depending upon the fan manufacturer’s test facilities,balancing equipment, and fan type and size.

As a guideline for fans in HVAC, agriculture, and industrialapplications, a peak velocity of 0.15 inches/second at thefactory test speed is usually adequate.

For those more familiar with using displacement as a measureof vibration, displacement units can be converted to velocityunits using the following equation:

šx F x DV = 1000

Where:V = velocity (in./sec.)

F = frequency in revolutions per second(RPM/60)

D = displacement, peak-to-peak, (mils)(1 mil = .001 inch)

Example:

Convert .6 mils displacement to velocity in in./sec. with thefan running at 1200 RPM.

3.1416 x 1200 x .6V = 60 x 1000

V = .0377 in./sec.

CONCLUSION

System designers and specifiers should observe the followingspecifications to ensure minimum, acceptable levels of fanvibration:

1. Wheels should be dynamically balanced prior toinstallation in the fan assembly to ISO 1940/ANSIS2.19 Quality grade G-6.3.

2. Fans should be given a run test and “trim balance”after wheel installation at the fan manufacturer’splant to decrease vibration caused by other fancomponents and the overall assembly processwhenever the fan configuration permits it.

3. Mounting structures must be rigid and sufficientlyheavy to properly support the fan. Structures musthave a natural frequency that is well out of the fan’soperating range.

4. For vibration sensitive applications, specialconsideration should be given to spring or rubber-in-shear isolation, or inertia bases.

5. Utilizing computerized fan selection programs andthe manufacturer’s representative, fans should beselected to avoid unstable operating points andresulting aerodynamic pulsation.

6. Alterations to the overall system design shouldinclude consideration of changes in the fan’s point ofoperation and possible aerodynamic pulsation.

7. Proper maintenance practice, including periodicwheel inspec t io ns and inspec t ion of drivecomponents and fasteners, will assure reducedvibration levels.

Form 507 DJK

Figure 6 – Three Axis of Measurement


STAINLESS STEEL SPECIFICATIONS FOR FAN EQUIPMENTINTRODUCTION

Specifiers and users of air-moving equipment are often facedwith the presence of corrosive, abrasive, or high temperatureconditions which may be detrimental to the service life ofstandard mild steel fan equipment. Recognizing the limitlessvariety of stainless steel alloys or polyester resin-basedmaterials of which fan components can be fabricated, andconsidering the multitude of special purpose paints and coatingscurrently marketed for such applications, the specification andselection of the single best combination can be a difficult task.The purpose of this Engineering Letter is to provide somegeneral guidelines to assist in the process. Refer to EngineeringLetters 16 and 18 for similar guidelines on corrosion-resistantcoatings and fiberglass-reinforced plastic (FRP) fanconstruction.

STAINLESS STEEL ALTERNATIVES

Often, low first cost plays an important role in the selection of aparticular type of corrosion-resistant construction; specialtycoatings usually offer the lowest initial cost, followed bystainless steel alloy construction, and finally FRP construction.However, this method of selection does not take into account lifecycle costing that could result in the least expense over theservice life of the product.

Stainless steel and FRP are generally superior to specialty paintsor coatings when it comes to corrosion resistance. FRP willusually exhibit the best corrosion-resistant characteristics and willhandle certain corrosive agents or reagents that stainless steelwill not, and in some sizes is as economical as stainless alloys.However, stainless steel alloys are capable of highertemperatures and will stand up much better to the impact ofnon-abrasive materials. Also, fabrication methods tend to limitthe availability of FRP fan equipment and certain performancerequirements may force the consideration of stainless steel alloyconstruction as an alternate to the superior corrosion-resistantqualities of FRP.

Neither FRP fan construction nor special duty paints or coatingsapplied to mild steel construction will provide any measure ofprolonged service life in an abrasive application when comparedto mild steel. Even stainless steel alloys with their seemingly“tough” close textured surface finish provide negligible

improvement over mild steel in abrasive applications. Thereare, however, special alloy steels classified in the “abrasionresistant” or “AR” grouping. Such AR steels are usuallymade to a minimum 321 brinell hardness specification where,for example, 304 stainless steel is rated at only 124 to 147brinell hardness while 316 stainless steel is only slightlyharder.

TEMPERATURE CONSIDERATIONS

Typically, mild steel’s strength decreases rapidly at elevatedtemperatures, affecting the maximum safe operating speed ofthe fan wheel and consequently reducing the effectiveperformance range of the fan. Beyond 800°F., mild steel andeven 304 stainless steel are not well suited for rotating parts.At temperatures up to 1000°F., 316 stainless steel should beconsidered first because of its cost and availability. Onlywhere 316 stainless steel does not allow adequate speeds atthe required temperature should 347 stainless steel wheelconstruction be specified. Refer to each fan line’s bulletinfor speed derate factors.

In all cases, the suitability of a particular fan to operate atthe required temperature is solely dependent upon theindividual fan design and construction. Maximum safeoperating temperatures for fan equipment range to 1000°F.but are also dependent upon the proximity of motors orbearings to hot airstream surfaces. Only where the productliterature expressly acknowledges the suitability of the basicfan construction for operation at the required temperaturecan stainless steel construction be used to obtain the requiredsafe speeds.

STAINLESS STEEL TYPES

Assuming that stainless steel is needed for a specificapplication, the next step is to determine the best stainlessalloy to use.

There currently exist more than 100 registered grades ofstainless steel. Certainly, not all of these various alloys can bemade available for all of the different sizes and types of fanequipment. To facilitate selection, specification, andproduction, the availability of stainless steel alloys for fanequipment must be selectively limited.

Basically, stainless steel can be divided into three categories;Martensitics - 12% chromium and iron with carbon in balancedproportion. Ferritics - with higher chromium content andcarbon content held low. Austenitics - with nickel added. . .often referred to as 18-8 stainless which is approximately 18%chromium content and 8% nickel content.

Martensitics have the least tendency to work harden. Theapplication of this alloy grouping is usually limited to that ofprecision parts such as surgical instruments, shear blades, and dies.

Ferritics exhibit the greatest degree of corrosion resistance inthis grouping but work harden quite readily and are usuallylimited to decorative applications such as interior architecturaltrim, kitchen trim or utensils, and fasteners.

Austenitics provide the best combination of corrosionresistance and ductility. The suitability of these alloys forwelding and fabrication methods common to the fan industryreflect the standardization by fan manufacturers.

The Summary of Austenitic Stainless Steel Types on page 3presents alloy composition, strength characteristics, andtypical applications for the various stainless steel alloys in theAustenitic category.

Of the Austenitic alloys shown, some further limitations areplaced on the fan manufacturer due to material availability,inventory needs and costs, and specific production methods.

Refer to Engineering Letters 16 and 18 for condensed guides tothe corrosion-resistant characteristics of stainless steel alloys.Note that these are condensed references and do not presentthe full extent of the corrosion-resistant characteristics of anygrade. Since the information is based on chemically purereagents, customer in-plant testing of a particular stainlessalloy in the actual environment is recommended to determinesuitability.

MANUFACTURING CONSIDERATIONS

The typical fan manufacturer rarely has the opportunity topurchase an adequate quantity of duplicate parts in the samestainless alloy construction that would warrant direct purchasesfrom the mill. Instead, “per job” purchases limit the fanmanufacturer to those alloys which are most readily availablefrom steel distributors.

Because of the dissimilar physical and mechanical properties ofthe stainless alloys, equipment fabrication methods often varyfrom the standards established for mild steel construction. Forexample, production equipment capable of handling 1/4"carbon steel may only be capable of handling 3/16" thickstainless. Likewise, the basic fan construction may involveheavy gauge components cut to size on standard flame burningequipment, but when stainless steel is required plasma-arccutting equipment becomes necessary. Typically, fan

construction involves spun-inlet venturi sections or spun wheelcomponents. It may be more economical to furnish all suchspinnings of one grade of stainless steel, allowing aninterchangeable inventory. Similarly, castings may befurnished of one grade of stainless instead of maintaining 3 or4 various grades and incurring added inventory expense.

Of the Austenitic alloys shown in the summary on page 4, 304,304L, 316, 316L, and 347 stainless steels provide an adequatevariety of corrosion resistance and strength characteristics andare readily available from steel distributor stock. These specificstainless steel alloys can be consolidated into versatile 304, 316,and 347 stainless steel construction groupings.

Recognizing the availability of these various stainless steelconstruction classifications, a determination can be maderegarding the suitability of a particular group for a givenapplication based on the following:

304 stainless steel - good corrosion resistance at aminimum price. Under this alloy grade, machined partssuch as shafting could be furnished from 304 stainlesssteel. However, in order to optimize production, nyb onlyoffers 316 stainless steel shafting. Welded parts such ashousings or wheels must be fabricated from 304L stainlesssteel. Beyond 800°F., the strength characteristics of 304stainless steel are not sufficient to warrantrecommendation.

316 stainless steel - better corrosion resistance than 304and good strength characteristics at elevated temperatures.Though higher in price, this alloy grade is the mostversatile. Welded components must be fabricated from316L stainless steel which is a low carbon grade stabilizedfor welding.

347 stainless steel - corrosion-resistant characteristicssimilar to 304 stainless steel but with the highest strengthcharacteristics at elevated temperatures. Since it is thehighest in initial cost and most difficult to obtain, 347stainless steel should only be used where rotating speeds andelevated temperatures demand its use for wheelconstruction.

The corrosion-resistance guide on page 3 provides a reference tothe corrosion-resistance characteristics of 304 and 316stainless steel alloys. For the purposes of this guide, thecorrosion-resistance of 347 stainless steel is considered similar to304 stainless steel and should only be used if high temperatureis a factor. Note that this is a condensed reference and does notrepresent the full extent of the corrosion-resistancecharacteristics of any grade. Because this information is basedon chemically pure reagents, customer in-plant testing of aparticular stainless alloy in the actual operating environment isrecommended to determine suitability.

CORROSION-RESISTANCE GUIDE

Stainless Steel Alloy Stainless Steel AlloyCorrosiveAgent 304* 316

CorrosiveAgent

304* 316

Acetic AcidAcetic AnhydrideAcetone

SSE

EEE

Lactic AcidMagnesium CarbonateMercuric Chloride

SEN

EEN

AcetyleneAluminum AcetateAluminum Chloride (dry)

EEN

EES

Methyl AlcoholMethyl Ethyl KetoneMineral Oil

EEE

EEE

Ammonia (dry)Ammonia (wet)Ammonium Sulfite

EES

EEE

MoistureNapthaNitric Acid

EEE

EEE

AnilineBarium ChlorideBenzene

EEE

EEE

OzonePerchloric AcidPhenol

SNE

SNE

Boric AcidBromine WaterButane

ENE

ENE

Phosphoric AcidPolyvinyl AcetatePotassium Chloride

SES

EEE

Calcium ChlorideCarbon Tetrachloride (dry)Chlorine Gas (dry)

SSS

SES

Potassium CyanidePotassium DichromatePotassium Hydroxide

EEE

EEE

ChlorobenzeneCitric AcidCopper Sulfate

SEE

SEE

PyridineSalt SpraySilver Nitrate

SSE

SSE

CyclohexaoneEthyl AcetateEthyl Alcohol

SSE

SEE

Sodium BicarbonateSodium ChlorideSodium Cyanide

ESE

EEE

Ethylene DichlorideEthylene OxideFerric Chloride

ESN

ESN

Sodium DichromateSodium HydroxideSodium Hypochlorite

SEN

SEN

Ferric NitrateFluorine Gas (dry)Formaldehyde

EEE

EEE

Sodium SulfateSteam VaporSulfamic Acid

EET

EES

Formic AcidGasolineGlycerine

SEE

EEE

Sulfur Dioxide (dry)Sulfur Dioxide (wet)Sulfuric Acid

SNN

ESS

Hydrochloric AcidHyfrofluoric AcidHydrogen Peroxide

NNE

NNE

Tannic AcidTolueneTrichloroethylene

SES

EES

Hydrogen Sulfide (dry)Hydrogen Sulfide (wet)Iodine

SNN

ESN

XyleneZinc ChlorideZinc Sulfate

ENE

ESE

E = Excellent S = Satisfactory N = Not Recommended T = Test data not available

* 347 stainless steel is considered to have the same corrosion-resistance characteristics as 304 stainless steel.

SPARK RESISTANCE

A common misapplication of stainless steel is in areas requiringnon-sparking materials. Since stainless steels are basicallyalloys of chromium and iron, or of chromium, iron, and nickel,they are considered ferrous and sparking. As a result, theavailability of SRC with stainless steel construction is verylimited. In some cases, a Monel shaft and/or Monel buffers

may be furnished to allow for some types of SRC construction.However, in other cases, all that is available are steps short ofSRC construction which can be added to the fan to minimizethe potential for generating sparks. The specific modificationsvary depending upon the product, so consult nyb for availability.

SPECIAL ALLOYS

Under the general description of stainless steel, there are manyother special alloys; some more corrosion resistant, and somemore abrasion resistant. These specialized alloys require carefulconsiderations of costs, availability, design suitability, andfabrication methods. Therefore, their selection and specificationshould be left to specific applications.

SUMMARY

Any equipment is only as good as its weakest component. If thecorrosive gas stream requires that 316 stainless steel be

specified, 304 or 347 stainless steel should not be substitutedbecause of limited corrosion resistance. If 304 stainless steelis all that is necessary to combat the corrosion and 316stainless steel wheel construction is adequate to obtain thesafe speed at the required temperature, there is no reason tosubstitute the more expensive 347 stainless steel alloy. The347 stainless steel alloy grade should never be specifiedbased solely upon its corrosion-resistant characteristics; itsonly advantage over 316 is higher rotating speeds at elevatedtemperatures.

SUMMARY OF AUSTENITIC STAINLESS STEEL TYPES

For m 607 GAW

Type 301. A 17% Cr., 7% Ni. grade used primarily instructural applications and where high strength plus highductility is required. Corrosion resistance is slightly less thanType 302.

Type 302. The basic 18% Cr. 8% Ni. possesses excellentcorrosion resistance to many organic and inorganic acids andtheir salts at ordinary temperatures. Also has good resistanceto oxidation at elevated temperatures. Can be readilyfabricated by all methods usually employed with carbonsteels. Cr-Ni grades are nonmagnetic in the fully annealedcondition and cannot be hardened by conventional heattreatment. Type 302 is subject to carbide precipitation due towelding.

Type 303. The basic 18-8 composition with the addition ofone or more other elements, usually phosphorus, sulfurand/or selenium to improve machinability. Also used whenminimum galling and seizing is desired. Corrosion resistanceunder certain conditions may be somewhat lower than Type302. Special precautions are necessary in welding Type 303.

Type 304. Similar to Type 302 in chemical analysis exceptcarbon is .8% max. The lower carbon decreases susceptibilityto carbide precipitation in the 800°F. to 1550°F. temperaturerange, making it useful over a wider range of corrosiveconditions than Type 302.

Type 304L. An extra low carbon analysis similar to Type304 except carbon is .3% max. Carbide precipitation doesnot occur if material is not held over two hours in the 800°F.to 1550°F. temperature range. Thus corrosion resistance isnot affected by normal welding and stress relievingapplications.

Type 305. A modified Type 304 grade of lower chromium,higher nickel content to reduce tendency to work hardenwhen severely cold worked. Particularly well suited fordifficult forming, perforating, etc., where rapid workhardening makes fabrication difficult.

Type 308. A 20% Cr. 10% Ni. grade providing somewhatbetter corrosion resistance than the 18-8 grades. Because ofits higher alloy content, it is less susceptible to carbideprecipitation than Type 304.

Type 309. A 24% Cr. 12% Ni. steel combining excellentresistance to oxidation with high tensile and creep strength atelevated temperatures. It resists oxidation at temperatures upto 2000°F. under normal conditions.

Type 310. A 25% Cr. 20% Ni. analysis having slightly higheroxidation resistance and creep values than Type 309. LowerCoefficient of Expansion gives less tendency to warp andthrow scale in fluctuating temperatures.

Type 314. Essentially Type 310 with the addition ofapproximately 2.50% silicon to increase resistance tooxidation and to retard carburization.

Type 316. A modified 18-8 grade containing approximately2.50% molybdenum. It is more resistant to corrosive action ofmost chemicals, especially sulfuric acid and fatty acids. Type316 is less susceptible to pitting and pin hole corrosion byacetic acid vapors, chloride solutions, etc. The tensile andcreep strength at elevated temperatures are also superior tothe other Cr-Ni types. Type 316 is subject to carbideprecipitation due to welding.

Type 316L. Similar to Type 316 in analysis except carbon is.3% max. It is immune to harmful intergranual corrosionproviding it is not held in the 800°F. - 1550°F. temperaturerange for over two hours.

Type 317. A modified 18-8 stainless containing approximately3.50% molybdenum. Resistance to corrosion is somewhatbetter and susceptibility to carbide precipitation is slightlyless than Type 316.

Type 321. A modified 18-8 analysis with titanium (five timescarbon content minimum) added to make it immune toharmful carbide precipitation. The corrosion resistance ofType 321 is the same as Types 347 and 304.

Type 347. A modified 18-8 formulation with columbium(two times carbon content minimum) added to make itimmune to harmful intergranular corrosion. The corrosionresistance of Type 347 is the same as Type 304.


PRACTICAL LIMITS OF SPARK-RESISTANT CONSTRUCTIONINTRODUCTION

Fan applications with airstreams of explosive or flammableparticles or gases require spark-resistant system components forthe safe handling of such airstreams. This includes componentssuch as ductwork, dampers, filter devices, heating or coolingcoils, and fans. This Engineering Letter presents practicalconsiderations and methods of providing fans with varyingtypes of Spark-Resistant Construction (SRC).

THE AMCA STANDARD

The Air Movement and Control Association (AMCA)established a standard set of Classifications for Spark-ResistantConstruction. For reference, that Standard is shown here in itsentirety.

AMCA STANDARD 99-0401-86Classification for Spark-Resistant Construction

Fan applications may involve the handling of potentially explosive or flammable particles, fumes, or vapors. Such applicationsrequire careful consideration of all system components to ensure the safe handling of such gas streams. This AMCA Standarddeals only with the fan unit installed in that system. The Standard contains guidelines which are to be used by both themanufacturer and user as a means of establishing general methods of construction. The exact method of construction and choiceof alloys is the responsibility of the manufacturer; however, the customer must accept both the type and design with fullrecognition of the potential hazard and the degree of protection required.

TYPE CONSTRUCTIONA All parts of the fan in contact with the air or gas being handled shall be made of nonferrous material. Steps must also

be taken to assure that the impeller, bearings, and shaft are adequately attached and/or restrained to prevent a lateral oraxial shift in these components.

B The fan shall have a nonferrous impeller and nonferrous ring about the opening through which the shaft passes.Ferrous hubs, shafts, and hardware are allowed, provided construction is such that a shift of impeller or shaft will notpermit two ferrous parts of the fan to rub or strike. Steps must also be taken to assure that the impeller, bearings, andshaft are adequately attached and/or restrained to prevent a lateral or axial shift in these components.

C The fan shall be so constructed that a shift of the impeller or shaft will not permit two ferrous parts of the fan to rub orstrike.

Notes1. No bearings, drive components, or electrical devices shall be placed in the air or gas stream unless they are

constructed or enclosed in such a manner that failure of that component cannot ignite the surrounding gas stream.

2. The user shall electrically ground all fan parts.

3. For this Standard, nonferrous material shall be any material with less than 5% iron or any other material withdemonstrated ability to be spark resistant.

4. The use of aluminum or aluminum alloys in the presence of steel which has been allowed to rust requires specialconsideration. Research by the U.S. Bureau of Mines and others has shown that aluminum impellers rubbing on rustysteel may cause high-intensity sparking.

The use of the above Standard in no way implies a guarantee of safety for any level of spark resistance. “Spark-resistantconstruction also does not protect against ignition of explosive gases caused by catastrophic failure or from any airstream materialthat may be present in a system.”

This Standard applies to: Centrifugal Fans; Axial and Propeller Fans; Power Roof Ventilators.

This Standard applies to ferrous and nonferrous metals. The potential questions which may be associated with fansconstructed of FRP, PVC, or any other plastic compound were not addressed.

THE LIMITATIONS OF SRCThe AMCA standard provides the system designer with auniform way to specify the system requirements and providesfan manufacturers with general guidelines. The fan manufacturermust still develop unique designs to deal with the physical andpractical limitations of fan equipment when developingconstruction methods to comply with AMCA.

A major limitation is the practical availability of truly“nonferrous” alloys that really can be used in fan construction.There are certain alloys or noble metals than are trulynonferrous, alloys that contain no iron, but for the most part theyare extremely expensive and/or difficult to obtain in forms andstrengths necessary for fan construction.

For most purposes, the fan manufacturer uses more readilyavailable alloys that are considered nominally nonferrous andwhich have strength and work properties suited to fanconstruction. The New York Blower Company’s list of usablealloys is shown in Figure 1.

Alloy % FE (iron)Aluminum 5052* 0.45Aluminum 6061* 0.70Brass CDA 360 0.00Bronze CDA 958 4.75Copper CDA 110 or 122 0.00Monel 400 Shafting 2.50

Note: Alternate alloys may be substituted; not to exceed 5%iron content. Hardware, such as setscrews or keys, may havean iron content greater than 5% provided they are recessedand relatively inaccessible.* Iron content in most aluminum alloys is actually a randomcontamination and not a predicted element of the alloy.

Figure 1 - Spark-Resistant Alloys used by nyb

Aluminum is the most frequently used alloy due to its low cost.However, as pointed out in the AMCA Standard, whenaluminum is in close proximity to steel, careful maintenanceprograms are necessary to prevent rust, because aluminumrubbing against rusty steel can cause high-intensity sparking.In applications where such maintenance is not possible, an SRCmethod that places steel in the airstream is not recommended.

Regardless of which classification is chosen, airborne foreign or“tramp” particles could either strike each other, or strike one ofthe components of the fan, causing a spark. Protection againstsuch occurrence cannot be built into the fan itself.

SRC does not eliminate the potential for spark generation. Fanswith any type of SRC are only intended to minimize thepotential that any two or more fan components might generatesparks within the airstream by rubbing or striking duringoperation. No type of SRC can be guaranteed to eliminate thepossibility of generating a spark, nor would any SRC typepreclude sparks resulting from any foreign influence such asairborne materials striking each other.

The AMCA Standard requires construction that will not permit awheel and/or shaft to shift due to some malfunction duringoperation. If two components are allowed to shift and rubagainst each other for any length of time, either sparks orfrictional heat could become a hazard in an explosive orflammable gas stream. Normally, standard procedures of

fastening the wheel to the shaft and locking the shaft in thebearings are sufficient. However, the degree of hazard in thesesituations dictates that extraordinary precautions to moresecurely prevent such shifting are in order, so further methods ofattachment or restraint are required.The following types of SRC are furnished by The New YorkBlower Company. These types meet the AMCA Standard, butgo a step further by explaining the specific construction methodsused to achieve SRC.NEW YORK BLOWER SRC STANDARDSAIRSTREAM-TYPE SRC - (AMCA Standard 99-0401-86,Type A) to include all airstream parts constructed of a spark-resistant† alloy. Bearing stop blocks and/or an aluminum shaftsleeve shall be provided to prevent contact of the shaft with thehousing at the shaft opening. Shaft set collars shall be providedto prevent axial movement of the shaft through the bearings.The fan wheel shall be secured to the shaft in such a manner thatit cannot shift axially on the shaft.WHEEL-TYPE SRC - (AMCA Standard 99-0401-86, Type B)to include the wheel constructed of a spark-resistant† alloy, anda buffer around the housing shaft opening. Bearing stop blocksand/or an aluminum shaft sleeve (in lieu of buffer) shall beprovided to prevent contact of the shaft with the housing at theshaft opening. Shaft set collars shall be provided to prevent axialmovement of the shaft through the bearings. The fan wheel shallbe secured to the shaft in such a manner that it cannot shiftaxially on the shaft.BUFFER-TYPE SRC - (AMCA Standard 99-0401-86, Type C)to include buffers constructed of a spark-resistant† alloyattached to the housing interior adjacent to the wheel front andback. Fan designs which incorporate a conical inlet venturiwithin the confines of the housing shall utilize a spun-aluminumventuri in lieu of a separate buffer on the inlet side. A buffer willalso be located at the housing shaft opening.† The term “spark-resistant alloy” may include, but is notlimited to, those alloys shown in Figure 1.WHAT THE NYB SRC TYPES OFFER AND HOW THEYARE ACCOMPLISHEDOne or more of these SRC types are offered on most New YorkBlower fans as indicated in the specific literature describingthose fans.Of these types, a fan furnished with AIRSTREAM-TYPE SRCshould provide the greatest degree of spark resistance. In theevent that two or more fan components in the airstream rub orstrike together, a properly maintained fan should be able tocontinue in operation for some reasonable period of time,without producing a spark. However, the severity of a hazardthat calls for AIRSTREAM-TYPE SRC dictates that the fanshould be closely monitored and shut down immediately uponsuch an occurrence. If allowed to operate, the rubbing or strikingof these fan components will generate frictional heat, quicklydeteriorate, and eventually catastrophically fail. Good safetypractice cannot be ignored!A fan furnished with WHEEL-TYPE SRC differs fromAIRSTREAM-TYPE SRC in that only the wheel itself isconstructed of a spark-resistant alloy. A spark-resistant buffer isadded around the housing opening through which the shaftpasses as shown in Figure 2. The remainder of the fancomponents are furnished in their standard material, usuallymild steel.

Fans furnished with WHEEL-TYPE SRC should notcontinue in operation for any length of time with the wheelrubbing any component or with the shaft striking thebuffer. Practically speaking, it is not possible to predict a“safe” length of time, because there may be other ferrouscomponents within the fan airstream which could be torn orjarred loose by the rubbing or striking of the wheel or shaft,and such loose ferrous objects could create a spark. Also, thebuffer cannot support the weight or withstand the forces of therotating shaft for any prolonged period of time.

The AIRSTREAM-TYPE and WHEEL-TYPE SRC specificationsgo further to minimize the potential for sparking by takingextraordinary precautions to minimize the potential for abnormalmovement or shift of the fan’s airstream components. Whilethe standard bearing mounting bolts will resist vertical or axialmovement, the addition of bearing stop blocks will resist horizontalmovement and effectively secure the bearings in place. Theaddition of shaft set collars as shown in Figure 3 will furtherresist shaft movement through the bearings. These combinedfeatures virtually eliminate the possibility of any movement inthe shaft and bearing assembly.

There are many ways to secure the fan wheel to the shaft, butstandard setscrews and keys are not enough for the moresevere applications. Figure 4 details one alternative whichincludes a bolted aluminum wheel retaining plate on the endof the shaft. Other methods might include countersinking theshaft to accept a setscrew, sweat-fitting, or tapered bores toprevent the wheel from slipping on the shaft axially. Theprecise method will vary by fan size and type.

The BUFFER-TYPE SRC specifications utilize standard,usually mild steel, airstream component parts and employspark-resistant plates or buffers to stop the wheel or shaft fromcoming into direct contact with other airstream components. Afan design which requires an inlet cone is usually furnishedwith an aluminum cone to act as the buffer on one side, asshown in Figure 5. Other designs might utilize a spark-resistant band or plate.

The BUFFER-TYPE SRC is intended to provide a low costalternative for non-critical applications. The user or specifiermust exercise caution in selecting this type so that the safetyof the installation is not compromised for the sake of initialcost.

Generally, aluminum wheel construction is utilized forAIRSTREAM-TYPE AND WHEEL-TYPE SRC. Because thematerial strength characteristics of aluminum decrease sharplyat elevated temperatures, it is not recommended for handlinganything other than nonabrasive airstreams at less than 200°F.In cases beyond these limits, BUFFER-TYPE SRC may be theonly readily available alternative.

As with the WHEEL-TYPE SRC, fans furnished withBUFFER-TYPE SRC should not continue to operate forany length of time with the wheel or shaft rubbing thebuffers. High speed fans will tend to wear away buffers morerapidly than slower speed fans, and thus BUFFER-TYPESRC should be used with caution on high speed fans. Thegreater wheel tip speeds and shaft surface speeds,combined with their corresponding weights and forces,reduce the amount of time available to react.

When a high speed fan application requires spark resistancebut AIRSTREAM- and WHEEL-TYPE SRC are not practical,The New York Blower Company will work with the systemdesigner to provide special spark-resistant features on a caseby case basis.

Periodic inspection of the fan, and particularly the airstream, isrecommended. The build-up of foreign material or rust, thepotential deterioration due to abrasion or corrosion, or theaccidental shifting of any fan part could lead to further hazardsof potential ignition or explosion.

Figure 2 – Spark Resistant Buffer

Figure 5 - Aluminum InletCone (1), Steel Wheel (2)

Figure 3 – Bearing Stop Blocks/Shaft Set Collars

Figure 4 - Wheel Retaining Plate

The centrifugal fan arrangements most compatible with theintended use of SRC are those in which the wheel is overhungon the shaft and the bearings are outside the airstream. Sucharrangements include Arrangements 1, 8, 9, and 10 asdescribed in AMCA Standard 99-2404-78.

One item mentioned in the AMCA Standard for SRC is thatthe user must electrically ground all fan parts. This is necessaryso that any electrical charge or static electricity that mightbuild up in operation can be safely conducted away. Thoughthere is probably sufficient electrical conductivity throughmost bearings to transmit any static charge to the bearingpedestal, brush type contacts on the pedestal may be a goodadded precaution. The pedestal can then be suitably grounded tothe support structure. Steps should be taken by the user toensure electrical conductivity to the connecting ductwork.

AXIAL FANS AND SRCPropeller Fans, Duct Fans, Vaneaxial Fans, and TubularCentrifugal Fans have the common difficulty of placing thebearings, and sometimes the drive components, either directly inthe airstream or in an inner tube construction that is locatedwithin the airstream as shown in Figure 6.

The New York Blower Company offers WHEEL- andBUFFER- TYPE SRC on its Duct, Tubeaxial, Vaneaxial, andTubular AcoustaFoil fan lines.BUFFER-TYPE SRC on these fans requires bearings and drivecomponents to be isolated from the airstream. To accomplishthis, the fans are furnished with shaft seals and all airstreamjunctions are continuously welded and/or gasketed withsuitable material. To prevent a shift of the impeller and/orshaft, a ceramic-felt shaft seal with retaining plates constructedof copper is used. For Tubeaxial and Vaneaxial fans, analuminum wheel is also required. On the Duct Fan, a partialaluminum wheel is used.

WHEEL-TYPE SRC utilizes all of the modifications ofBUFFER-TYPE SRC. The addition of a wheel retainer, setcollars, and bearing stop blocks help prevent a lateral or axialshift of the wheel, bearings, and shaft.

FIBERGLASS-REINFORCED PLASTIC AND SRCCentrifugal fans made of FRP material present an excellentdegree of spark resistance as FRP materials are nonsparking.However, FRP is also a nonconductor so the possibility ofbuilding and retaining a static charge is greater and must beaccounted for. Adding graphite to the final resin finish willprovide the necessary conductivity to alleviate this situation.The special construction features of FRP fans may also call forother considerations in dealing with hazardous fumes. SeeEngineering Letter 20.

WHERE TO AVOID ATTEMPTING SRCThe basic requirement that bearings should not be placed inhazardous airstreams eliminates several centrifugal fanarrangements from consideration. Single-width or double-width fans in either Arrangement 3 or Arrangement 7, wherethe fan bearings are located in the inlet, should not be furnishedfor such service. See Figure 7.

CONCLUSION

Moving explosive or flammable gas streams through fansrequires the utmost care in system design and equipmentselection. The system designer must weigh the total systemfrom all angles to minimize risk, particularly when the systemcomponents and/or fans are in environments that are located inareas where people are likely to be working or passing.The explosiveness of the gas mixture, the people factor, andthe potential for foreign or “tramp” elements to enter thesystem, are all necessary concerns in determining to whatdegree special-material construction should be used. Vibrationdetectors to warn of impending malfunction of bearings orrotating assemblies are a good preventive measure to forestallthe actual rubbing or impact of two parts in any mechanicalequipment, and should certainly be considered in “severe risk”situations. The extraordinary measures to pre-vent wheel andshaft movement offered in nyb’s AIRSTREAM-TYPE SRCand WHEEL-TYPE SRC are features to help minimize thepotential of allowing two parts to strike.

The three classifications of spark-resistant construction inAMCA’s Standard and the specific construction methodsoffered by New York Blower provide only degrees of resistanceto sparking. They have been used, and are continuing in use, asdeterrents to possible sparking and ignition in hazardoussystems. Care must be taken to recognize that there are noabsolute guarantees.

Therefore, in particularly hazardous applications, the location ofthe fan and perhaps the entire system should be a majorconsideration. In some cases, protective enclosures around thefan or other mechanical parts in the system may be anotherprotective step to lessen the danger in the event that a sparkmight occur in spite of the precautions taken. The systemdesigner is in the best position to weigh the alternatives andspecify the required fan equipment.

Form 607 GAW

Figure 7 - Arrangement 3 Double-Width Fan

Figure 6 – Tubeaxial Fan


CORROSION-RESISTANT COATINGS FOR FAN EQUIPMENT

INTRODUCTION

This Engineering Letter provides basic information regardingthe different types of corrosion-resistant coatings readilyavailable for fan equipment. The coatings are described hereaccording to generic classifications having similar characteristicssuch as curing methods, adhesion qualities, chemical resistance,and temperature limitations. Coating manufacturers offer avariety of brand name coatings which can be categorized bythese generic classifications.

The service life of air-moving equipment constructed of carbonsteel may be significantly reduced when corrosives are allowedto attack the surface of the metal through chemical or electro-chemical action. One method of inhibiting this corrosive action isby applying a protective coating to the area in contact with thecorrosives. Protective coatings act as a barrier between thecorrosive and the parent material. A wide range of protectivecoating systems is available to provide protection from a varietyof corrosives including acids, alkalis, solvents, salts, and oils.Although other materials of construction, such as special alloys(see Engineering Letter 14) and fiberglass-reinforced plastic -FRP (see separate Engineering Letters) are available, protectivecoatings can offer a low-initial-cost solution to the corrosionproblem.

The selection of a protective coating is critical in determiningthe service life of the equipment. The selection process mustconsider the actual chemical composition of the gas stream. Toevaluate the corrosive nature of the gas stream completely, theconcentrations and temperatures of the chemicals present mustalso be considered.

COATING INGREDIENTS

Although protective coatings are differentiated by their specificchemical composition, the most common consist of three basicingredients; a binder, a flow control agent, and a pigment orfiller. When these ingredients are combined, they can range inconsistency from thin liquids to semi-solid pastes in a variety ofcolors.

The binder is the film-forming ingredient in the coating. Itconsists of either a drying oil or a polymeric substance.

Drying oils form a hard film by reacting with oxygen in the air.Coatings with this type of binder are usually cured by air dryingbut in some cases may be baked in order to cure more rapidly.Coatings that utilize a polymeric substance as the binder requirea “thermoset” cure. Thermosetting can be accomplished bybaking the applied coating in some cases or by adding a catalystin other cases. The type of thermoset is dependent upon thecharacteristics of the polymeric substance itself.

A flow control agent, or solvent, is combined with the binder toform the liquid portion of the coating. The solvent prevents thebinder from solidifying prematurely and ensures uniformdispersion over the surface. This combination of binder andsolvent is called the vehicle portion of the coating.The pigment is any substance, usually a powder, which givescolor to the mixture. Most pigments are insoluble in solventsand are not affected by the vehicle portion of the coating.

The generic coating classifications are differentiated by theirchemical composition. While the chemical composition alone isnot sufficient in determining which protective coating is selectedfor a specific application, it can be useful in determining thegeneric group of a particular brand name coating.COATING TYPES

The following descriptions of generic coatings present curingmethods, adhesion qualities, chemical resistance, andtemperature limitations. This information can be used as a guideto specifying and selecting corrosion-resistant coatings for fanequipment. For chemical resistance to specific applications refer tothe Corrosion-Resistance Table beginning on page 3.

Phenolic - resin systems include any of the several types ofthermosetting resins obtained by the condensation of phenol orsubstituted phenols with aldehydes, such as formaldehyde,acetaldehyde, or furfural. Phenolic resins can be cured bybaking, air-drying, or catalyzation. These curing processesremove the solvents and oxidize the oils contained in the resin toproduce coatings with an extremely hard finish. Phenoliccoatings possess excellent resistance to moisture, solvents, and awide variety of concentrated acids at temperatures to 150°F. air-dried or 400°F. baked.

Epoxy - coatings are derived from a thermosetting resin basedon the reactivity of the epoxide group. The most common formof this resin stems from a reaction between epichlorohydrin andbisphenol A. Another type is formed from the oxidation ofpolyolefins with peracetic acid. Epoxy resins can be cured bybaking or catalyzing. When cured, these coatings have a supplefinish and superior adhesion qualities. Epoxy coatings arecharacterized by their excellent resistance to a variety ofcorrosive chemicals, including acids, alkalis, and salts withtemperature limitations between 200°F. and 300°F.

Epoxy-Phenolic - coatings are modified phenolic coatingscreated by blending phenolic resins with resins from the epoxidegroup. Epoxy-phenolics can be cured by baking or by theutilization of a catalyst. Catalyzed epoxy-phenolic coatingsrequire a longer curing time and have lower chemical resistancethan the baked epoxy-phenolic coatings. They can be applied ingreater thickness to attain virtually the same

performance characteristics as the baked epoxy-phenoliccoatings. These coatings are used mainly for alkali-resistancein moderate temperatures up to 400°F.

Inorganic Zinc - coatings are formulated by adding zinc dustto inorganic binders. These binders give the zinc coatingstheir corrosion-resistant qualities, while the zinc adds cathodicprotection (alters the rate of electron flow which can producecorrosion) to metals below it in the galvanic series. (However,the zinc-rich coatings are not recommended for use overaluminum substrates.) These coatings, which are cured by air-drying, are not subject to ultraviolet degradation and may beused without a top coat for severe weathering conditions. Theinorganic zinc coatings have good solvent-resistant properties,but may require an appropriate protective top coat for acid oralkali-resistant applications. Inorganic zinc coatings aresuitable for temperatures to 750°F.

Vinyl - coatings use resins from the vinyl-resin family as themajor portion of the binder. These resins are formed by areaction between acetylene and an acid. They consist largelyof vinyl acetate, vinyl chloride, and vinyl copolymer. Vinylcoatings can be cured by baking or air-drying, and have excellentadhesion qualities to steel. As the vinyl dries, the film remainsnon-brittle and will easily follow the expansion and contractionof the underlying surface. Vinyl coatings are unique in thatthey possess superior corrosion-resistant performance over abroad range of corrosive combinations. The vinyl coatingswill give satisfactory results for most corrosive fume applicationsbelow 200°F., but are not recommended for solvent-ladenenvironments.

Coal Tar Epoxy - coatings are formed by combining coal tar, ablack liquid obtained from the distillation of coal during theconversion of coke, and a resin from the epoxide group.These coatings, which are cured by air-drying or catalyzing,adhere well to metal surfaces. The blend of coal tar and theepoxy resin forms a coating which has good water-resistancecharacteristics and is resistant to acids and alkali fumes attemperatures to 250°F.

Alkyds - are actually a type of polyester resin modified by theaddition of fatty acids or drying oils. These resins are aproduct of the thermosetting reaction between polyhydricalcohol and a poly-basic acid. Alkyd resins are cured bycatalyzation or air-drying. They have the ability to harden atroom temperatures in a very short time. These coatings arenot generally selected or specified for corrosion-resistantapplications, but are normally required for color-matchingpurposes.

Silicone - coatings are polymeric silicones formed by heatingsilicon in methyl chloride to yield methylchlorosilanes whichare separated and purified by distillation. The desiredcompound is then mixed with water. Silicone coatings can becured by baking or air-drying. Formulated for medium to hightemperature service where temperatures seldom fall below200°F. to 300°F., these coatings normally exhibit good toexcellent fume resistance to acids, alkalis, solvents, salts, andwater, but are not recommended for areas subjected to acid oralkali splash or spillage. Silicone coatings possess good

weathering characteristics, but an inorganic zinc primer willgreatly extend the coating’s service life when applied to steel,especially if service temperatures fall below 300°F. andmoisture is present. The maximum temperature limitation forthese coatings varies according to each specific manufacturer’srecommendation.

Polyurethane - coatings are derived from prepolymerscontaining isocyanate groups and hydroxyl containingmaterials such as polyols and drying oils. Polyurethanecoatings, which are cured by air-drying or catalyzing, arefrequently applied over zinc and epoxy primers. Thesecoatings produce an extremely hard, yet flexible, high glossfinish that is resistant to weathering, ultraviolet degradation,acids, and alkalis at temperatures to 200°F.

Polyester - resins are thermosetting synthetic resins formed bythe polycondensation of dicarboxilic acids and dihydroxyalcohols. Polyester resins are characterized by their ability tocure at room temperatures in a very short time after beingcatalyzed. They also have excellent adhesion qualities. Thepolyester coatings are resistant to mild traces of acids, alkalis,and solvents. The maximum temperature limitation for thesecoatings varies according to each specific manufacturer’srecommendation.

Vinyl Ester - resins are combined with a special curing systemand inert flake pigment. Vinyl ester coatings, which are curedby air-drying or catalyzing, provide excellent chemical resistanceto organic and inorganic acids, oxidizing agents, salts, and awide range of solvents. Vinyl ester coatings are applied at 35to 40 mils DFT.

Although these coatings cover a broad range of generic types,they by no means cover them all. Types mentioned here are themost commonly specified and selected generic types offeredfor use on fan equipment.

Selecting the proper coating system for the application is notenough to ensure its success. Proper surface preparation isessential to the effectiveness of any coating system.

COATING SURFACE PREPARATION

Surface preparation not only ensures that the coating willadhere adequately but also removes contaminants which couldbe detrimental to the service life of the equipment. The SteelStructures Painting Council further defines various types ofsurface preparation as shown in Engineering Letter 17. Coatingmanufacturers then suggest the recommended degree of surfacepreparation for each of their brand name coatings.

Based on the surface preparation necessary for each coatingspecification, nyb will either apply the coatings in its facilitiesor have them sent to an outside applicator. Most coatingsapplied by nyb receive a combination of phosphate wash andhand tool cleaning. This procedure removes all oil, dirt, grease,loose rust, and mill scale that hinders the effectiveness of thecoating. This method of surface preparation is equivalent to acombination of Solvent Cleaning (SSPC-SP1) and Hand ToolCleaning (SSPC-SP2). The application of a coating whichrequires any degree of sandblasting is handled by an outsideapplicator. Sandblasting is further defined in EngineeringLetter 17.

Airstream, exterior, and all surfaces are common arearequirements for coatings. Airstream surfaces coated -includes interior of housing, entire wheel, that portion of theshaft in contact with the airstream, airstream areas of collar,inlet ring and/or inlet plate, and all surfaces of the inlet cone.Exterior surfaces coated - includes all outside surfaces,except bearings, motor, and the shaft. All surfaces coated -includes all surfaces inside and outside, except bearings,motor, and that portion of the shaft not in contact with theairstreamAPPLICATION AND SELECTION GUIDEThe table below provides a condensed guide to the corrosion-resistant properties of generic coatings commonly available onfan equipment. Each coating should be chosen according tothe specific corrosive chemical or chemicals involved in theapplication. The customer is responsible for selecting the

coating which will provide proper protection. nyb can onlywarrant that the coating will be applied according to the coatingmanufacturer’s instructions.

Fume-and aerosol-contaminated air has been used as the basisfor this guide. The fumes or aerosols of a substance areeffectively diluted by air, reducing the chemical concentration toa level significantly lower than the liquid solution. Because thisguide is based on dilute concentrations of fumes and aerosols,relatively few chemicals are listed as unsatisfactory for use withthese protective coating systems.

Protective coatings play an important role in corrosion-resistantconstruction. They often have the lowest first cost. Specialalloy and fiberglass-reinforced plastic construction are alsoavailable for corrosive applications. Special alloy and FRPconstruction are able to handle a wider range of corrosives, arefar superior when it comes to corrosion resistance, and manytimes result in the lowest life cycle cost.

CORROSION-RESISTANCE GUIDE TO GENERIC COATINGS AND ASSORTED METALS

COATINGS METALS

CorrosiveAgent

Bak

edP

hen

olic

Air

-dri

edP

hen

olic

Cat

alyz

edE

pox

yB

aked

Epo

xy-

Ph

eno

licC

atal

yzed

Ep

oxy-

Ph

eno

lic

Ino

rgan

icZ

inc

Air

-dri

edV

inyl

Co

alT

arE

poxy

Alk

yd

Air

-dri

edS

ilic

one

Po

lyu

reth

ane

Po

lyes

ter

Vin

ylE

ster

Car

bon

Ste

el

Alu

min

um

304*

Sta

inle

ssS

teel

316

Sta

inle

ssS

teel

Acetic AcidAcetic AnhydrideAcetone

EEE

NNN

NNS

TTS

NNS

NNE

EEN

NSN

NNN

SSS

NNE

ENN

ESN

NNE

SEE

SSE

EEE

AcetyleneAluminum AcetateAluminum Chloride (dry)

EEE

EES

SNS

SSE

ENS

ESN

SSS

TTS

NNS

SSS

SSS

TSE

EEE

STN

ESS

EEN

EES

Ammonia (dry)Ammonia (wet)Ammonium Sulfite

EEE

SNS

NNS

EES

SSS

NNN

NNN

SNS

NNN

SNS

SSS

SNS

EEE

SSN

EEN

EES

EEE

AnilineBarium ChlorideBenzene

EEE

SSN

NSS

TEE

NES

ESE

SEN

NSN

NNN

SSS

SSS

SSS

SEN

NNN

NNS

EEE

EEE

Boric AcidBromine WaterButane

ENE

ENS

SSN

ENE

ENS

NNS

ENS

SST

NNS

SNS

SSS

SST

ESE

NNE

SNS

ENE

ENE

Calcium ChlorideCarbon Tetrachloride (dry)Chlorine Gas (dry)

EES

EES

ENS

EET

ESS

NSN

ENN

SSS

NNN

SSS

ESS

ESN

ESE

NSS

SNN

SSS

SES

ChlorobenzeneCitric AcidCopper Sulfate

SEE

NES

NNS

ESE

SEE

SNS

SEE

NSS

NNN

SSS

SSS

SSS

SEE

SNN

SSN

SEE

SEE

CyclohexanoneEthyl AcetateEthyl Alcohol

EEE

NNS

NSS

SSE

NSS

SSE

SSS

NSS

NNN

SSS

SSS

SNS

SES

NSS

SSS

SEE

SEE

Ethylene DichlorideEthylene OxideFerric Chloride

ESS

NNS

NNE

STE

SNE

SES

NSE

NNS

NNN

SSE

NNS

SNS

NNE

SSN

SSN

ESN

ESN

Ferric NitrateFluorine Gas (dry)Formaldehyde

SNE

SNE

NNS

TTE

SNS

NNE

SSE

TTN

NNN

SSS

SNS

TTN

EES

NNS

NES

EEE

EEE


* 347 stainless steel has the same corrosion-resistance characteristics as 304 stainless steel.

The suitability of the coatings found in this table has been based on fume concentration effectively diluted by air at 70°F. Highchemical concentra tion and/or elevated temperatures and/or moisture may significantly reduce a coating’s suitability.

CORROSION-RESISTANCE GUIDE TO GENERIC COATINGS AND ASSORTED METALS

COATINGS METALS

CorrosiveAgent

Bak

ed

Phe

nol

icA

ir-d

ried

Phe

nol

icC

atal

yzed

Epo

xyB

aked

Ep

oxy-

Phe

nol

icC

atal

yzed

Ep

oxy

-P

hen

olic

Inor

gan

icZi

nc

Air

-dri

edV

inyl

Coa

lTar

Epo

xy

Alk

yd

Air

-dri

edS

ilico

ne

Pol

yure

than

e

Po

lyes

ter

Vin

ylE

ster

Car

bon

Ste

el

Alu

min

um

304*

Sta

inle

ssS

teel

316

Sta

inle

ssS

teel

Formic AcidGasolineGlycerine

SEE

NSE

NSE

NEE

NES

NES

NSS

NNS

NSN

SSS

SES

NSS

EEE

NES

NEE

SEE

EEE

Hydrochloric AcidHydrofluoric AcidHydrogen Peroxide

ENN

SNN

NNS

SNN

SNN

NNN

SSS

NNN

NNN

SNS

SNS

NNS

ESE

NNN

NNE

NNE

NNE

Hydrogen Sulfide (dry)Hydrogen Sulfide (wet)Iodine

ESE

SNT

SNN

EST

SNN

NNN

ENN

SSN

NNN

SNS

SNT

SNS

EEE

SNN

SSE

SNN

ESN

Lactic AcidMagnesium CarbonateMercuric Chloride

EEE

SEE

SEE

SEE

EEE

NSS

EET

SSS

NSN

SSS

SSE

SSS

EEE

NNN

SSN

SEN

EEN

Methyl AlcoholMethyl Ethyl KetoneMineral Oil

EEE

SNE

NSE

ESE

SNE

EEE

ENE

NNE

NNS

SSS

EEE

SNS

SNE

SSS

SSS

EEE

EEE

MoistureNapthaNitric Acid

EEE

EEN

ESN

EEN

ESN

EEN

ENS

ESN

ESN

SSS

ESN

ESS

ESS

SSN

SEN

EEE

EEE

OzonePerchloric AcidPhenol

NSE

NNN

NNS

NSS

NNN

SNS

SES

NNN

NNN

SSS

NNS

TSS

SSS

NNS

SNE

SNE

SNE

Phosphoric AcidPolyvinyl AcetatePotassium Chloride

EEE

SNS

NEN

STE

SNE

NNS

ENS

NTS

NNN

SSS

ETS

SNS

EEE

NNS

NTN

SES

EEE

Potassium CyanidePotassium DichromatePotassium Hydroxide

ESN

SSN

NNS

EEN

SSS

NSS

SSE

SSS

NNN

SSS

TSS

SSS

EEE

SSS

NEN

EEE

EEE

PyridineSalt SpraySilver Nitrate

EEE

TSS

NEN

EET

NSN

SEN

TEE

NET

NSN

SES

NES

NST

NEE

SNN

SSN

SSE

SSE

Sodium BicarbonateSodium ChlorideSodium Cyanide

EEN

ESS

SES

EES

EES

NSN

EES

EES

SSN

SES

SSS

SSS

EEE

NNS

SNN

ESE

EEE

Sodium DichromateSodium HydroxideSodium Hypochlorite

NNN

SNN

SSN

SSN

NNN

SSN

SEN

ESS

NNN

SNN

SES

SSN

SEE

SSN

SNN

SEN

SEN

Sodium SulfateSteam VaporSulfamic Acid

EEE

SNS

SSS

EET

ENS

NNN

ESS

SSS

NSN

SES

SNS

SSS

EES

SSN

ESN

EET

EES

Sulfur Dioxide (dry)Sulfur Dioxide (wet)Sulfuric Acid

EEE

SSS

SSN

SSS

SNS

NNN

NNS

SSN

NNN

SNS

TTE

SNN

ESE

ENN

SNN

SNN

ESS

Tannic AcidTolueneTrichloroethylene

EEE

ENN

SEN

TEE

ENN

NES

ENN

ESN

NNN

SSS

SEN

TSN

ESN

NES

NES

SES

EES

XyleneZinc ChlorideZinc Sulfate

EEE

NSS

ESS

EEE

SEE

ENN

NEE

SSS

NSS

SES

ESS

SSS

EEE

SNN

ENS

ENE

ESE


* 347 stainless steel has the same corrosion-resistance characteristics as 304 stainless steel.

The suitability of the coatings found in this table has been based on fume concentration effectively diluted by air at 70°F. Highchemical concentration and/or elevated temperatures and/or moisture may significantly reduce a coating’s suitability.

Form 607 GAW


COATING SURFACE PREPARATION SPECIFICATIONSINTRODUCTION

This Engineering Letter is intended to be an aid for selecting theproper surface preparation specifications for a givenapplication. It also provides a better understanding of the SteelStructures Painting Council (SSPC) surface preparationspecifications, which are the most commonly used. In addition,surface preparation standards published by the NationalAssociation of Corrosion Engineers (NACE) are cross-referenced where applicable.

The life of a coating depends as much on surface preparation ason the subsequent coating system. Surface preparation,therefore, requires thorough consideration. The primaryfunctions of surface preparation are:

To remove surface contaminants and imperfections,such as oil, grease, dust, rust, weld spatter, etc., thatwill affect the performance of a coating.

To provide an anchor pattern or surface profile whichimproves the mechanical bonding of a coating to theprepared surface by increasing the surface area.

Note that all coating systems will fail eventually. However,most premature coating failure can be attributed to inadequatesurface preparation or lack of coating adhesion.

SUMMARY OF COMMON SURFACE PREPARATION SPECIFICATIONS

SSPCSTANDARD DESCRIPTION

SP1 - SOLVENT CLEANING Removal of oil, grease, dirt, soil, salts, and contaminants by cleaningwith solvent, vapor, alkali, emulsion, or steam.

SP2 - HAND TOOL CLEANING Removal of loose rust, loose mill scale, and loose paint by hand chipping,scraping, sanding, and wire brushing.

SP3 - POWER TOOL CLEANING Removal of loose rust, loose mill scale, and loose paint by power toolchipping, descaling, sanding, wire brushing, and grinding.

SP5 - WHITE-METAL BLAST CLEANINGRemoval of all visible rust, mill scale, paint, and foreign matter by blastcleaning. (For very corrosive atmospheres where the high cost of cleaningis warranted).

SP6 - COMMERCIAL BLAST CLEANINGBlast cleaning until at least two-thirds of the surface area is free of allvisible residues. (For conditions where thoroughly cleaned surface isrequired).

SP7 - BRUSH-OFF BLAST CLEANINGBlast cleaning of all except tightly adhering residues of mill scale, rust,and coatings, exposing numerous evenly distributed flecks of underlyingmetal.

SP10 - NEAR-WHITE BLAST CLEANINGBlast cleaning until at least 95% of the surface area is free of all visibleresidues. (For high humidity, chemical atmosphere, marine, or othercorrosive environments).

SSPC-SP1, “SOLVENT CLEANING”

This specification includes simple solvent wiping, immersion insolvent, solvent spray, vapor degreasing, steam cleaning,emulsion cleaning, chemical paint stripping, and alkalinecleaners. Solvent Cleaning is used primarily to remove oil,grease, dirt, soil, drawing compounds, and other similarorganic compounds.

SSPC-SP2, “HAND TOOL CLEANING”

Hand Tool Cleaning is an acceptable method of surfacepreparation for normal atmospheric exposures, for interiors,and for maintenance painting when using paints with goodwetting ability. This specification includes hand chipping,scraping, sanding, and wire brushing. Hand Tool Cleaning isused primarily to remove loose rust, loose mill scale, and loosepaint after all oil, grease, and salts are removed as specified inSSPC-SP1, “Solvent Cleaning.”*

SSPC-SP3, “POWER TOOL CLEANING”

Power Tool Cleaning provides a better foundation for thepriming paint than Hand Tool Cleaning. This specificationincludes power tool chipping, descaling, sanding, wire brushing,and grinding. Power Tool Cleaning is used primarily toremove loose rust, loose mill scale, and loose paint after alloil, grease, and salts are removed as specified in SSPC-SP1 -Solvent Cleaning.

SSPC-SP5, “WHITE-METAL BLAST CLEANING”

This blast cleaning method is generally used for exposures invery corrosive atmospheres and for immersion service wherethe highest degree of cleaning is required and a high surfacepreparation cost is warranted. Blast cleaning by wheel ornozzle (dry or wet) using sand, grit, or shot to white metal willresult in high performance of the paint systems due to thecomplete removal of all rust, mill scale, and foreign matter orcontaminants from the surface. In ordinary atmospheres andgeneral use, White-Metal Blast Cleaning is seldom warranted.Meets requirements of NACE Standard #1.

SSPC-SP6, “COMMERCIAL BLAST CLEANING”

The most common type of blast cleaning should be employedfor all general purposes where a high, but not perfect, degree ofblast cleaning is required. It will remove all rust, mill scale, andother detrimental matter from at least two-thirds of the surfacearea. The advantage of Commercial Blast Cleaning lies in thelower cost for satisfactory surface preparation for the majorityof cases where blast cleaning is believed to be necessary. If thecleaning done according to this specification is likely to result ina surface unsatisfactory for severe service, then Near-WhiteBlast Cleaning (SSPC-SP10) or White-Metal Blast Cleaning(SSPC-SP5) should be specified. Meets requirements ofNACE Standard #3.

SSPC-SP7, “BRUSH-OFF BLAST CLEANING”

This method of blast cleaning should be used when theenvironment is mild enough to permit tight mill scale, paint,and minor amounts of tight rust and other foreign matter toremain on the surface. The surface resulting from this method ofsurface preparation should be free of all loose mill scale andloose rust with the small amount of remaining rust serving as anintegral part of the surface. Brush-off Blast Cleaning is notintended for very severe surroundings. It is generally intendedto supplant Power Tool Cleaning where facilities are availablefor blast cleaning. Meets requirements of NACE Standard #4.

SSPC-SP10, “NEAR-WHITE BLAST CLEANING”

This type of blast cleaning is generally employed for allgeneral-purpose applications where a high degree of blastcleaning is required to remove all rust, mill scale, and otherdetrimental matter from at least 95% of the surface area.Exposures include high humidity, chemical atmosphere,marine, or other corrosive environments. Blast cleaning tonear-white metal was developed to fill the need for a grade ofblast cleaning beyond that of Commercial (SSPC-SP6) but lessthan White Metal (SSPC-SP5). The advantage of Near - WhiteBlast Cleaning lies in the lower cost for surface preparationthat is satisfactory for all but the most severe serviceconditions. Meets requirements of NACE Standard #2.

* nyb’s standard surface preparation is a high-pressure chemical wash followed by SSPC-SP2 - Hand Tool Cleaning or SSPC-SP3 - Power Tool Cleaning as required.

Form 607 GAW

CORROSION RESISTANCE OF FRP FANS

ENGINEERING LETTER 18The New York Blower Company ● 7660 Quincy Street, Willowbrook, Illinois 60521-5530

INTRODUCTION

Process applications involve a wide variety of corrosive gas streams.Selecting the best materials of construction for air handlingequipment can be difficult. This Engineering Letter providesinformation about the corrosion resistance of the resins used tomanufacture standard nyb FRP fans.

GAS STREAM TYPES

Fumes are the dry vapors evolved from acids, solvents, etc. Anexample is the dry acid vapor scavenged from a process using acid.As a generalization, fumes are not as corrosive as aerosols.

Aerosols are suspensions of liquids or solids in a gas stream. For thepurpose of this discussion, aerosols are considered as being wet.Water as fog is an example of an aerosol. Another example is themist of acid present in air scavenged from a process where acid isbeing used as a spray wash. As a generalization for the purpose ofestimating corrosiveness, aerosols in fan-driven systems can beconsidered as being dilute concentrations of the chemicalscomposing the aerosols.

An example of the distinction between fumes and aerosols is asystem where sulfuric acid fumes are collected by hoods andscrubbed. The dry fumes entering the scrubber could be quiteconcentrated but have a relatively mild corrosive effect on the hoodand duct material. On the other hand, the wet gas down stream fromthe scrubber could be quite dilute but more corrosive because of thescrubber’s converting the fumes to an aerosol.

CORROSION-RESISTANCE GUIDE

The corrosion-resistance guide presented in this Engineering Letteris adapted from the literature published by the manufacturers of theresins used in the standard construction of nyb FRP fans. The guideprovides data for aerosols being handled in fan-system gas streams.Data for chemicals that are potentially damaging as aerosols aremarked “fumes only”.

Where the user is unsure of the nature of the chemicals involved, orof the corrosive effect of the combination of chemicals involved, it isadvisable to insert resin test coupons, as well as coupons of possiblealternate materials of construction, into the gas stream forobservation.

Derakane® is a registered trademark of Dow Chemical CompanyHetron ® is a registered trademark of Ashland, Inc.Nexus ® is a registered trademark of Precision Fabrics Group, Inc.

TYPES OF RESIN USED BY nyb

All polyesters and vinyl esters are corrosion resistant to somedegree. The resins used by nyb are at the highly resistant end ofthe scale of corrosion resistance. (The opposite end from general-purpose resins, sometimes called “boat resins”.)

Standard FRP construction consists of Hetron® 92FR, a goodchemical-duty grade of polyester made by Ashland ChemicalCompany, for the housing and all non-rotating parts. Wheels aremade of Derakane® 510A40, a top-quality chemical-duty vinylester made by Dow Chemical Company. (See separateEngineering Letter for a full description of nyb resins.)

All-vinyl ester airstream construction consists of all FRP partsbeing made from Derakane 510A40.

SURFACE VEIL

Standard nyb construction does not include the use of surfaceveil. Years of service prove this construction to be cost-effectiveand functionally successful.

However, the general approach to the design of most FRPchemical-process equipment, such as storage tanks, is to usesurface veil. Therefore, the ASTM standard specification for FRPFans and Blowers, D4167, calls for a layer of surface veil on theinner surface of the fan housing. If required, nyb will construct afan with synthetic veil on the housing airstream surfaces to meetASTM D4167.

Synthetic veil such as Nexus®, a polyester veil made by PrecisionFabrics Group, Inc. and used exclusively by nyb, is advantageousin helping to build a relatively thick surface layer (approximately10 mils) that protects the glass structure from attack by chemicalsthat are particularly aggressive toward glass. Where the use ofsynthetic veil is advisable, the corrosion-resistance guide is sonoted. For more information on surface veil and its uses, refer toEngineering Letter 21.

CUSTOMER RESPONSIBILITY

nyb will provide quality FRP construction using either of theabove resin types as specified by the customer. This EngineeringLetter and any discussions between nyb representatives and thecustomer should not be construed as a warranty of materialsuitability for a particular application. The system designer shouldhave sufficient knowledge of, or experience with, the applicationto select the appropriate resin or alternate material.

Acetaldehyde R*n R*n N R* R R R* R Calcium Hydroxide TV RV N N R* R R RAcetic Acid R R N R* R* R R* R Calcium Hypochlorite X X N N N R* N RAcetic Acid, Glacial N N N R* R* R R R Calcium Sulfate† R R N R* R* R* R* R*Acetic: HCl: H2O R R N N N N R* R* Carbon Dioxide R R R* R R R R* RAcetic Acid: HCl R R N N N N R* R* Carbon Disulfide Vapor R* R* R* R R* R* R* R*Acetic: H2O2 R R N N N N R* R* Carbon Tetrachloride R R R* N R* R R RAcetic Anhydride N N N R R* R R* R Cascade Solution R R R* R* R R T R*Acetone N R R R R R R R Chlorine Dioxide T R N R* N N T RAcetyl Chloride N R N N R R R* R Chlorine Gas, Dry RV RV R* N R* R* R RAcetylene T T R* R R R R* R* Chlorine Gas, Wet RV RV N N N N N RAcrylic Acid R R N T R R T N Chloroacetic Acid R* R* N N R* R R* RAcrylonitrile N R³ R* R* R R R* R* Chlorobenzene N N R* R* R* R* R* RAerosol®, Wetting Agent R R T T T T T T Chlorofluorocarbon R R T T T T T TAlmond Oil R R R* R R R R* R* Chloroform N R* N R* R R R RAluminum Acetate† T R T R* R R R R Chlorosulfonic Acid N N R* R* N N R* RAluminum Chloride (dry)† T R N R* N R* R* R Chlorotoluene N R* N N N R* R* NAluminum Fluoride† RVn RVn N R* N N R* R* Chrome-Plating Bath R* R N N R* R* N RAluminum Sulfate† R R N N R R R* R Chromic Acid R R N R* R* R R* RAmmonia R* R R* R R R R* R* Chromic Acid + Sulfuric R R N N N R* R* RAmmonium Carbonate† R* R R* R R R R R Citric Acid R R N R* R R R* RAmmonium Chloride† R R N N R R R* T Cooling Towers R R T T T T T TAmmonium Hydroxide R*V RV N R* R R N T Copper Chloride R R N N N N N R*Ammonium Nitrate† R R N N R R R* T Copper Cyanide R R N N R R N RAmmonium Persulfate R R N R* R R N T Copper Nitrate R R N N R R N RAmmonium Sulfate† R R N R* R* R* R* R* Copper Oxychloride R R N N N N N R*Ammonium Sulfite R R N N R* R R* R Copper Sulfate† R R N N R R N RAmyl Acetate R n R N R R R R R Cyclohexane R R R* R* R* R* R R*Amyl Alcohol R n R n R* R* R R R R DDT, Insecticide Solution R R T T T T T TAniline N N N N R R R* R* Dichlorobenzene N N R* R* R* R* R* RAniline Sulfate R R N N T T T T Dichloroethylene N N N R* R* R* R* R*Anthracene Oil R R T T T T T T Dichlorophenoxyacetic R* Rn T T T T T TAntimony Pentachloride R R N N N N T T Dichloropropane N N T T T T T TAntimony Trichloride† R R N N N N T T Dichlorotoluene N Rn N N N T T NAqua Regia (HNO3 - HCl) R R N N N N T N Diesel Fuel R R R R R R R* R*Arsenious Acid R R N N R R T R* Diethyl Ether N N R* R* R* R* R* R*Barium Carbonate† R R R* N R R T R* Diethyl Glycol R R R* R* R R R* R*Barium Chloride† R R N N R R RT R* Diethyl Ketone N N T T T T T TBarium Hydroxide† N R n N N R* R* R* R* Diethyl Maleate: Water R R T T T T T TBeer R n R N R R R* R R Diethylbenzene T R N T N T T TBenzaldehyde N R* N R* R* T R Diisobutyl Ketone N N T T T T T TBenzene N N N R* R R R* R* Diisobutylene T R T T T T T TBenzene, Sulfonic Acid R R N N R* R* T R* Dimethyl Sulfide N N T T T T T TBenzoic Acid R R N R* R R R R* Dimethyl Sulfoxide R* R* T T T T T TBenzoyl Chloride N R N T R R T T Dimethylformamide R* R* N R R* R* R* TBenzyl Alcohol N N R* R* R R R R* Dimethylamine R* R* T R R R T TBenzyl Chloride N R* N N N R* R* N Dipropylene Glycol R R T T T T T TBoric Acid R R N R* R R R* R Divinyl Benzene T R T T T T T TBromine, Dry Gas R R N N N N T R Dodecene N R T T T T T TBromine, Moist Gas R R N N N N N R Dodecylbenzenesulfonic R R T T T T T TButane T T R R* R R R R* Acid: H2SO4: H2O: oilButyl Acetate R* R* N T R R T T Esters, Fatty Acid R R N R R* R* R* RButyl Alcohol T R n R* R* R R T R* Ethanol Chloride N R T T T T T TButyl Hypochlorite X X T T T T T T Ether R* R* R* R* R R R* R*Butylene Glycol R R T T T T T T Ethyl Acetate N R R* R* R R R* R*Butylene Oxide N N T T T T T T Ethyl Alcohol R R R* R* R R R* RButyric Acid R R T R* R R R* R Ethyl Acrylate N N N R* R* R* R* RCalcium Chlorate† R R T T R R R* R Ethyl Benzene N Rn N R* N R* R* RCalcium Chloride† R R N R* R* R* R R Ethyl Chloride R* R* R* R* R R R* R*† These compounds are normally solids; considered here as being water solutions. R - Recommended 120°F. maximumR* - Recommended for fumes only. Care must be taken to prevent formation of condensate on wheel or in housing N - Not recommendedT - Test data not available V - Surface Veil required D - Double layer of surface veil required X - Consult New York BlowerAerosol ® is a registered trademark of American Cyanamid Co.

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CORROSION-RESISTANCE GUIDE TO FUME AND AEROSOL CONTAMINATED AIRFOR nyb FRP CONSTRUCTION AND ASSORTED METALS

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Ethyl Ether N R* R* R* R* R* R* R* Linseed Oil R R R* R* R R R* R*Ethylene Chlorohydrin R R R* R* R* R* R* R* Lithium Carbonate N RV T T T T T TEthylenediamine Tetra N Rn T T T T T T Lithium Chloride† R R R* N R R R R

Acetic Acid Lithium Hydroxide N RV N N R* R* R* R*Ethylene Dibromide N R* N R* R* R* R* R* Lithium Hypochlorite X X T T T T T TEthylene Dichloride N N R* R* R R R R* Magnesium Carbonate† R R N R* R R R* R*Ethylene Glycol R R R* R R* R* R* R Magnesium Chloride† R R N N N R* R* REthylene Oxide N N R* R* R* R* R* R Magnesium Sulfate† R R R* R* R* R* R R*Fatty Acids R R N R R* R* R* R Malathion R R T T T T T TFerric Chloride† R R N N N N N R* Maleic Acid R R N R* R* R* R* R*Ferric Nitrate† R R N N R R N R* Mercapto Acetic Acid N N T T T T T TFerric Sulfate† R R N N R* R R* R Mercuric Chloride† R R N N N N N R*Ferrous Chloride† R R N N N N N R* Mercurous Chloride† R R N N N N N R*Ferrous Sulfate† R R N R* R* R* R* R* Mercury R R R* N R R R* RFlue Gas, (wet) R R N R* T R* T R Methacrylic Acid R R T T R R T NFluoboric Acid RV RV R* N R* R R* R Methyl Alcohol R* R* R* R* R R R* RFluorine Gas N RDn N R R R R R* Methyl Bromide R* R* R* N R* R* T TFluosilicic Acid† RV RV N N N R* R R* Methyl Chloride Rn R N N R R R* R*Fluosulfonic Acid† R* R T T T T T T Methyl Ethyl Ketone N N R* R* R R R* RFormaldehyde R R R* R* R R R R* Methylene Chloride N N R* R* R* R* R* RFormic Acid R R N N R* R R* R Mineral Oil R R R* R* R R R R*Fuel Oil R R R* R* R R R* R Monochloroacetic Acid R* R* N N R R T RFungicides R R T T T T T T Monochlorobenzene N N R* R* R* R* R* R*Fungus, 95% Relative Rn R T T T T T T Monoethanolamine N N R* R* R R R* R*

Humidity Naphtha R R R* R R R R* R*Furfural R*n R*n R* R R* R* R* R* Naphthalene R R R* R R R R* R*Gasoline R R R R R R R R Nickel Chloride† R R N N R* R R* RGasoline, Aviation R R R R R R R R Nickel Nitrate† R R N N R R R* R*Glycerine R R R* R R R R R Nickel Sulfate† R R N N R* R* R* R*Glycolic Acid R R N R* R* R* R* R* Nitric, Red Fuming N R* N R R R N R*Heptane R R R* R R R R* R Nitric Acid R R N N R R N RHexane R R R* R R R R* R Nitrobenzene N N R* R* R* R* R* R*Hexachlorocyclopentadiene R R T T T T T T Nitrogen R R R R R R R RHexachloroethane N N T T T T T T p-Nitrotoluene Sulfonic Acid R* R T T T T T THexamethylenetetramine R R T T T T T T Nitrous Acid Rn R N N R* R N THydrazine N N T T R R T T Nut Oil, Ground R R R* R R R R* R*Hydrochloric Acid Fumes RD RD N N N N R* R Oakite Stripper SA® R R T T T T T THydrochloric Acid + Free Cl2 RD RD N N N N R* R Oleic Acid R R R* R R R R RHydrocyanic Acid R R N R R* R R R Oleoparathion R R T T T T T THydrofluoric Acid R*D RD³ N N N N R R* Oleum N N R* R* R* R* N RHydrogen R R R R R R R R Olive Oil R R R R R R R* RHydrogen Bromide R R N N N N N R* Oxalic Acid† R R N R* R R R* R*Hydrogen Chloride R R N N N N R* R Oxidizing Gases R R T T T T T THydrogen Fluoride R*V RVn N N N N R R* Ozone T R* N R* R* R* R* THydrogen Peroxide R R N R R R R* R Palmitic Acid R R N R* R* R R* R*Hydrogen Sulfide R R R* R* R* R R* R Parathion, Wet R R T T T T T THydroxyacetic Acid R R N R* R* R* R* R* Perchloroethylene R* R R* R* R R R R*Hypochlorous Acid R R N N N N N R* Perchloric Acid Rn R* N N N N N R*Insecticides R R T T T T T T Petroleum Ether R R R* R R R R RIodine R* R N R N N R* R Phenol N N R* R R R R R*Iron Perchloride† R R T T T T T T Phenol, Sulfonic Acid N N T T R R T TIsobutyl Alcohol T R T T T T T T Phosphate Salts† R R N R* R R R* RIsopropyl Alcohol T Rn R* R* R R R* R Phosphoric Acid R R N N R* R R* RIsopropyl Amine T R T T T T T T Phosphorous Acid R R R* R* R R R RKerosene R R R* R R R R* R* Phosphorous Oxychloride N T N R* N N R* R*Lactic Acid R R N R* R* R N R* Phosphorous Trichloride N N R* N R R T R*Lead Acetate† R R N N R* R* R* R* Phthalic Acid R R N R* R* R R* R*Leather Dyeing & Finishing R R T R R R T R* Phthalic Anhydride R R R* R R R R R† These compounds are normally solids; considered here as being water solutions. R - Recommended 120°F. maximumR* - Recommended for fumes only. Care must be taken to prevent formation of condensate on wheel or in housing N - Not recommendedT - Test data not available V - Surface Veil required D - Double layer of surface veil required X - Consult New York BlowerAerosol ® is a registered trademark of American Cyanamid Co.

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Picric Acid in Alcohol R* R* N N R* R* N R* Sulfamic Acid R* R* N N T R* T TPolychlorocyclohexane R R T T T T T T Sulfated Detergents R R N N T T T TPolyvinyl Acetate Emulsions R R N T R R T T Sulfate Liquors R R N N R* R* R* R*Polyvinyl Alcohol R R N T T T T T Sulfite Liquors R R N N R* R* N RPolyvinylidene Chloride R R T T T T T T Sulfur, Wettable, Fungicide R R N R R R R RPotassium Bicarbonate† R*V RV R* N R* R* R* R* Sulfur Dichloride R R N N N R* N R*Potassium Carbonate† R*V RV R* N R* R* R* R* Sulfur Dioxide R* R R* R* R* R* R* R*Potassium Chloride† R R R* N R* R R R Sulfur Trioxide Rn R R* R* R* R* R* R*Potassium Cyanide N R*n R* N R R R* R* Sulfuric Acid R R N N N R* R* R*Potassium Dichromate† R R R* R R R R* R* Sulfuric Acid: Phosphoric R R N N N R* R* R*Potassium Ferrocyanide† R R R* R* R* R* R* R* AcidPotassium Hydroxide N RDn R* N R R R R Sulfuric: Nitric Acids R R N N N R* N R*Potassium Nitrate† R R N R* R* R* R* R* Sulfurous Acid R* R* N R* N R* N R*Potassium Permanganate† R R R* R R R N R Sulfuryl Chloride N R T R* T T R* RPotassium Persulfate† R R T T T R* T R Sweet Oil R R R R R R R* RPotassium Sulfate† R R R* R* R R R R Tannic Acid R R N N R* R R* R*Propionic Acid Rn R T T T R* N R Tar Camphor R R R* R R R R* R*Propionyl Chloride N T T T T T T T Tartaric Acid R R N R* R R R* R*Propylene Glycol R R R* R* R* R* R* R* Tetrachloroethane T R* R* N R* R T RPulp and Paper Mill RVn RV T T T T T T Tetrachloropyridine T Rn T T T T T T

Blow Down Gases Tetrapotassium R R T T T T T TPyridine N N R* R* R* R* R* R* Pyrophosphate†Rayon Spin Bath R R N T N R* T R* Thionyl Chloride N N N N N N R* TSelenious Acid T R N N T T N T Tin, Molten, Fumes R R N N N N T RSewage Treatment R R N R* R* R* T R* Toluene R R R R R R R RSilver Nitrate† R R N N R R N R* Toluene Sulfonic Acid R* R T T T T T TSodium Acetate† R R N R R* R* R* R Tolyl Chloride N RV T T T T T TSodium Benzoate† R R T R T T R* R* Trichloroacetaldehyde N R T T T T T TSodium Bicarbonate† RV RV N R* R R R R Trichloroacetic Acid R R N N N N R* R*Sodium Bisulfate† R R N N R* R* R* R* Trichloroethane N Rn T N R* R* T RSodium Bisulfite† R R N N R* R* R* R* Trichloroethylene N R* R* R* R* R* R RSodium Borate† R R N R* R* R* R* R Trichloromonofluoromethane TV RVn T T T T T TSodium Bromide† R R N N R* R* R* R* Trichlorophenol N T T T T T T TSodium Carbonate† RV RV R* N R R R* R Triethanolamine T R R* R* R* R* R* R*Sodium Chloride† R R N N R* R R R Trimethylene Chlorobromide N N T T T T T TSodium Chloride, pH 10.5 R* R N N N N N R Trisodium Phosphate† R R R* N R* R* R* R

Cl2 Sat. Turpentine N R R* R* R R R R*Sodium Chlorite† R R N T R* R T R* Urea T R R* R* R* R* R* R*Sodium Cyanide T R R* N R R N R* Urotropine R R T T T T T TSodium Dichromate R R R* R* R* R* T R Vinegar R R N R* R R R* RSodium Ferricyanide† R R T R R* R* R* R Vinyl Chloride N N T R* R* R* T R*Sodium Hydroxide† RVn RV R* N R R R R* Vinyl Toluene T R³ T T T T T TSodium Hypochlorite† X X N N N N N R* Waste, Organic, H2O, RVn RVn N N N N N RSodium Nitrate† R R R* R R R R* R* HCl, Cl2 VaporsSodium Sulfate† R R R* R R R R R Water, Deionized R R N R* N R* T R*Sodium Sulfide† Rn R R* N R* R R* R* Water, Demineralized R R N R* R R T R*Sodium Sulfite† R R R* R* R R R* R* Water, Distilled R R N N R* R* N RSodium Xylenesulfonate R R T T T T T T Water, Sea R R N R* R* R* R* RStannic Chloride† R R N N N N N R* Water, Steam Condensate R R R* R* R R R* RStannous Fluoride: N RVn N N N N T T Xylene N R* R* R R R R* R

Hydrofluoric Acid Zinc Chloride† R R N N N R* R* R*Stearic Acid R R N R* R* R R* R Zinc Hydrosulfite† R R N N R R T TStyrene R* R* R* R R R T N† These compounds are normally solids; considered here as being water solutions. R - Recommended 120°F. maximumR* - Recommended for fumes only. Care must be taken to prevent formation of condensate on wheel or in housing N - Not recommendedT - Test data not available V - Surface Veil required D - Double layer of surface veil required X - Consult New York BlowerAerosol ® is a registered trademark of American Cyanamid Co.

Form 607 GAW

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D E S I G N A N D C O N S T R U C T I O N O F n y b F R P F A N SINTRODUCTION

Fiberglass-reinforced plastic (FRP) made from chemical-gradepolyester or vinyl ester resin resists corrosion as well as, or insome cases better than, high-priced materials such as titanium orhigh-nickel alloys. In general, FRP (also known as RTP, orreinforced-thermoset plastic) is widely used in handling thefumes of acids and of many inorganic and organic chemicalswhere service temperatures do not exceed 250°F.

COMMON USES FOR FRP FANS

Potential applications for FRP fans include any process inwhich corrosive fumes must be captured, moved, cleaned, orvented. FRP fans are most often used in fume-scrubber systemswhere the scrubber itself may be constructed of FRP or anexotic alloy, but where FRP is the preferred fan material.Galvanizing and etching processes often have FRP exhausthoods and ducts, and many of the fans used to convey fumesin such systems are also built of FRP. Wastewater-treatmentplants and laboratory exhaust systems are other applicationsfor which FRP fans are being used with increasing frequency.

When FRP is the selected material for an air-handling system, itis logical that the fan also be made of FRP. For example, theacids used in the pickling of stainless steel are necessarily thosethat attack stainless steel. In such a system, the acid-holdingtanks, fume-control hoods, ducts, scubbers, and fans are oftenmade of FRP because FRP resists acid corrosion and costs lessthan metal alloys having comparable resistance.

In summary, FRP fans may be an economical alternative tostainless steel or other metal-alloy fans when corrosion is aconcern and temperature is below 250°F. In addition to theeconomic advantage, FRP fans often provide better performancethan special alloys in handling airstreams that are particularlycorrosive to metals.

COMPARISON OF FRP FANS TO FANS OF OTHERMATERIALS

A comparison of the corrosion resistance and economics offans made of various materials leads to these generalizations:

Coated steel fans vary greatly in the degree of corrosionprotection provided and cost. Coatings run the scale fromlittle-different-than-ordinary machinery enamel to baked-onphenolics applied to sandblasted metal. Costs for coated fansrun from about one-third that of FRP fans for the least-resistantcoated-steel fans to about half the cost for the baked phenolic-coated fans.

Monel ® is a registered trademark of Inco Alloys International, Inc.Hastelloy ® is a registered trademark of Haynes International, Inc.Derakane ® is a registered trademark of Ashland, Inc

Coated fans, regardless of the inherent corrosion resistance ofthe coating, have the potential of coating failure and resultantrapid deterioration of the base metal. Failures occur whencoatings are physically damaged, and when corrosive attackpermeates the coating to attack the metal. Refer to EngineeringLetter 16 for additional information on corrosion-resistantcoatings for fan equipment.

Stainless steel is susceptible to attack by chlorides andresultant physical failure by stress cracking. Residential hot-water heaters are never made of stainless steel because thecombination of small amounts of chlorine in the water, modesttemperatures, and the stresses caused by changes in waterpressure results in rapid failure of the stainless steel.As noted earlier, stainless steel is also much more susceptible tocorrosive attack by most acids than is FRP. Refer toEngineering Letter 14 for additional information on the use ofstainless steel in fan construction.

Costs for fans made of 3 16L stainless vary from about three-fourths that of FRP fans for small Class I fans to almost twicethe price of FRP fans for large Class III fans.

Fans made of Monel®, titanium, and the high-nickel alloysmay be more or less corrosion resistant than FRP dependingon the chemistry and temperature involved. Figure 1 shows theeffect of simultaneously submerging a coupon of a high-nickelalloy (Hastelloy® C-276), 316 stainless steel and FRP(Derakane® 510A40) in a bath of nitric and hydrochloric acid(aqua regia). While the 316 stainless was destroyed and the276 alloy severely corroded, the FRP was untouched. The costof fans made of such alloys is usually several times the cost offans made of FRP.

Figure 1 - High-nickel, stainless, and FRP coupon in bath ofnitric and hydrochloric acid.

Fans made of rigid polyvinyl chloride (PVC) have good all-around corrosion resistance and generally cost less than fansmade of FRP. However, PVC has two significant physicalweaknesses that severely limit its use in fans: PVC becomesquite brittle at temperatures below freezing, and PVC loses itsstrength so rapidly with increasing temperatures that evenordinary summer rooftop operating conditions are marginal.Wheels sag and go out of balance and strike housings. PVC is athermoplastic material that ™remembers ~ its original shape atabout 150°F., and needs to reach only about 300°F. for it tohave the zero strength needed for vacuum forming.

Numerous users have disavowed the use of PVC fans becauseof their experiences with failures resulting from PVC's low-temperature brittleness and high-temperature weakness.

The use of PVC equipment involves some safety considerationsas well. PVC does not burn, but because it is a low-temperaturethermoplastic it collapses early in a fire and will drip moltenPVC. Thus, rather than containing a potential fire within theduct system, as fire-retardant FRP will do, PVC tends toexpand the fire into other areas, even though it is not inherentlycombustible. In addition, PVC releases highly toxichydrochloric acid fumes when exposed to flame even though itis a self-extinguishing material.

PVC, like FRP, is an insulating material and inherently spark-resistant. However, unlike FRP, it cannot be made electricallyconductive to control static electricity.

UNDERSTANDING FRP

The term FRP describes a broad spectrum of fiber-reinforcedplastic materials. For example, cabinets for office machinesmight be made of non-corrosion-resistant plastics reinforcedwith mica and loosely called FRP. However, the FRP used inmaking process vessels and equipment such as fans iscomposed of about 30% by weight of glass or other fibers thathave been given a coating (sizing) to enhance their bondingwith the resin, and about 70% by weight of corrosion-resistantpolyester or vinyl ester resin.

The fibers provide physical strength, and the resin provides thecorrosion resistance and rigidity that make FRP a workablesolid. Sometimes, non-glass-fiber materials are used in FRP toimpart special properties. For example, graphite fibers addtensile strength, and aramid fibers (Kevlar ®) add toughness.But FRP for process equipment usually has glass fibersbecause they are more economical and easier to work with;graphite fibers, for example, are more difficult to handle anddo not bond as well as glass.

Glass fibers are available in a variety of forms, includingcontinuous -strand roving, woven roving, continuous -strandmat, chopped-strand mat, chopped fibers, and milled fibers.nyb uses all of the above except woven roving and continuous-strand mat in the construction of its FRP products.

Continuous-strand roving is used in the chopper guns forspray-molding of non-moving parts such as housings, inletcones, inlet boxes, damper frames, and outlet transitions.

Kevlar ® is a registered trademark of E.I. DuPont De Nemours & Company

Chopped-strand mat, consisting of Type E glass of 1 1/2 ouncesper square foot in weight, is used in hand lay-up of housingsand Fume Exhauster wheel blades among other products.Castings such as FPB and RFE wheels and seal housings aremade with chopped fibers. Milled fibers are primarily used tomake putty for filling cracks, turning sharp angles into smoothradii, and encapsulating wheel hubs.

The corrosion resistance of FRP depends on the resin. Resinsused in FRP for process equipment are formulated for maximumcorrosion resistance, and are consequently two or three timesas costly as those used in everyday products such as boat hullsor auto body parts.

FRP fan manufacturers normally use two types of resin in theconstruction of their products. Polyester is the resin of choicefor non-moving components such as housings and inlet conesbecause it provides excellent corrosion resistance for mostFRP applications at a relatively low cost. Unfortunately, thistype of resin cannot withstand the dynamic stresses inherent inrotating parts such as wheels. Therefore, FRP wheel constructiondictates the use of vinyl ester resins which are much strongerand more flexible than polyester resins.

The strength and elasticity of vinyl ester resins enable FRPwheels to achieve maximum safe speeds comparable to similar-sized steel wheels at 70°F. As with steel and other alloys, thestrength and flexibility of vinyl ester is compromised at elevatedtemperatures, resulting in safe speed derate factors above150°F. Refer to specific product bulletins for maximum safespeeds and applicable derate factors.

FIRE RETARDANCE OF FRP RESINS

Since many FRP applications involve a mixture of combustiblechemicals and air, nyb FRP fans are made of fire-retardantresins. Fire retardance is measured by the ASTM E-84 testmethod, which determines ™flame spread ratings~ by comparingthe rate at which flame spreads when material is fired in along, narrow furnace with flowing air. (The test is also calledthe ™tunnel test~ and is recognized by Underwriters'Laboratory and the National Fire Protection Association.)Completely incombustible materials, such as cement board, arerated zero flame spread. Red oak is used as the comparativevalue of a combustible material and is rated at 100. A flame-spread rating of 25 or lower is considered non-combustible.(Resin systems rated at 25 or less are often referred to as ClassI.)

A flame-spread rating of 50 means that the material willgradually, but steadily, extinguish itself. (Resin systems rated at26 to 50 are often referred to as Class II. Class III and IVdenote less fire-resistant ratings.)

Resins for chemical duty can be made fire retardant by formulatingthe resins to include adequate molecularly bound halogens,such as chlorine or bromine, or by the use of smaller amounts ofhalogens but with the addition of antimony trioxide. The firstmethod is more costly but provides a clear resin that improvesquality control of the product being manufactured since theworkers and the inspectors can see into the finished

product. Antimony trioxide is a white pigment which producesan opaque product that reduces the ability to visually checkproduct quality. Further, antimony additives can reduce thecorrosion resistance and strength of the resin. nyb uses resinsthat are fire retardant without the use of antimony trioxide.

STANDARD nyb RESINS

Housings and other non-moving parts are made with anAshland Chemical Company resin, Hetron® 92FR. This is achlorendic, halogenated polyester resin known for its ease offabrication. It is well suited for high temperature applications,and is highly fire retardant with an ASTM E-84 tunnel testrating of 25 or less, meeting requirements for ASTM E-84Class I performance. This satisfies the most stringent concernfor the containment of ventilation-system fires. Hetron 92FR isa clear resin that does not require the use of antimony trioxide toachieve its rating. The clarity makes it possible to maintainextremely high standards of quality control.

All nyb FRP wheels are made with Derakane 510A40, abrominated epoxy vinyl ester resin manufactured by Ash-landChemical Company. This resin offers the flexibility, lowshrinkage, and excellent secondary bonding necessary towithstand the vibrational stress and fatigue of dynamic loadinginherent in rotating wheels. Derakane 510A40 has an ASTME-84 tunnel test rating of 25, offering Class I fire retardancewithout the use of additives which could compromise its superiortoughness and corrosion resistance.

As with Hetron 92FR, the clarity of Derakane 510A40 enablesthe fabricator and inspectors to locate and eliminate air inclusionsin the laminate, thus maintaining high standards of qualitycontrol of a critical fan component.

Since additives tend to adversely affect a resin's chemicalresistance, nyb FRP products do not contain ultraviolet (UV)inhibitors. These additives have a tendency to inhibit resincure and lose their effectiveness after long exposure to ultravioletradiation. In order to prevent UV degradation, nyb applies acoating to the exterior surfaces of all FRP components. Years ofsuccessful outdoor service prove that this method of protectionis superior to adding UV inhibitors to the resin.

Corrosion resistance is the main concern when selecting aresin. nyb's standard housing resin has very good corrosionresistance to a broad spectrum of corrosive environments.When necessary, the entire FRP airstream may be constructedof the more corrosion-resistant vinyl ester resin. The data oncorrosion resistance to various chemicals, presented inEngineering Letter 18, were derived from tests of these resins.

CONSTRUCTION OF FRP FANS

The fabrication of FRP is similar to the casting of metal. Apattern is used to make a mold for the FRP part. In a fan, theairstream surfaces of the housing should be smooth to minimizeresistance and prevent build-up of airborne contaminants.Thus, male molds are required rather than female ones. Thesmooth outside surface of the mold shapes the inside surface ofthe housing.

Parts made with male molds must be removable, so FRP fanhousings are usually made in two halves with matchingflanges. In larger fans, these two halves are bonded togetherby means of FRP filler between the flanges, as shown inFigure 2. A lamination laid over the joint on the inside of thehousing provides a smooth surface. The joined flanges form aridge that adds rigidity to the housing. The inlet subassembly isbolted into place to allow access for installation of the wheel.

Smaller FRP fan housings are also molded in halves, but theyare typically bolted together as shown in Figure 3. Removingthe inlet side of the housing allows installation or removal of thefan wheel.

Fan wheel construction is also different for large and smallFRP fans. Small wheels, such as nyb's Fiberglass PressureBlower, are made by casting or press-forming in fullyenclosed molds; Figure 4 shows an example. Larger wheels,such as nyb's Fume Exhauster, are made by assembling andbonding molded parts (wheel blades, frontplates, and back-plates) with layers of laminate construction so as to makestrong, smooth joints. See Figure 4. All FRP wheels are oven-cured for several hours to improve physical strength andcorrosion resistance of the FRP laminate.

Figure 2 - Fiberglass Fume Exhauster Figure 3 - Fiberglass Pressure Blower

Figure 4Fiberglass Pressure Blower wheel, upper left-

Fiberglass Fume Exhauster wheel,lower right

Hetron ® is a registered trademark of Ashland, Inc.

Metal parts that are incorporated in the FRP parts, shafts,wheel hubs, and studs, are encapsulated in FRP so no metal isexposed to the gas stream. Shafts are encapsulated in an FRPsleeve that extends through a close-fitting opening in the sideof the housing. (Shaft seals that can be lubricated are availableas an option.) Bolts used to fasten smaller fan housing halvestogether are of 316 stainless steel.

Neoprene foam gasketing is used between bolted housing sub-assemblies and under access doors, inspection ports, and shaftseal assemblies.

FRP fan wheels are permanently bonded to the fan shafts, theshafts encapsulated in FRP, and the assembly balanced as aunit. After the fan is assembled it is test run as a final checkto guarantee smooth operation.

Exterior surfaces of completed nyb FRP fans are coated withgray epoxy enamel.

APPLICABLE STANDARDS FOR FRP FANS

The purchaser of FRP fans should consider the importance oftwo published standards: one, ASTM D4167, covers theconstruction of FRP fans; the other, AMCA Standard 210,describes how fans are to be tested for air performance. TheAMCA Certified Ratings Program is the method by whichmanufacturers certify their products' aerodynamic performance.

The ASTM standard is concerned with the structural reliabilityof the fan. If the fan in a fume-control system fails, the entireprocess may come to a halt. The importance of reliability hasled to development of a standard for FRP fansAmericanSociety for Testing and Materials (ASTM) D4167, StandardSpecification for Fiber-Reinforced Plastic Fans and Blowers.This standard defines minimum specifications for constructionof major fan elements. Here are six of the details:

1. Fan housing construction must conform to the ASTMC582 specification which applies to all FRP processequipment. (nyb standard construction with optional veilcomplies with ASTM C582.) The same resin must beused throughout the housing unless the manufacturer anduser agree to use different resins in different layers of thelaminate. (nyb does not back up the corrosion barrierwith less costly resin. nyb uses premium quality resinthroughout.) The structural rigidity of the housing (or aprototype) is tested by running the fan with the inlet closedand the outlet open. Inward flexing may be no greater than0.5% of the fan-wheel diameter.

2. Fasteners, hubs, and shafts exposed to corrosives must beeither corrosion resistant or encapsulated in a material thatis.

3. The ASTM standard prohibits the use of additives in fanwheel resins that obscure visual inspection of wheelparts, including the use of antimony trioxide. nyb standardconstruction provides fire retardance without the use ofadditives.

4. Safe wheel operating speed is determined either by pastexperience or by destructive testingi.e., running the fanwheel at increasing speeds until it fails, and applying asafety factor to the failure speed.

5. Spark resistance. FRP is spark-resistant in the sense thatcontact of FRP parts does not generally produce sparks.However, FRP fans handling dry air can developelectrostatic charges on wheel and housing surfaces becauseFRP is a non-conductor. Still, an FRP fan can be madespark-resistant by incorporating graphite flakes in thewheel and housing airstream surfaces to make themconductive, and grounding the surface layers of thehousing as shown in Figure 5. ASTM D4167 definesacceptable resistivity as no greater than 100 megohmsbetween all points on the airstream surfaces and ground.

6. Dynamic balance is achieved either by balancing thewheel-shaft assembly as a separate unit or by balancingthe wheel once it is installed in the fan (nyb does both).Unbalance is corrected by adding FRP weights.

The Air Movement and Control Association (AMCA) CertifiedRatings Program is concerned with accurate performance ratings.The manufacturer submits published performance ratings toAMCA and fans for test in the AMCA Laboratory. Deviationsare determined by plotting actual performance against the fan'scataloged performance .

Manufacturers displaying the AMCA Certified Ratings Sealon their products, and in their literature, have agreed to asystem of check testing in the AMCA Laboratory. If a productfails to perform within the tolerances specified by the program,the manufacturer must either republish the literature with correctratings or republish without the seal.

Form 507 DJK

Figure 5 - A graphite-impregnated FRPfan for spark resistance.


ACCESSORIES AND CONSTRUCTION MODIFICATIONSFOR FRP FANS

INTRODUCTION

The applicability of corrosion-resistant FRP fans to a widerrange of applications is enhanced through the use of accessoriesand construction modifications. The purpose of this EngineeringLetter is to provide supplemental information concerningaccessories and modifications that are unique to FRP fans.

ACCESSORIESSHAFT SEALS are used where the standard close-clearanceshaft opening is not deemed to be adequate. (Standardconstruction on nyb FRP fans have shaft openings fitted withTeflon® membranes that have shaft holes 1/32" larger than theFRP shaft sleeves.)

nyb’s standard shaft seal for FRP fans utilizes a pair of Viton®

lip seal elements pressed into an FRP casing. As an option,Teflon shaft seal elements can be provided for more corrosiveapplications. The seal assembly is secured to the fan housingwith 316 stainless steel studs. The heads of the studs areencapsulated in FRP to eliminate exposure to airstreamcorrosives. See Figure 1.

Because the seals must ride on a smooth, heat-conductivesurface, the standard construction of the shaft encapsulated inFRP is not suitable. Therefore, the seal assembly includes thesubstitution of a 316 stainless steel sleeve for the standard FRPsleeve. As an option, Hastelloy® C-276 sleeves are availablefor those cases where the corrosive environment makesstainless steel unacceptable.

The seal assembly is lubricated with “Never-Seez®,” a graphitecompound.

Teflon® and Viton® are registered trademarks of E.I. DuPont de Nemours & Company.Hastelloy ® is a registered trademark of Haynes International, Inc.Hetron® is a registered trademark of Ashland, Inc.

Seals are recommended wherever corrosive or toxic gases arebeing handled, or when outside air is to be kept from enteringthe fan and contaminating a process. It is difficult to predictthe conditions that increase leakage into or out of the fanaround the shaft opening. However, as a general rule, higherpositive or negative pressure differentials will result in greaterleakage.

OUTLET DAMPERS are designed to bolt directly to theoutlet flange on FRP fans. RFE and FPB dampers are round,with one blade. FE and GFE dampers are rectangular, withparallel blades, and are available for MP fans only. See Figure 2.

Casings and blades are constructed of Hetron® 92 FR, asstandard, with Derakane® 51 0A40 available as an alternate.All damper parts are constructed of FRP except the 316 stainlesssteel control quadrant and hardware, and the corrosion-resistant,injection-molded bearings.

Damper casing halves are bolted together to allow for easyreplacement of damper vanes and bearings. All componentscan be disassembled except vanes from rods.

Never-Seez ® is a registered trademark of Bostik.Derakane® is a registered trademark of Ashland, Inc.

Figure 1 - Photo of FRP shaft seal mounted and diagramillustrating lubricated lip seal elements.

Figure 2 - Three types of FRP outlet dampers asmanufactured by nyb.

INLET BOXES are used to accomplish a 90° turn at the faninlet when space is limited. Fan applications typically involveless than ideal connections between the fan and the process.When the connections cause other than straight, uniform flowinto the fan inlet, the fan suffers performance losses beyondthose determined by ordinary duct-resistance calculations orpressure drop measurements. (See Engineering Letter 5 for adescription of the effects of inlet connections.) Therefore, it isadvantageous to use nyb test-rated inlet boxes to reduce flowlosses, and to make those losses predictable for inclusion insystem design calculations. See Figure 3.

Inlet Boxes are available for Fume Exhausters and General-Purpose Fume Exhausters. See Figure 4.

Construction of FRP inlet boxes is similar to that of FRP FumeExhausters. Standard construction is with Hetron 92FRpolyester resin, with Derakane 510A40 vinyl ester resinavailable as an option. Inlet boxes are made in two sectionsbolted together with 316 stainless steel hardware.

THREADED FRP DRAIN with PVC plug, 1" npt, is bondedto the lowest point in the housing scroll.

COMPANION FLANGES are available with FPB and RFEfans for those applications where a flexible or slip connection tothe fan inlet and/or outlet is required. Companion flanges arecommonly used on fans furnished with vibration isolation.

INSPECTION PORTS are used for periodic maintenancechecks on the wheel and the housing interior. They areavailable on all FRP fans, and are located on the drive sidehalf of the housing (GFE and FE fans) or the inlet side half ofthe housing (FPB and RFE fans), at either the 2 o' clock or the10 o’clock position, opposite the fan discharge.

RAISED BOLTED CLEANOUT DOORS are available on GFEand FE fans. They are located above the fan centerline at either the 2o’clock or the 10 o’clock position, opposite the fan discharge.

OUTLET TRANSITIONS provide for a rectangular-to-roundtransition on the outlets of various GFE and FE fan sizes. Theyare available on GFE and FE Sizes 18 through 36 and 48 (MPfans only). The I.D. of the round outlet is equal to that of thefan inlet, and also to the transition length.

MODIFICATIONS

ALL-VINYL ESTER AIRSTREAM provides increasedresistance to certain corrosives. Engineering Letter 18 providesdata for the corrosion resistance of the standard constructionand of the all-vinyl ester construction.

Standard construction uses vinyl ester resin for wheels. Allother FRP parts are made of polyester resin. When an all-vinylester airstream is specified, parts normally made of polyesterare made of vinyl ester. See Engineering Letter 19 for moredetails.

SURFACE VEIL is used to reinforce the surface layer ofresin for added resistance to specific corrosives or to meet thespecification of ASTM D4167. Veil may be applied to just thewheel, or to just the housing, or to the entire airstream. nybuses a synthetic surface veil that is described in detail inEngineering Letter 21.

GRAPHITE IMPREGNATION of the final resin coat onairstream surfaces provides for static grounding. Thisimportant modification allows the fan to handle gas fumes thatare not only corrosive but also potentially explosive.

FRP is inherently non-sparking and the electrical resistance ofFRP may be considered infinitely high since it is essentially anon-conductive or non-metallic material. Because FRP is non-metallic, the physical contact of two FRP parts or a metallic

High, unpredictable effect on fan selection andsystem performance.

Minimum, calculable effect on fan selection andsystem performance.

Figure 3 - Inlet Connections

Figure 4 - FRP Fume Exhauster with Inlet Box

part with an FRP part will not produce a spark. However, FRPdoes have the tendency to hold a charge of static electricity.This charge can be generated by a dry gas or airstream passingover FRP. The fan can ultimately become a capacitor capable ofdischarging high-voltage, low-amperage sparks.

The static electricity or charge which builds up on theairstream surface of the FRP part must be eliminated inapplications where the fumes are potentially explosive. This canbe accomplished by making the surface electricallyconductive, providing an electrical path to dissipate the relativelylow-current static charge.

STATIC GROUNDING - FRP fans can be effectively groundedfor the removal and control of static electricity by incorporatinggraphite in the airstream layer of resin. See Figure 5.

The proper application of the graphite-resin coat is critical ifstatic grounding is to be achieved. Airstream and relatedsurfaces are coated with a mixture of graphite flakes and resin toform a smooth, continuous graphite surface. FPB, RFE, andnon-rotatable GFE and FE fans are furnished with contactswhich are imbedded in the graphite layer to accommodategrounding straps made of twisted, bare copper wire. The strapsare attached to the fan base on FPB and RFE fans and to inletside angles on the large Fume Exhausters. Rotatable GFE andFE fans do not require grounding straps. These fans arecompletely grounded to the pedestal through the mountingstuds on the housing. This design effectively grounds theairstream to the steel base of the fan. However, it is essentialthat the customer ground the fan base at theinstallation.

GROUNDING FEATURES - Surface resistivity of not morethan 1 megohm from any point on the airstream to ground isgenerally considered adequate. nyb’s process of staticgrounding by graphite impregnation provides surface resistivitywell below the 1-megohm figure.

Tests of nyb FRP fans equipped for static grounding indicatethat there is sufficient conductivity through the bearings toeliminate the need for supplemental brush-type contacts toground the wheel and shaft assembly for most applications.However, the burden of determining whether this is the casefor a particular installation and lubrication system rests withthe customer.

Static grounding by graphite impregnation does not interferewith the corrosion-resistant properties of the fan. Graphite isextremely corrosion resistant. However, the addition of thegraphite makes the surface softer than normal and prevents thenormal checking of the surface for Barcol-hardness readings.

FRP fans are often the best alternative for those applicationswhich require the handling of explosive, as well as corrosivegas fumes. However, care must be taken to realize that therecan be no guarantees against possible sparking or ignition insuch airstreams. All aspects of the application, the systemcomponents, and even the potential for sparks resulting from“tramp” or “foreign” elements in the airstream must beconsidered to ensure the safety of the installation.

FLANGE-DRILLING PATTERNS for round inlet andround outlet flanges are in accordance with the NationalBureau of Standards Voluntary Product Standard PS 15-69.This drilling pattern was developed by members of the FRPindustry for FRP ductwork and specifies bolt hole diametersappropriate for bolting FRP ducts to FRP fans.

nyb FRP fans that have both round inlets and round outlets arealso available with flanges drilled to ANSI 150. Because ANSI150 is intended for bolting together heavy metal pipe, it usesbolts that are unnecessarily large for FRP. Although nybcharges the same for drilling to PS 15-69 or ANSI 150, thecost to the user can be substantially different. Flanges areusually fastened together with corrosion-resistant alloy bolts,nuts, and washers. The cost difference between the sizesrequired for PS 15-69 and ANSI 150 can be significant. Forexample, a 12" inside-diameter PS 15-69 flange would have7/16" diameter holes for twelve 3/8" bolts. An ANSI 150flange would have 1" diameter holes for 7/8" bolts. Thedifference in cost can be $50 or more per flange for 316stainless steel hardware and much more for higher-alloyhardware.

Since PS 15-69 and ANSI 150 drilling patterns only pertain toround flanges, they do not apply to FE and GFE outlet flanges.Therefore, nyb has developed a standard for drilling rectangularoutlet flanges which provides holes drilled on 4" centers,straddling the flange centerlines.

Form 507 DJK

Figure 5 – FRP Radial Fume Exhauster with graphiteimpregnation and copper grounding straps.

Page 3


S U R F AC E V E I L F O R F R P F AN SINTRODUCTION

This Engineering Letter has several functions: to describe nybsurface veil characteristics, define the purposes of surface veil,explain the relationship between surface veil and ASTMD4167, detail the specific corrosive agents that require adouble layer of veil, and describe the special constructionrequirements involving hypochlorite applications.

SURFACE VEIL CHARACTERISTICS

The synthetic surface veil used exclusively by nyb is Nexus®. Itis a non-woven formed fabric produced from Dacron® 106homopolymer. This binder-free polyester fiber has an apertured(perforated) design that provides the necessary flexibility forthe fabrication of fans. Each layer of surface veil contains about90% resin and 10% veil material and is applied at a minimum of10 mils.

PURPOSES OF SURFACE VEIL

One purpose of surface veil, also referred to as surfacing mator tissue, is to prevent protrusion of the chopped-strand matglass fibers to the surface which could allow chemical wickingto occur.

A second, but equally important, purpose is to providereinforcement to the surface layer of resin to prevent crackingand crazing.

Finally, the addition of surface veil allows nyb fans to meetthe requirements of ASTM Standard D4167.

Dacron® is a registered trademark of E.I. DuPont de Nemours & Company, Inc.Nexus ® is a registered trademark of Precision Fabrics Group, Inc.Lupersol ® is a registered trademark of Elf Atochem North America, Inc.

ASTM STANDARD D4167

The American Society for Testing and Materials (ASTM) D4167, Standard Specification for Fiber-Reinforced Plastic Fansand Blowers defines the basic guidelines for the constructionof FRP fans handling corrosive fumes. One of thespecifications within D4167 is that the laminate construction bein accordance with another ASTM standard, C582. Thatstandard specifies that the working surface (the surface to be incontact with corrosives) of the laminate consist of one layer ofsurface veil backed by two layers of chopped-strand mat orequivalent from a chopper gun, followed by the structurallayers. Therefore, in order to comply with ASTM D4167, allFRP fans must be furnished with at least one layer of surfaceveil on all housing surfaces.

CORROSIVES REQUIRING A DOUBLE LAYER OFSURFACE VEIL

There are some chemical agents that are aggressive towardglass. For these specific corrosives, nyb’s resin suppliersrecommend the addition of a layer of surface veil for increasedcorrosion resistance. Additionally, in those applications wherethe corrosive agent is extremely aggressive, a second layer ofveil is required. The corrosion-resistance guide found inEngineering Letter 18, Corrosion Resistance of FRP Fans,indicates where one or two layers of veil are required. Thecorrosives listed as requiring a double layer of surface veilinclude fluorine gas, hydrochloric acid, hydrofluoric acid,hydrogen fluoride, potassium hydroxide, and varioushypochlorite compounds.

ADDITIONAL REQUIREMENTS FOR HYPOCHLORITEAPPLICATIONS

Applications involving butyl hypochlorite, calciumhypochlorite, lithium hypochlorite, or sodium hypochloriterequire special FRP construction considerations. In addition tothe aforementioned double layer of surface veil, resin suppliersrecommend a substitution for nyb’s standard catalyst andpromoter.

nyb’s standard polyester and vinyl ester resins both use cobaltnapthanate (CoNap) as a promoter and Lupersol®, a methylethyl ketone peroxide (MEKP), as a catalyst. Gas streamscontaining hypochlorites attack CoNap whenever MEKP isused as a catalyst. Therefore, a benzoyl peroxide (BPO)catalyst is recommended for these applications because it

Nexus Synthetic Surface Veil

does not use CoNap as its promoter. The BPO catalyst requiresdimethyl aniline (DMA), which is unaffected by hypochlorites,as its accelerator in lieu of the CoNap.

Special BPO/DMA construction is limited in that it cannot beused for FRP wheel construction. It can only be used to applysurface veil to the wheel. All other FRP components can beconstructed using this special catalyst/accelerator system.

In addition, note that due to reactivity between BPO catalystsand graphite, which reduces the graphite’s conductivity, static

grounding by graphite impregnation is not an available optionwhen used in conjunction with a BPO catalyzed resin.

CUSTOMER RESPONSIBILITY

This Engineering Letter and any discussions between nybrepresentatives and the customer should not be construed as awarranty of material suitability for a particular application. Thesystem designer should have sufficient knowledge of, orexperience with, the application to select the appropriate resin oralternate material.

Form 607 GAW

ENGINEERING LETTER 22The New York Blower Company ● 7660 Quincy Street, Willowbrook, Ill inois 60527-5530

I N T E G R A L M O T O R S F O R C E N T R I F U G A L F A N SINTRODUCTION

The most common power source for fans is the electric motor.A motor’s service life is largely dependent upon properselection and installation. Since the motor and its controlcircuitry represent a substantial portion of the cost of many fansystems, they deserve careful consideration. This Letterintroduces some of the more important matters for consideration.

SELECTION CRITERIA

The selection of the proper motor is based on numerous criteria.Included are horsepower, service factor, enclosure, ambienttemperature, phase and voltage, speed, and efficiency.

Horsepower. If all air-handling systems had exactly the samevolume/pressure relationship the designer anticipated, allmotors could be selected merely to cover the fan brake horse-power (BHP) calculated. However, system design usuallyinvolves some estimating, and systems are not always installedexactly as intended by their designers.

With all centrifugal fans, the fan speed must be increased tohandle the desired volume when the system resistance is higherthan anticipated, creating a substantially higher fan BHPrequirement. For radial and forward-curved wheels, if the systemresistance is lower than anticipated, fan BHP will increasewith the greater volume of air being handled. Refer to Figure 1.

The major difference in the BHP curve for backwardly-inclined fans is its “non-overloading” characteristic. Figure 2illustrates a BHP curve that reaches a peak and then drops offas the volume continues to increase. This makes it possible toselect a motor for the maximum BHP at a given speed withoutfear of overload despite any variance in the volume/pressurerelationship of the installed system. Since BHP varies withchanges in fan speed, the non-overloading characteristic onlyapplies to a given fixed speed.

The fan capacity table (Figure 3) shows the fan BHP for a givenvolume/pressure relationship. However, it is not uncommon tosize the motor for a static pressure 5% to 10% higher thandesign to allow for variances in the installed system. Thesystem designer should also be prepared to reduce fan speed ifresistance is lower than anticipated.

Motors should be selected so that the fan BHP rating for therequired volume and pressure is less than the rated motorhorsepower. The rated motor horsepower is the mechanicalpower available at the motor shaft at full-load speed withoutexceeding the motor’s maximum temperature rise.

4” SP 41/2” SP 5” SP 5 1/2” SPCFM OV

RPM BHP RPM BHP RPM BHP RPM BHP10856 2100 1140 9.29 1194 10.4 1248 11.6 1300 12.811373 2200 1154 9.79 1205 10.9 1257 12.1 1308 13.311890 2300 1167 10.2 1218 11.4 1269 12.7 1318 13.912607 2400 1183 10.8 1232 12.0 1282 13.3 1330 14.612924 2500 1200 11.4 1248 12.6 1296 13.9 1342 15.213441 2600 1217 12.0 1263 13.3 1310 14.5 1355 15.9

Figure 3 - At 12,924 CFM and 5" SP,the BHP required is 13.9. With an additional 10% system

resistance margin (51 /2" SP), the BHP required is 15.2.

Service Factor. Integral open-dripproof and totally enclosedmotors usually have a service factor of 1.15, while explosion-proof motors usually have a 1.0 service factor. When themotor nameplate voltage and frequency are maintained, themotor can be run up to the capacity obtained by multiplyingthe rated horsepower by the safety factor shown on the motornameplate.

For example, a fan in a given system might require 5.0 BHPaccording to original estimates, but minor system changescould increase the demand to 5.25 BHP. In this case, a 5 HPopen motor rated with a 1.15 service factor could still be used(5 HP x 1.15 = 5.75 HP) without detrimental overheating.

Enclosure. The selection of a motor enclosure depends uponambient conditions. Electric motors are air-cooled machinesand their service life depends greatly upon protecting themotor from contaminated surroundings. Basically, all motorenclosures can be divided into two categories: open andtotally enclosed.

OPEN MOTORS - This type is recommended for relativelyclean environments since the ventilating openings permitpassage of external cooling air over and around the motorwindings. Open motors are usually less expensive than otherenclosures.

DRIPPROOF MOTORS - These are open motors withventilating openings so constructed and positioned thatoperation is not hampered when drops of liquid or solidparticles strike the enclosure at any angle from 0° to 15°downward from the vertical axis. The standard insulation isClass B with a 1.15 service factor rating.

WPI AND WPII MOTORS - These are essentially openmotors with vacuum-pressure impregnation (VPI) windingtreatment for moisture resistance and weather protection. WPImotors are equipped with space heaters. WPII motors haveventilating openings arranged so that high-velocity air and/orairborne contaminants blown into the motor during storms orhigh winds can be discharged without entering the internalelectrical parts of the motor. Generally, the weather protectedmotors are only available in frame sizes larger than NEMAstandard and they are less expensive than totally enclosedmotors in those cases.

TOTALLY ENCLOSED - This type is recommended for anyinstallation where dirt or contaminants can collect in or aroundthe motor. They are constructed in a manner that prevents thefree exchange of air between the inside and outside of themotor case, but they are not airtight.

TOTALLY ENCLOSED FAN-COOLED MOTORS - Theseare totally enclosed motors equipped with a cooling fan, orfans, integral with the motor assembly but external to theenclosed parts. These motors should be installed so that theintake of the cooling fan is not blocked or impeded. Thestandard insulation is Class F with a 1.15 service factor rating.

TOTALLY ENCLOSED AIR-OVER MOTORS - Thesespecial-purpose totally enclosed motors are intended for use infan applications where the fan provides sufficient coolingairflow over the surface of the motor. However, they are notself-cooling, so they should only be used when airflow ispresent at or above the velocities necessary for continuousoperation within the rated motor temperature rise.

TEFC SEVERE DUTY MOTORS - These special purposeTEFC motors are intended for use in contaminated environmentssuch as in the paper, metal, or chemical industries. Special featuresinclude cast-iron frame, end brackets, conduit box and fancover, plated hardware, and stainless steel nameplates. Theyare also rated with 1.15 service factors and Class F insulation.Some trade names include “Mill and Chemical,” “Dirty Duty,”“Extra Tough,” and “Chemical Duty.”

TOTALLY ENCLOSED NON-VENTILATED MOTORS -These are basically totally enclosed motors with larger frames todissipate heat, but no cooling fan. Typically offered in thesmaller fractional horsepowers, these motors should only beused in open, well-ventilated areas.

EXPLOSION-PROOF MOTORS - These special-purposetotally enclosed motors are designed to withstand internalexplosions of gases or vapors, and to prevent the ignition ofgases or vapors surrounding the motor. Refer to EngineeringLetter 23, Electric Motor Codes and Standards, for details.

Insulation. Various motor insulation systems are available.The rated temperature for a given insulation classification isthe maximum temperature for sustained operation. Threecommon insulation classes are shown in Figure 4.

INSULATIONNEMAClass

AmbientTemperature*

Hot-SpotTemperature

B 40°C. 130°C (266°F.)F 41° - 65°C. 155°C. (311°F.)H 66° - 90°C. 180°C. (356°F.)

*Note that these ratings apply to 1.0 service factor only.Figure 4

Not all parts of the motor windings operate at the sametemperature. The temperature at the center of the coil is thehottest, and is commonly referred to as “hot-spot temperature.”This hot-spot temperature is used to establish the rating of aninsulation class. The actual temperature is the sum of all theheat-producing factors including the ambient temperature,motor induced temperature rise, and the hot-spot allowance.

Ambient temperatures. Whenever possible it is best to selecta motor with the appropriate insulation for the specific ambientconditions. For example, a TEFC motor with Class F insulationis suitable for ambient temperatures of 40°C. (104°F.) with1.15 service factor or 65°C. (149°F.) with 1.0 service factor. Ifthis same motor is used in an ambient of 75°C. (167°F.)

continuously, the life of the motor will be greatly reduced.

Phase and voltage. Although these are limited to the powersupply available at the installation site, the general rule ofthumb is to use polyphase (three phase) motors of the highestavailable voltage in order to achieve the most economicalequipment and installation costs. Single phase motors typicallycost more than polyphase because of the need for capacitors,centrifugal switches, etc. Higher voltage ratings can reduceinstallation costs by reducing the required electrical line size.

In most U.S. and Canadian industrial sites, the power supplytypically found for the average polyphase motor is 230 or 460volts (U.S.) and 575 volts (Canada) at 60 Hertz (cycles persecond) generation. In many large cities where 120/208 voltnetworks are employed, commercial and small industrial loadsrequire motors rated for 200 volts.

Motors for 2300 volts can be furnished in motor frames 445Tand larger. Because of the cost of starting equipment for thishigher voltage, 2300 volt motors are not generally availablebelow 200 HP.

Single phase motors are available for service on 115/230 voltsfor 3 HP and smaller. Motors up to 10 HP are available for230 volt service in single phase.The standard motor frequencies are 60 and 50 cycles per second,or “Hertz.” The prevailing frequency in the United States andCanada is 60 Hertz. Most of Europe, the Middle East, and thePacific Rim have 50 Hertz service. Many motors specified for50 Hertz will require 380 volts, 440 volts, or 220/380 volts . . .all of which are considered standard by motor manufacturers.Although motors built for 50 Hertz are becoming more readilyavailable in this country, consideration should be given to theaccepted practice of derating 60 Hertz motor speed and horse-power. Ratings can be derated by a factor of .833 (50/60) todetermine the operating characteristics in 50 Hertz service.

For example:60 Hertz - 10 HP, 1800 RPM, 3/60/230/460

50 Hertz* - 8.3 HP, 1500 RPM, 3/50/190-380* Note: This does not apply to single phase or explosion-proofmotors. RPM and Voltage rounded to standard nomenclature.NEMA standards state that motors must be capable of deliveringtheir rated horsepower at a variance of nameplate voltage of ±10% voltage, although not necessarily at the standard ratedtemperature rise. One exception is a motor nameplated as208-230/460 volts. The ± 10% voltage only applies to 230 or460, and thus requires very good voltage regulation for operationin a 208 volt network. Another exception is 60 Hertz motorsderated for 50 Hertz operation.A 208 volt network requires a 200 or 200/208 volt motor.Note that the 200/208 does not mean dual voltage, (as with astandard 23 0/460 rating), but is simply a 200 volt motor ratedand recommended for 208 volt service.The NEMA standard 230/460 volt rating is not generallyrecommended for 208 volt service unless authorized by themotor manufacturer. Motors for use in a 208 volt networkshould be ordered with a 200 volt rating, with windings andnameplate so designed and stamped.

Speed. The general rule of thumb is to select the highestpractical motor speed to reduce the size, weight, and cost of themotor.

However, belt-drive fan applications are usually limited to1800 RPM motors when the horsepower requirements are 25and up. Generally, TS (short shaft) frames are used on larger3600 RPM motors, and these are not well-suited to belt-drivearrangements. Although T frame motors are available for largerhorsepower 3600 RPM motors, they are not standard, so longprocurement lead times and cost can be prohibitive.The majority of electric motors used in fan applications aresingle speed. However, multispeed motors are available ineither single phase or three phase.The motor synchronous speed is expressed as:

120 x FSynchronous RPM = Pwhere: F = supply frequency in Hertz

P = number of poles in motor winding

The actual full load RPM (nominal speed) will be somewhatbelow the synchronous speed. The percentage in speed isknown as the percent slip. Thus, an 1800 RPM (4 pole) motorwith a 2.8% slip will have a full load nominal speed of 1750RPM (1800 - 50 = 1750).The exact slip percentage will vary from one motor size andtype to another. Slip is also somewhat dependent upon load. Apartially loaded motor will run slightly faster than a fullyloaded motor. Since calculating the precise nominal speed foreach application would be impractical, the Air Movement andControl Association (AMCA) has established nominal speeds tobe used uniformly to determine fan performance. See Figure 5.

NOMINAL SPEEDS FOR 60 HERTZ MOTORSNumberof Poles

SynchronousSpeed (RPM)

NominalSpeed (RPM)

2-pole: 3600thru 1 HP 345011/2thru 25 HP 350030 HP and up 3550

4-pole: 1800thru 3/4 HP 17251 thru 20 HP 175025 HP and up 1770

6-pole: 1200thru 3 HP 11505 HPand up 1175

8-pole: 900thru 1/8 HP 8501/2 HP and up 875

Note: 50 Hz motor speeds can be determined bymultiplying the above ratings by .833 (50/60).

Figure 5

Motor Efficiency. The continued increase in energy costs andemergence of energy savings programs have heightenedconcern for electrical usage and motor efficiency. Goodsystem design necessitates the selection of the most efficientmotor for a given application.Motor manufacturers are able to improve motor efficiency byaltering any number of design factors. The use of thinner steellaminations in the stator and rotor core, using better grades ofsteel, more copper in the stator, and more efficient, smallercooling fans are just a few examples.

In an effort to distinguish one manufacturer’s motor fromanother, motor manufacturers use a number of names, such asstandard, high, premium, etc., to qualify published efficiencyvalues. The generally accepted basis for comparison ofefficiency values is the “guaranteed minimum efficiency”based on NEMA recommendations. Motor efficiency can becalculated by the following formula:

746 x HP outputMotor Efficiency =

Watts InputWhen comparing motor efficiencies, the power factor mustalso be considered. At a given efficiency, a higher powerfactor results in a lower current demand. The power factor isthe ratio of real current (current required to run the motor) tothe total current (real current plus the reactive current that createsthe magnetic field). The power factor for a given motorshould be obtained from the specific motor manufacturer, but itcan be calculated by the following formula:Power Factor = Watts Input

Volts x Amps x 1.73** For 3-phase motors only.

SPECIAL CONSIDERATIONSIn addition to the previous selection criteria, there are severalother special considerations that affect proper motor selection.These include high or low voltage, starting times, minimumsheave diameters, heavy cycling, and excessive loading.High or low voltage. Motor service life can be shortenedconsiderably if the motor is operated outside the ± 10%voltage variance range.

With low voltage, motor torque decreases. The motor is thereforeforced to slow down to develop the required torque. Thiscauses increased current draw which creates additional heat inthe motor winding. In addition, at the slower speed ventilationis reduced and heat will not be dissipated as rapidly.High voltage will cause an increase in magnetizing current inthe motor. This causes additional heating in the motor windings.Particularly with older motors, increased voltage can breakdown the motor insulation by breaching its insulating capability.Starting times. Whenever an electric motor is used to drive acentrifugal fan, both the fan’s maximum power demand andthe motor starting torque characteristics must be considered.Where larger centrifugal fans are to be driven by relativelysmall motors, it is possible that the motor will not be capable ofovercoming the fan’s inertia to bring it up to the requiredspeed in a reasonable time. Excessive starting time, generallygreater than 10 to 15 seconds, will raise the temperature of themotor windings to a point where circuit breakers can trip out,or the motor itself can be damaged. The user must be awareof this problem when selecting the fan and motor combination.

The two main factors to be considered are the fan wheel inertia(WR 2 or WK 2) and the starting torque characteristics of themotor. Exact curves of the motor starting torque, as a percentageof full load torque at a given speed, are available from themotor manufacturer.Many fan applications require a fan speed other than a nominalmotor speed, so a belt-drive configuration is used. In thesecases, the WR2 must be corrected to include the effects of thefan shaft and fan sheave.

It is best to consult the fan manufacturer for confirmation ofquestionable fan/motor combinations, i.e. large fans with smallmotors. If the combination has an unacceptable starting time,the solution could be to use a larger motor, damper the fan forreduced load starting, or in some cases consider clutchingsystems so the fan can be brought up to speed without trippingelectrical breakers or damaging the motor.

Minimum Sheave Diameters. Special consideration should begiven to the diameter of drive sheaves used on motors. As belttension must increase to avoid slippage with small diametersheaves, the radial load imposed on the motor bearingbecomes significant. The motor manufacturer can providespecific recommendations for minimum sheave diameters.Some general recommendations are shown in EngineeringLetter 23 - Electric Motor Codes and Standards.

Heavy Cycling. When a motor is started and stopped frequently,heat build-up from the heavy starting current cannot beadequately dissipated. Heat will build up on successive startsand the temperature will rise even after the motor is stoppedbecause air movement is not present for heat dissipation. Thistype of operation poses unusual problems in the selection ofproper protective devices. Thermal protectors located in themotor starter will cool more rapidly than the motor windings, soprotection is compromised. Internal temperature sensors,known as thermal overload detectors, can be embedded in themotor windings to provide the best form of protection formotors subjected to heavy cycling.

Generally, standard integral motors are designed for continuousoperation. Cyclic service of any fan/motor combinationdemands special consideration. Such situations should beexplained and carefully reviewed with the fan and motormanufacturers.

Excessive Loading. When too much is demanded of a motor, itwill attempt to compensate by drawing more current. Heatbuild-up is proportional to the square of the increase in current.Proper overload protection will guard against excessive heatbuild-up; however, it is unwise to use overcurrent protectorswith automatic resets because the motor can cycle untilenough heat builds up to damage the windings.

The potential problems of excessive loading are often dealtwith by using backwardly inclined fan designs. As explainedpreviously, it is possible to select a motor for a backwardlyinclined fan that will not overload at a fixed speed, regardless ofany changes in system resistance.

CONCLUSION

The New York Blower Company frequently supplies the entirefan, drive, and motor package. However, because motor selec-tion is dependent upon the actual location, environment, andintended service, and since only the system designer or enduser can be fully aware of these variables, nyb cannot beexpected to select or recommend motor specifications.

The information contained in this Letter provides the systemdesigner or user with fundamental information to aid in theselection and application of motors. Further information canbe obtained by contacting motor manufacturers directly.

Form 507 DJK


E L E C T R I C M O T O R C O D E S A N D S T A N D A R D SINTRODUCTION

Electric motors are often required to meet various industrystandards and national codes in addition to specific applicationrequirements. The more common of these standards and codesare explained in this Engineering Letter. Also included are somegeneral motor dimensions and weights for reference purposes.

In the early days of electric motors, motors were built to thespecifications and standards of individual manufacturers. Eachbrand usually had its own unique nomenclature, dimensions,ratings, etc., thus interchangeability was seldom possible.

Currently, a number of independent groups and several specialinterest organizations provide uniform specifications to whichmotor manufacturers can comply on a selective or voluntarybasis. Some of the more common of these are listed below.

MOTOR STANDARDS ORGANIZATIONS

AIM - Automotive Industrial Motors include specific brandsmanufactured on a selective basis to meet the specificationsestablished by the automotive industry. Examples includeGMC - 7EQ, Ford EM-1, and Chrysler NPEM-100.

ANSI - American National Standards Institute representsmanufacturers, distributors, and consumers. A wide variety ofsubjects are covered, such as dimensions, material specifications,test methods, and performance. Standards frequently referencethose adopted by NEMA and IEEE.

CSA - Canadian Standards Association provides materialstandardization services for Canada. It develops or adoptsstandards for safety, quality, and performance.

IEC - International Electrotechnical Commission definesmetric equivalancies to some NEMA standards, such asenclosures, frame sizes, conduit box locations, and mountingarrangements.

IEEE - Institute of Electrical and Electronics Engineers coverssuch fundamentals as basic standards for temperature rise,classification of insulating materials, and the appropriate testcodes and rating methods.

ISO - International Standards Organization establishes uniformterminology, units, and equivalancies in international metricterms.

NEC - National Electric Code is an ANSI standard sponsoredby the National Fire Protection Association for the purpose ofsafeguarding persons and property from electrical hazards. Thecode covers wiring methods and materials, protection ofbranch circuits, motors and controls, grounding, hazardouslocations, and recommendations. See Figure 1.

NEMA - National Electrical Manufacturers Association is atrade association organized and supported by manufacturers ofelectrical equipment and supplies. Voluntary standards defineproducts, processes, and procedures with reference tonomenclature, construction, dimensions, tolerances, operatingcharacteristics, performance, testing, and rating. The standardscover such matters as motor-frame sizes and designations,circuit connections, lead markings, torque classifications, and abasis for ratings. Some of the more important itemsstandardized by NEMA are:

Speeds - see Figure 3.Horsepower Ratings - see Figure 3.Frame Sizes and Dimensions - see Figure 4.Conduit Box Locations - see Figure 5.Standard Voltages and Frequencies*Service Factors*Enclosures*Starting CurrentTorques

* Note: refer to Engineering Letter 22 - Integral Motors forCentrifugal Fans.

UL - Underwriters Laboratories, Inc. is an independent testingorganization specializing in testing products, systems, andmaterials with particular reference to life, fire, and casualtyhazards. Standards have been developed for motors andcontrols in cooperation with the manufacturers. The variety ofstandards for motors compliance include:

1. Motors for use where explosive vapors, combustibledusts, or easily ignitible flyings exist…as adopted byNEC.

2. Motor-operated appliances.3. Motor overload protection devices.

NEC EXPLOSION-PROOF MOTOR DESIGNATIONS

The National Electrical Code categorizes common hazardous atmospheres and locations. Classification of hazards might bedefined by the plant safety engineer or by the insurance company. Since the type and degree of hazard varies widely according tothe materials encountered and their probable presence in hazardous quantities, the following methods of identification are used:Class - materials are “classed” as flammable vapors or gases (Class I); or as combustible dusts (Class II).

Group - materials are “grouped” according to their relative degree of hazard with Groups C and D applicable to vapors or gases,and Groups E through G applicable to combustible dusts.

Division - the containment aspects are defined by “divisions” according to the likely concentration of the hazard. Division 1 isapplicable to routine or periodic exposure, while Division 2 refers to a hazard that is normally confined within a system orcontainer and which would only escape in the event of some abnormal circumstance or equipment failure. NEC requires the useof explosion-proof motors for all Division 1 locations.

Class I Group C - Atmospheres containing ethyl vapors, ethylene, or cyclopropane.Class I Group D - Atmospheres containing gasoline, hexane, naptha, benzine, butane, alcohol, acetone, benzol,

lacquer-solvent vapors, or natural gas.Class II Group E - Atmospheres containing metal dust.Class II Group F - Atmospheres containing carbon black, coal, or coke dust.Class II Group G - Atmospheres containing flour, starch, or grain dust.

The specific motor Class and Group must be determined for the particular hazard involved. Motors designed and rated for onetype of hazard or location are not necessarily suitable for use in another situation . . . consult the motor manufacturer for specificapplication information.The explosion-proof motor ratings normally stocked by motor manufacturers or distributors are Class I, Group D and Class II,Groups F and G, Division 1. Other ratings, such as Class I, Group C or Class II, Group E, Division 1 are non-standard but areavailable on special order.

Figure 1

COMMON MOTOR WEIGHTS AND SHEAVE LIMITS

Motor Weights (lbs.)1 Sheave Limitations (Inches) 2

ODP TE Maximum WidthFrameMin./Max. Min./Max.

Min.Pitch Dia. Narrow Conven.

143T 26/41 28/65 2.2 21/4 41/4145T 33/55 35/70 2.4 21/4 41/4182T 50/105 55/111 2.6 23/4 51/4184T 60/120 70/125 3.0 23/4 51/4213T 90/137 99/197 3.0 33/8 61/2215T 100/160 121/224 3.8 33/8 61/2254T 145/275 23 1/384 4.4 4 73/4256T 160/3 10 265/415 4.6 4 73/4284T 228/374 359/495 5.0 45/8 9284TS 225/372 356/425 * * *286T 275/409 390/499 5.4 45/8 9286TS 250/380 380/475 * * *324T 366/495 490/700 6.0 51/4 101/4324TS 333/478 458/671 * * *326T 415/600 526/766 6.8 51/4 101/4326TS 406/565 490/73 8 * * *364T 580/792 748/948 7.4 57/8 111/2364TS 519/777 730/916 * * *365T 620/835 804/1040 9.0 57/8 111/2365TS 600/821 777/1004 * * *404T 845/1110 1100/1220 9.0 71/4 141/4404TS 750/1108 1000/1211 * * *405T 816/1163 1049/1368 11.5 71/4 141/4405TS 800/1150 907/1312 * * *444T 1122/1528 1400/1820 11.0 81/2 163/4444TS 1100/1515 1365/1799 * * *445T 1250/1750 1500/2458 13.2 81/2 163/4445TS 1200/1600 1481/2300 * * *

* Not recommended for belt driveFigure 2

NEMA STANDARD FRAME SIZES

Synchronous Speed (RPM) 3

3600 1800 1200Rating(HP)

ODP TEFC ODP TEFC ODP TEFC3/4 -- -- -- -- 143T 143T1 -- -- 143T 143T 145T 145T

11/2 143T 143T 145T 145T 182T 182T2 145T 145T 145T 145T 184T 184T3 145T 182T 182T 182T 213T 213T5 182T 184T 184T 184T 215T 215T

71/2 184T 213T 213T 213T 254T 254T10 213T 215T 215T 215T 256T 256T15 215T 254T 254T 254T 284T 284T20 254T 256T 256T 256T 286T 286T25 256T 284TS 284T 284T 324T 324T30 284TS 286TS 286T 286T 326T 326T40 286TS 324TS 324T 324T 364T 364T50 324TS 326TS 326T 326T 365T 365T60 326TS 364TS 364T 364T 404T 404T70 364TS 365TS 365T 365T 405T 405T100 365TS 405TS 404T 405T 444T 444T125 404TS 444TS 405T 444T 445T 445T150 405TS 445TS 444T 445T 445T 445T200 444TS 447TS 445T 445T 445T 447T

Figure 31. Motor weights are not standardized and vary with manufacturer,

enclosure, frame, etc. The minimum and maximum weights shownare representative of the range available from several majormanufacturers as of November, 1995. Where exact weights arerequired, consult the specific manufacturer.

2. The sheave limitations shown represent the most restrictedparameters from several major manufacturers. It may be possibleto exceed these parameters for a given situation by consulting thespecific manufacturer.

3. Motor frame sizes may vary with special features or characteristics.Refer to Engineering Letter 22 - Integral Motors for CentrifugalFans for nominal speeds.

NEMA STANDARD DIMENSIONS (Inches)

C-ODP1

1C-TE1

Frame BA D* E F U V †Min. Max. Min. Max.

143T 2.25 3.50 2.75 2.00 .875 2.00 10.99 12.82 10.45 13.35145T 2.25 3.50 2.75 2.50 .875 2.00 11.12 12.99 11.45 14.35182T 2.75 4.50 3.75 2.25 1.125 2.50 12.36 14.72 13.55 17.15184T 2.75 4.50 3.75 2.75 1.125 2.50 13.56 16.50 13.55 17.15213T 3.50 5.25 4.25 2.75 1.375 3.13 15.50 18.19 17.18 20.28215T 3.50 5.25 4.25 3.50 1.375 3.13 16.31 18.19 17.18 20.28254T 4.25 6.25 5.00 4.12 1.625 3.75 20.00 22.32 21.50 25.60256T 4.25 6.25 5.00 5.00 1.625 3.75 21.69 23.19 23.20 25.60284T 4.75 7.00 5.50 4.75 1.875 4.38 23.19 25.94 25.33 28.93284TS 4.75 7.00 5.50 4.75 1.625 3.00 21.82 22.44 23.95 27.55286T 4.75 7.00 5.50 5.50 1.875 4.38 23.81 25.06 26.83 28.93286TS 4.75 7.00 5.50 5.50 1.625 3.00 22.44 23.69 25.45 27.55324T 5.25 8.00 6.25 5.25 2.125 5.00 21.38 27.25 28.15 32.25324TS 5.25 8.00 6.25 5.25 1.875 3.50 21.38 25.75 26.65 30.75326T 5.25 8.00 6.25 6.00 2.125 5.00 26.69 28.50 29.65 32.25326TS 5.25 8.00 6.25 6.00 1.875 3.50 25.19 27.00 28.15 30.75364T 5.88 9.00 7.00 5.62 2.375 5.63 28.62 29.69 31.28 34.28364TS 5.88 9.00 7.00 5.62 1.875 3.50 26.50 29.70 29.15 32.15365T 5.88 9.00 7.00 6.12 2.375 5.63 26.57 29.69 31.28 34.28365TS 5.88 9.00 7.00 6.12 1.875 3.50 27.50 29.81 29.15 32.15404T 6.63 10.00 8.00 6.12 2.875 7.00 32.38 34.19 33.88 39.91404TS 6.63 10.00 8.00 6.12 2.125 4.00 29.38 31.19 30.89 36.91405T 6.63 10.00 8.00 6.87 2.875 7.00 33.88 34.19 36.85 41.95405TS 6.63 10.00 8.00 6.87 2.125 4.00 30.88 31.19 33.85 38.95444T 7.50 11.00 9.00 7.25 3.375 8.25 37.56 39.94 39.56 46.68444TS 7.50 11.00 9.00 7.25 2.375 4.50 33.81 36.18 35.31 42.93445T 7.50 11.00 9.00 8.25 3.375 8.25 38.62 39.94 39.56 48.68445TS 7.50 11.00 9.00 8.25 2.375 4.50 35.87 36.18 35.31 44.93

*Tolerance: 8" or less, + .000, - .03 1, Over 8", +.000, - .062. Tolerance: 11/2" dia. or less +.0000, - .0005; Over 11/2" dia. + .000, - .001.† V is usable shaft length.

Figure 4

1. The overall motor length is uniformly designated asNEMA “C,” but the dimension itself varies betweenmanufacturers. The “C” dimensions shown arerepresentative of the range available from severalmanufacturers as of November, 1995. Where exactdimensions are required, consult the specific motormanufacturer.

2. The distance from the center of the motor shaft to theoutside edge of the conduit box is known as NEMA“AB.” Since this dimension varies with manufacturer,enclosure, frame, etc., consult the specific motormanufacturer.

Form 507 DJK

MOTOR ROTATION DESIGNATIONS

Figure 6

Conduit-box locations . . . the standard location forfloor mounted motors is designated as F-1, where theconduit box is on the right when viewing the endopposite the shaft. Although other arrangements areavailable as indicated, they are non-standard andrequire special production and extended deliveryschedules in most cases. Thus, the F-1 is used forthe majority of fan applications regardless of fanarrangement. See Figure 6.

Assembly F-1, W-2, W-3, W-6, W-8 and C-2 =Standard Lead Location.

Assembly F-2, W-1, W-4, W-5, W-7, and C-1 = LeadLocation Opposite Standard.

Motor rotation . . . the direction of the motor rotationcan be significant, particularly in large fan-cooledmotors. The increasing demand for energy-efficientand quiet-operating motors has forced motormanufacturers to use uni-directional cooling fans inmany cases. Thus, the motor manufacturer will needto know the required rotation in many cases.

Most motor manufacturers specify CW or CCWwhen viewing the end opposite the shaft. Therefore,the motor rotation will be the same as the fan rotationin Arrangements 4, 7, 8, and 1 or 3 with motorpositions X and Y. The motor rotation will beopposite the fan’s in Arrangements 9, 10, and 1 or 3with motor positions W and Z. This may differ withsome manufacturers, since there is no formal standard.


F U N D A M E N T A L S O F S T E A MINTRODUCTION

A good knowledge of steam heating, for process work and airhandling/ventilation systems, is important to design engineers,building owners, and maintenance personnel who mayencounter steam systems. This Engineering Letter was writtenas a basic reference tool, primarily for those who have not beenregularly involved in designing and operating steam-heatingsystems.

LATENT HEAT

One of the factors important in holding the earth’s surfacetemperature within its rather narrow bounds is the fact thatwhile it takes about 1 Btu to change the temperature of apound of liquid water by 1 °F., it takes 144 Btu to freeze onepound of water (latent heat of fusion) and about 1000 Btu toconvert one pound of water to steam (latent heat of evapo-ration). The relatively large amount of heat change required toconvert water into either ice or steam acts to keep the earth’stemperature moderate.

Heating water from 32°F. to its boiling point, 2 12°F. at sealevel, requires about 180 Btu per pound (one Btu per degree).This is referred to as sensible heat. Converting the water at212°F. to steam at the same temperature requires about 1000Btu per pound. This is the heat applied in a steam boiler.

Conversely, when the latent heat is extracted from the steam,perhaps by condensing it in a section of STEELfin coil, the1000 Btu per pound is given up by the steam without anychange in temperature.

Figure 1 shows how the temperature of one pound of waterwould vary if subjected to a constant rate of Btu input. Noticethat it would stay at 32°F. and 2 12°F. (at sea level) until, ineach case the latent heat conversions had taken place for theentire pound of water.

SATURATION

If a container of water is heated sufficiently at a constantpressure, the water temperature will rise until the boiling pointis reached. While boiling, the temperature will remain constantuntil all the water has been converted to steam. Then thetemperature will rise again as the steam is further heated, asshown in Figure 1. Steam at the temperature at which it co-existswith water is called saturated steam. The temperature is calledthe saturation temperature. The saturation temperature varieswith the pressure. An increase in pressure increases thetemperature at which the latent heat transfer takes place. Thepressure at which the latent heat transfer takes place (at agiven temperature) is called the saturation pressure.

BTU to raise the temperature of one pound of water.Figure 1

For example, at sea level normal atmospheric pressure is 14.7psia (absolute pressure). The saturation temperature is 212°F. At2 12°F. the saturation pressure is also 14.7 psia (which is also 0psig see Pressures, below). Almost all useful steam-heattransfer work takes place at the latent heat-saturationtemperature and pressure point. Saturation pressures,temperatures, and latent heat values are shown in Figure 2.

GaugePressure

Temp.°F.

LatentHeat

GaugePressure

Temp.°F.

LatentHeat

2 219 966 80 324 8915 227 960 90 331 886

10 239 953 100 338 88015 250 945 110 344 87520 259 939 120 350 87125 267 933 130 356 86630 274 929 140 361 86140 286 920 150 366 85750 298 912 175 377 84760 307 905 200 388 83770 316 898

Steam gauge pressures, saturation temperatures, and latent heatvalues at sea level, standard barometric pressure of

29.92" Hg = 14.7 psia.Figure 2

PRESSURES

In the English system of measure, steam pressures aremeasured in pounds per square inch. In international units,steam pressures are measured in pascals or kilopascals where 1psi is equal to 6894.7 pascals. For the sake of simplicity,English units are used in this Engineering Letter.

There are, necessarily, two reference levels for measuringpressure. One is the pressure above atmospheric. This is theboiler pressure, commonly called gauge pressure andabbreviated either psi or psig. Because of the variable nature ofatmospheric pressure, steam pressures are more accuratelydescribed in terms of their absolute pressure. This is the totalamount of pressure above a perfect vacuum. At sea level,atmospheric pressure is 14.7 psia. Hence gauge pressure (psig)+ 14.7 = absolute pressure (psia).

SUPERHEAT

Steam is a gas. As in the case of any gas, it can be heatedabove the boiling point. Once it is past the saturationtemperature it requires only about .5 Btu per pound to increase itstemperature 1°F.

The increase in temperature above the saturation temperatureis called superheat. Steam that has a small amount of super-heat is called dry steam. If heated more than a few degreesabove the saturation temperature it is referred to assuperheated steam. Obviously, neither dry nor superheatedsteam can co-exist with liquid water. Since steam is a gas ittends to expand with a direct relation to temperature. Theincreased volume and small amount of extra heat value makessuperheat a relatively worthless factor in steam heating. Itsonly real value is to ensure that there will be dry steam at thepoint where the steam is to be used. In other words, a fewdegrees of superheat at the boiler will minimize condensationin the supply lines to the steam coils.

CONDENSATION

When steam gives up its latent heat and changes from saturatedsteam to water at the same temperature, it condenses. Thewater is referred to as condensate.

HEAT TRANSFER

Figure 3 shows the cross section of a typical steam coil. Theheat produced by the condensation of the steam travelsthrough the boundary layer of steam, through thecondensation that forms on the inside of the tube, through thetube itself, out into the fins, and through the boundary layer ofair on the fins’ surfaces and into the passing stream of air.

All steam coils are 100% efficient in the sense that the heatreleased by condensing steam within the coil has nowhere togo but into the air surrounding the coil. Tube-and-fin material,fin spacing, air velocity, and some other factors affect the rateat which the heat transfer (and therefore the condensation)takes place but they cannot alter the fact the steam’s latentheat has only one place to go: into the airstream.

Form 607 GAW

Steam coil cross-section showing the temperature gradientwith 5 psig steam (227°F. saturation temperature)

heating air to 90°F.Figure 3


INDUSTRIAL STEAM HEATING SYSTEMSINTRODUCTION

Reduced to its barest elements, a steam heating system consistsof a boiler to convert water to steam, piping to conduct thesteam to where it is to be used, a coil or other surface forcondensing the steam and transferring the latent heat from thesteam to the air, a trap to prevent the steam from passingthrough the coil before it is condensed, and return piping tobring the condensate back to the boiler. The purpose of thisEngineering Letter is to provide a basic overview of the majorelements found in typical industrial steam heating systems.

SYSTEM COMPONENTS

Boilers

While the boiler and its attachments are major factors in thesteam heating system, it is not the intent of this Letter to domore than point out that boilers are generally divided into“Low Pressure” and “High Pressure” designs. Low pressureboilers, running up to 15 psig, are generally used for spaceheating with unit heaters, make-up air units, heating andventilating units, etc. There is no benefit in raising the steampressure or temperature much beyond the minimum needed toboil water and to provide the pressure necessary to drive thesteam through the piping system. Higher pressures not onlyrequire more expensive piping and fittings but the addeddanger involved in higher pressures and temperatures has givenrise to municipal and insurance codes requiring additional safetyfeatures, licensed operators, etc.

High pressure boilers generate more than 15 psig. High-pressure systems are used either to provide adequate pressurefor long runs of steam piping or to develop higher temperaturesfor process systems. The air passing across a steam coil cannotbe heated any higher than the steam temperature. At 5 psig thesteam temperature is 227°F. At 200 psig it is 388°F. There islittle difference between the amount of total heat at 5 psig and at200 psig but the fact that the heat is released at a highertemperature gives the capability of producing substantially higherfinal air temperatures.

Piping

Piping is addressed on page 3.

Steam Coils

The steam coil is the part of the system designed to condensethe steam and transfer the latent heat to the airstream. If all coilsare 100% efficient, then what differentiates a good steam coilfrom a poor one? Here are some important factors:

1. The metal or metals of which a steam coil is manufacturedare relatively unimportant insofar as heating capacities areconcerned but may be extremely important in determining thelife of the coil. Coils have been successfully made from almostevery conceivable metal. Copper tubes have long been a favoritebecause of copper’s supposed corrosion resistance and ease ofsoldering, brazing, and forming. However, other tubes,particularly steel, are quite adaptable to the manufacture ofsteam coils. Conventional copper or steel tube coils are usuallyadequate for commercial heating installations.

2. Industrial heating and process applications demand themost rugged possible coil construction. The most practical coilis one using heavy-gauge, welded-steel tubes with an oval-shaped cross-section. The resultant strength is several timesthat of light-gauge copper or steel tubing. A round tube willsp lit when filled with water and frozen, as so often happenswhen the condensate return system fails for one reason oranother. An oval tube deforms slightly, increasing its cross-sectional area, but rupture normally will not occur if the ovaltube is made of heavy-gauge, high-strength steel.

3. Condensate is water and it runs downhill. The condensatedrains from the coil’s tubes by gravity. Good coil installationproduces an almost uniform pressure through the coil. Thesteam pressure cannot and does not force the condensatethrough the tubes. For high heating capacities, the tubes shouldbe vertical. This allows quick drainage and clearing of thetubes. In addition to reducing the possibility of freezing, thewashing action brought about by the quick drainage alsoreduces the boundary layer of water in the tubes and improvesheat transfer.

An advantage of vertical-tube coils, often overlooked, is theirlack of susceptibility to water hammer. Water is virtuallyincompressible. When driven through a pipe or co il tube at thevelocity of steam, it “hammers” the turns in the pipe or the endof the coil tube. Vertical drainage eliminates water hammer invertical-tube coils. Horizontal-tube coils are destroyed byrepeated water hammer. Typically, water hammer results in afairly uniform bulge, or rounding, at the end of the steam coiltube. When the bulge finally ruptures it is frequently mistakenfor failure due to freezing. The visual distinction between theresults of the two kinds of failures is that water hammer gives asymmetrical bulge at the end of the tube, where freezing gives anon-symmetrical distortion.

4. Lack of maintenance, particularly in industrial plants, cancause deterioration of the coil and of its capacity. Coils withthin copper tubes and thin aluminum or copper fins arephysically weak. Normal industrial cleaning methods can betoo rough. Cleaning aluminum or copper fins with an air hose isalmost certain to deform the fins and result in a loss of heatingcapacity. Welded-steel tubes with steel fins bonded to them andreinforced with hot-dipped galvanizing offer the physicalstrength to withstand scrubbing or high-pressure air-hosecleaning.

Although not precisely related to the subject of this Engineer-ing Letter, it seems worth recording the “Steam Formula”, theequation used to predict coil performance at one steam pressureand entering air temperature from the performance of the samecoil at the same standard air velocity but at a different steampressure and/or entering air temperature:

TR1 ST1 - EAT1

TR2= ST2 - EAT2

, where

TR is air temperature rise through the coil,ST is steam saturation temperature,EAT is entering air temperature.

Traps

All steam traps serve the same basic purposes:

1. The trap prevents the higher steam supply pressure frompassing directly to the return line. If the supply pressure hadready access to the return piping, the whole system would be atthe same pressure and there would be reduced steam flow.

2. The steam must not be allowed to pass through the trapuntil it has condensed in the coil. The whole purpose of thesteam heating system is to condense the steam in the coil, andnowhere else.

3. When a steam heating system is started up, the system isfilled with air. The water used to produce steam containsdissolved air, which is released when the water is heated. Itmay also contain nascent oxygen and noncondensable gaseswhich can form CO2 and which, if not released immediatelyfrom the coil, will inhibit heat transfer and may attack the tubewalls. The air and gases must be allowed to pass through thecoil and out of the trap. On high-pressure steam systems, thetrap may not have enough air-venting capacity. Refer to note 3on condensate piping later on in this Letter.

All traps are rated on the basis of constant steam and condensateflow at a differential in pressure across the trap. In practice,constant flow rates are seldom encountered. Temperaturecontrol variations are the principal cause of uneven flow rates.All steam traps should be sized to handle three times theanticipated maximum condensate rate to ensure condensateremoval under surge-load conditions and cold startups.

Condensate will not flow from one side of a trap’s orifice tothe other without a pressure differential.

For systems with non-modulating types of steam control, thetrap must at least be below the coil to ensure that the waterlevel in the trap is below the coil.

For systems with modulating types of steam control, the trapshould be at least twelve inches below the coil to ensure thetrap of a water head when the modulating valve has throttleddown to 0 psig at the coil. Therefore, for modulating systems,the trap should be sized to handle the maximum condensateload at the pressure available in the water leg only. For atwelve-inch leg, this would be .43 psi.

The two types of traps of most interest for industrial heatingand process work are described below:

The Float and Thermostatic Trap shown in Figure 1 is theclosest thing to a general-purpose trap for industrial heatingand process work. F and T traps function well over broadranges of pressure and steam volume. They are especiallysuitable for low to medium pressures up to about 20 psig.However, they should not be used on systems involving steamthat is superheated more than a few degrees. In operation, air isvented through the thermostatic element on systems with under20 psi steam pressure at the coil. Condensate raises the float,opening the lower port.

The Inverted Bucket Trap of Figure 2 should, generally, takethe place of the F and T trap for both high pressure steam andfor superheated steam systems. There are other types of traps,but they should not be used as condensate traps on heating andventilating systems.

Float and Thermostatic Trap (Courtesy of Sarco Co.)Figure 1

Inverted Bucket Trap (Courtesy of Sarco Co.)Figure 2

PIPING

The key to successful steam piping requires that these twoprinciples be kept in mind:

A. Steam is a gas and can flow in any direction, butcondensate, a liquid, flows downhill.

C. Both steam and condensate cause friction when theyflow. As with air flowing in ducts, consideration must begiven to velocity, pipe size, and pressure drop.

Bringing the steam to the coil is not nearly so difficult nortroublesome as getting the condensate from the coil back tothe boiler. Because “steam” is the working element in thesystem and condensate is, after all, only ordinary water, wetend to concentrate our attention on the steam piping andignore the condensate piping. We should do just the opposite.Although the following discussion treats steam piping first, it isthe return piping that demands most careful attention.

Referring to Figures 3 (Low Pressure) and 4 (High Pressure)the elements of a good steam-piping system are:

A. Steam mains must be sized based on the steam pressure,how much of the pressure may be used to overcome frictiondrop, and the length of the longest run. (System designersaccustomed to air-duct design will recognize the basic similarity.)A nomograph for sizing steam pipes is contained on page 8.

Pipe expands when heated. The increase is .00008 in./ft.°F. A100-foot long main for 50 psig steam would expand.00008(100) (298-70) = 1.82". Piping must be installed so theexpansion may take place without placing stress on the pipe orthe equipment to which it is connected. Some of the methodsemployed to accommodate expansion are metal bellowsexpansion joints, expansion loops (Figure 5), swing connections(Figure 6), and pipe-support brackets employing rollers.

Some steam condenses in the steam mains. The amount maybe minimized by insulating the pipes and by using superheat,but all steam supply piping should provide for condensatedrainage.

Vertical-steam pipes cause no particular problem if the steamif flowing down, but long up-flowing steam lines can betroublesome. Water hammer can be avoided by installing ashort horizontal swing connection and drip leg every 20 to 40feet.

The condensate that forms in the steam pipe is passed through atrap to the return (condensate) line. (Sometimes the connectionand trap are called the drip leg and drip trap.) See Figure 7.

Figure 7 -Drip Leg and Trap Systems Use Swing Connections

The purpose of good drainage and drip lines is to avoid waterhammer. Steam traveling at high velocity has the capability ofscooping up condensate and driving it, in slugs, against a pipeturn, valve, coil, etc. The hammering effect can be violentenough to burst pipes. The only prevention of water hammer isto keep the steam lines “dry”, i.e., clear them of condensate atfrequent intervals.

B. Steam supply to the coils should consist of thesecomponents:

1. A drip line and trap should parallel the coils unless thecoils are located quite close to a drip line on the main. Thesteam supply should rise above the drip line, as it approachesthe coil, for best drainage.

2. Swing connections, see Figures 3, 4, and 6, from the mainto the branch and from the branch to the coils.

3. A strainer to keep foreign matter out of the valves, coils,and traps. See Figure 8.

4. A shutoff valve for possible maintenance use.

5. A pressure-control valve.

6. A union. By putting unions and shutoff valves on bothsides of coils and traps, an individual coil or trap may beremoved without shutting down the entire system.

Steam or Condensate Strainers (Courtesy of Sarco Co.)Figure 8

C. Condensate (return) piping should include:

1. A stub pipe or “dirt pocket”, at least 8" long, directly belowthe coil. This is simply a settling spot for dirt and scale, andshould be periodically emptied.

2. The strainer, Figure 8, with the dirt pocket, keepsextraneous matter from the mechanism of the trap. Boilers,pipes, and coils are apt to contain small particles of scale, weld-spatter or thread-turnings. The strainer in the condensate line isintended primarily to pick up dirt, pipe dope, etc., that find theirway into the system during installation. The element should beremoved from the condensate strainer assembly after thesystem is fully in operation. It should not be replaced. Thestrainer on the supply side of the coil is adequate for the entiresystem. Since “high pressure” steam implies high velocity andrapid scouring of dirt from pipes, especially when the system isnew, it may be best to use strainers that are available withaccessory blow-down valves for frequent and quick cleaning.

3. On high-pressure systems, over 15 psig, it is desirable toprovide more air-venting capacity than is incorporated in thetrap. This may be done in one of two ways:

a. With an air eliminator, which is a thermostatic vent. Thistype should be used only if it can be guaranteed to operateat the elevated temperature corresponding to the steamtemperature.

b. By means of a petcock left continuously open. The loststeam is far less costly than the damage done to coils byinadequate venting.

Improper venting of high pressure systems is a major cause ofcoil problems. The high-temperature gases entrained in thesteam, if not eliminated, may combine with the condensate toform acids.

4. The traps, described in a previous section, must be installedbelow the coils. Water flows downhill. Overhead return lines(Figure 9) are perhaps the biggest single cause of freezing, waterhammer, coil corrosion, and trap failure.

While it is theoretically possible for the steam pressure in thecoil to push (lift) water into an overhead return line there arejust too many reasons why the pressure may not be availablewhen most needed. Consider, for example, a 25 psi boilersystem. Assuming a 5 psi drop through the lines, the remaining20 psi should be able to raise water 46 feet. (One sea levelatmosphere is equal to 14.7 psi. This is, in turn, equivalent to a“head” of 34 feet of water. Stated differently, standardbarometric pressure at sea level is 34 feet of water. Since 14.7psi will “lift” 34 feet of water then 1 psi will lift 2.3 feet andthe 20 psi in the example will lift 46 feet.) On this basis a 15-foot “lift” into an overhead line would seem reasonable.

But, on the first cold Monday morning of winter, when theplant heating and process systems were shut down over theweekend, every terminal on the steam system will be atmaximum demand. The boiler may develop only 20 psi.

The steam will travel at higher-than-ordinary speeds, and thepressure drop may become 10 psi. The steam coils, normallythought of as having negligible pressure drop, will be temporarilystarved for steam. The steam will condense so rapidly in thecold coils that the 10 psi at the coil inlets might drop to 5 psi inthe coils. Five psi will lift water 111/2 feet, but cannot buck the15-foot rise. The trap and coil will become waterlogged. Waterhammer may be severe in horizontal tube coils. If the coil ishandling air below 32°F. the coil will freeze.

Or, consider shutting down the same system at the end of theheating season. As the steam pressure drops, a point is reachedwhere the coil is again waterlogged. A stagnant water level in acoil is an invitation to corrosion.

5. Not shown in Figures 3 or 4, but often advantageous, is an“aquastat” strapped to the return line just beyond the trap. It isset so cold temperature, indicating no condensate flow, shutsoff the fan and thereby prevents freezing air from passing over awater-filled coil. It does not prevent the occurrence of waterhammer in horizontal tube coils.

6. Where overhead returns are unavoidable, the only goodsolution is to drop first into a vented reservoir (sometimescalled receiver) and use a motor-driven condensate pump to liftthe water into the overhead line. This relieves the trap and coilof the dangers of waterlogging.

Despite all of the reasons for not using overhead returns withoutcondensate pumps, such installations are found. In fact, theyare so common that they will be discussed here. This is bestdone by differentiating between those systems that usemodulating steam control and those that use non-modulatingcontrol.

a. Non-modulating control systems may be calculated asthe steam pressure is always great enough to overcomethe rise in the return line. It can be argued that there aresteam systems that do not involve handling low-temperature air and therefore present no problems offreezing. Such a system might be a process systemcompletely enclosed within a manufacturing plant.However, even in such a system there comes a timewhen the steam valve is shut off. The condensate that is,at the moment, on the supply side of the trap cannot bedischarged from the system (unless fitted with anothersmall trap line that can drain this trapped water into asewer) and if the water level happens to be such that itsettles out across a coil and is allowed to sit there for anylength of time, the coil is apt to corrode at the watersurface.

b. Modulating systems present a unique situation in thatunder most conditions the only pressure available at thetrap is the water leg between the coil and the trap. Forexample, a coil that will heat from -10°F. to 60°F. with 5psig (227°F.) steam will heat from -4°F. to 60°F. with 0psig (212°F.) steam. This not only makes controldifficult but aggravates the condensate removal problem.Therefore, a modulating system must be provided with avacuum-breaker on the return side of the coil to ensurethat the trap will at least have equal pressure on theupstream and downstream sides - plus the maximumwater head over twelve inches that space will allow. (Avacuum-breaker is just a swing check valve installed so itopens into the system.) Obviously overhead returnscannot be tolerated on this type system without the use ofa vented reservoir and condensate pump.

c. Due to the difference in volume between water andsteam, condensate pipes may be sized at 60% of thediameter of the steam pipe, for gravity-return systems.Pumped systems may be sized at 40% of the steam pipediameter.

CONTROL METHODS

Control, when referring to steam, means control of the airtemperature leaving the coil. Proponents of other heatingmethods point out that temperature control is difficult withsteam. This is a fair criticism. Compare a steam coil to a gasburner, for example. The heat released by the gas burner is amore or less direct function of the amount of gas burned.

Contrast this to a 5 psig steam system. The maximumtemperature of the coil, at 5 psig, is 227°F. By throttling thesteam pressure down to 0 psig the temperature can be reducedonly to 212°F. This difference doesn’t allow good control.Attempting to go to a lower temperature necessitates operatingat a less than atmospheric pressure and introducing more airinto the coil through the vacuum-breaker. This raises the verysort of condensate drainage problems that were discussed inthe previous section.

However, there are methods of obtaining satisfactory control.

A. On-Off. Two-position control is relatively trouble-free butgives the least desirable type of temperature control. In UnitHeaters it is accomplished by leaving the steam “on” all thetime and turning the fan on or off as required by a thermostat.In Make-Up Air and most process and ventilation systems,where constant airflow is desired, the steam is turned full-onor full-off. Before dismissing such systems as too primitive,recognize that most residential heating is done by basically on-off systems. On-off steam systems have one great advantage -full steam pressure is available at all times to operate traps and(despite warnings) overhead return lines, and to minimize thedanger of freezing.

B. Face and Bypass. By allowing some air to bypass thecoils, and thereby remain unheated, and by blending the “face”and “bypass” airstreams it is possible to obtain good temperaturecontrol and still maintain full steam pressure on the coils. Thisis the system best-suited for steam Make-Up Air (See Figure10). Face and bypass systems may be built-up (plenum) orpackaged. Both may have the disadvantages listed below but,generally speaking, built-up systems can be designed to avoidthem.

1. The presence of steam in the coils generally precludes thepossibility of handling 100% bypass air without a temperaturerise of a few degrees.

2. Most packaged units are designed with less bypass areathan is desirable for 100% bypass flow. Most manufacturers

assume that summer operation will be with the steam off andair flowing through the face. Most customers seem to preferlow unit height to full bypass capability.

3. The two different temperature airstreams force the fan(generally downstream of the coils) to operate with inletstratification. This damages fan performance.

One important factor often overlooked in the selection ordesign of face-and-bypass systems is that the damper bladesshould have their axis of rotation perpendicular to the axis of thecoil tubes. Imagine horizontal dampers and horizontal tubesand you can see that in a partly throttled condition, air wouldbe directed towards some tubes and away from others. Usingvertical tubes and horizontal dampers gives the best possiblecombination.

C. Modulating Valves. Since the heat comes from the steam,it seems reasonable to control the heat by throttling the incomingsteam. By now the reader has been through the previousdiscussion of the difficulties involved in operating with thissort of control that results in poor drainage. In addition to thedanger of freezing, there is the possibility that horizontal coilsand long tubes can set up water hammer that will ruin the coil.

D. Preheat-Reheat. Two coils in series can be used to givegood temperature control and a reasonable measure of freezeprotection (See Figure 11). The coils must be accurately sized.The first “preheat” coil is selected to raise the entering airtemperature to about 40°F. to 50°F. The second “reheat” coilraises the air to the desired final temperature. The preheat coil issupplied with a snap action on-off steam valve. The reheat coilhas a modulating steam valve. Under maximum conditions,with the coldest (design) entering air temperature, both coils willbe under maximum pressure. The thermostatic controls are set tothrottle the reheat coil until it is fully closed. The preheat coil issized so that it will not overheat at full pressure.

E. Combinations can be made of preheat-reheat with faceand bypass. Fresh air and recirculating dampers may be usedto exercise some control by closing down on fresh air in coldweather. Caution should be used in designing combinationsystems. Complex control systems are often maintenanceheadaches. Keep it simple.

F. High-pressure steam presents the special problems ofsuperheat and “flashing”.

The high temperature of high pressure steam can aggravate theproblems of control. One solution is to pass the steam through apressure-reducing valve before it gets to the coil or temperaturecontrol valve. Reducing the pressure reduces the temperature atwhich the latent heat will be released and makes controleasier. However, reducing the pressure does not, in itself,extract any heat from the steam - so the reduced pressuresteam is superheated. Reducing saturated 150 psi steam, at366°F., to 25 psi steam, at 266°F., gives steam with up to 52°of superheat. Since superheated steam is just another gas until ithas been cooled to saturation temperature, it is necessary toincrease the size of the coil. The added coil face may bethought of as room for the superheated steam to sit and cool tothe saturation temperature. Dry superheated steam has a lowerfilm coefficient than does the wet saturated steam. This also

adversely affects the overall coefficient of heat transfer. Agood rule of thumb is to increase the coil area by 10% for each100° of superheat.

When high pressure steam is used, without pressure reduction,the condensate temperature may be high enough to cause someof the condensate to flash back into steam as it enters the low-pressure condensate line, downstream of the trap. Not all thecondensate flashes - just a small part of it, enough to absorbthe amount of heat needed to produce a stable mixture of steamand water. The mixture is therefore at a lower temperature thanthe high-pressure condensate.

G. Vacuum-steam systems. One-pipe steam systems andsome other variations were, and sometimes still are, used forsmall space heating installations. They are seldom of muchinterest in industrial heating or process work.

CONCLUSION

A knowledge of the fundamentals of steam heating is still anecessity in some process applications and building heatingsystems. The purpose of this Engineering Letter was to providea basic overview. Engineers and designers of steam-heatingsystems are encouraged to seek out additional training andresources to build their knowledge base.

ENGINEERING LETTER E The New York Blower Company ● 7660 Quincy Street, Willowbrook, Illinois 60521 - 5530

MISCELLANEOUS ENGINEERING DATA

The purpose of this Engineering Letter is to provide reference data commonly required in routine fan system computations.

UNITS COMMONLY USED IN FAN APPLICATIONS

Pressure In. WG Pascals Psi In. HG mm WG mm HG Atm 1 248.36 .03602 .07334 25.400 1.8628 .00245 .00403 1 .00015 .00030 .10227 .00750 .00001 27.761 6894.7 1 2.0360 705.13 51.715 .06805 13.635 3386.4 .49116 1 346.33 25.400 .03342 .03937 9.7779 .00142 .00289 1 .07334 .00010 .53681 133.32 .01934 .03937 13.635 1 .00132 407.98 101325 14.696 29.921 10363 760.00 1

BASIC FAN LAWS

Variable When Speed Changes When Density Changes

Volume CFM2 = CFM1 ( RPM2 ) Does Not Change RPM1

Pressure P2 = P1 ( RPM2 )2

P2 = P1 (D2 ) RPM1 D1

Horsepower BHP2 = BHP1 ( RPM2 ) 3 BHP2 = BHP1 (D2) RPM1 D1

FAN EFFICIENCY

Mechanical Efficiency =

Air Horsepower outx 100%

Shaft Horsepower in

Mechanical Efficiency

= TP x CFM

x 100% 6356 x BHP

Static Efficiency

= SP x CFM

x 100% 6356 x BHP

Volume Flow

CFM m3/s m3/min. m3/hr. l/s l/min.

1 .000472 .02832 1.6990 .47195 28.317 2118.9 1 60.000 3600.0 1000.0 60000 35.314 .01667 1 60.000 16.667 1000 .58858 .00028 .01667 1 .27778 16.667 2.1189 .00100 .06000 3.6000 1 60.000 .03531 .00002 .00100 .06000 .01667 1

Velocity

ft./min. m/s m/min. m/hr. mph Knots

1 .00508 .30480 18.288 .01136 .00987 196.85 1 60.000 3600.0 2.2369 1.9425 3.2808 .01667 1 60.000 .03728 .03238 .05468 .00028 .01667 1 .00062 .00054 88.000 .44704 26.822 1609.4 1 .86839 101.34 .51479 30.887 1853.2 1.1516 1

Rotating Speed

RPM rps

1 60.000

.01667 1

Density

lbs./ft.3 Kg/m3

1 .06243

16.018 1

Power HP Watts

1 1.341

.7457 1

PRESSURE EQUIVALENTS

Inches Water

Inches Mercury

Ounces Per Sq. In.

Pounds Per Sq. In.

MillimetersWater

1 .0733 .5763 .0360 25.42 .1467 1.153 .0720 50.83 .2200 1.729 .1081 76.24 .2934 2.305 .1441 101.65 .3667 2.882 .1801 127.06 .4400 3.458 .2161 152.47 .5134 4.034 .2522 177.88 .5867 4.611 .2882 203.29 .6601 5.187 .3242 228.6

10 .7334 5.763 .3602 254.011 .8067 6.340 .3962 279.412 .8801 6.916 .4323 304.813 .9534 7.493 .4683 330.214 1.027 8.069 .5043 355.615 1.100 8.645 .5403 381.016 1.173 9.222 .5763 406.417 1.247 9.798 .6124 431.818 1.320 10.374 .6484 457.219 1.393 10.951 .6844 482.620 1.467 11.527 .7204 508.021 1.540 12.103 .7565 533.422 1.613 12.680 .7925 558.823 1.687 13.256 .8285 584.224 1.760 13.832 .8645 609.625 1.834 14.409 .9005 635.026 1.907 14.985 .9366 660.427 1.980 15.561 .9726 685.828 2.054 16.238 1.009 711.229 2.127 16.714 1.045 736.630 2.200 17.290 1.081 762.031 2.274 17.867 1.117 787.432 2.347 18.443 1.153 812.833 2.420 19.019 1.189 838.234 2.494 19.596 1.225 863.635 2.567 20.172 1.261 889.036 2.640 20.748 1.297 914.437 2.714 21.325 1.333 939.838 2.787 21.901 1.369 965.239 2.860 22.478 1.405 990.640 2.934 23.054 1.441 1016.041 3.007 23.630 1.477 1041.442 3.080 24.207 1.513 1066.843 3.154 24.783 1.549 1092.244 3.227 25.359 1.585 1117.645 3.300 25.936 1.621 1143.0

DENSITIES OF SATURATED AIR

Temp. (°F.)

Density (lbs./ft.3 )

Temp. (°F.)

Density (lbs./ft.3 )

-20 .09027 100 .0619-10 .08824 110 .06741

0 .08632 120 .06552 10 .08445 130 .06349 20 .08264 140 .06132 30 .08090 150 .05895 40 .07921 160 .05634 50 .07753 170 .05346 60 .07589 180 .05036 70 .07425 190 .04667 80 .07262 200 .04270 90 .07094 212 .03730

VELOCITY PRESSURES (At Standard Density .075 lbs./ft.3 )

Velocity (FPM)

VP (In. Water)

Velocity (FPM)

VP (In. Water)

500 .016 3000 .561600 .022 3200 .638 700 .031 3400 .721 800 .040 3600 .808 900 .050 3800 .900

1000 .062 4000 .998 1100 .075 4200 1.10 1200 .090 4400 1.21 1300 .105 4600 1.32 1400 .122 4800 1.44 1500 .140 5000 1.56 1600 .160 5200 1.69 1700 .180 5400 1.82 1800 .202 5600 1.96 2000 .249 5800 2.10 2200 .302 6000 2.24 2400 .359 6200 2.40 2600 .421 6400 2.55 2800 .489 6600 2.72

6800 2.88

FAN SYSTEM EFFECT FACTORS

Pressure Drop, Inches Water Gauge

Air Velocity(FPM)

Round, Mitred Elbow Square-Duct Elbow

Two-piece

Multi-piece W/Out TurningVanes

With TurningVanes

R

= 1R

= 2 R

= 1 R

= 2 R

= 1R

= 2 D D D D D D

Elbow On The Inlet 3000 4000 5000

1.8 3.2 5.0

0.7 1.3 1.8

0.6 1.0 1.5

0.7 1.3 1.8

0.5 0.8 1.3

0.3 0.6 0.8

0.1 0.3 0.4

Elbow (2) Duct Diameters From The Inlet 3000 1.2 0.4 0.3 0.4 0.3 0.2 0.1

0.2 0.3

4000 2.0 0.7 0.6 0.7 0.5 0.45000 3.0 1.0 0.8 1.1 0.7 0.5

Elbow (5) Duct Diameters From The Inlet 3000 0.6 0.2 0.2 0.2 0.1 0.1 0.0

0.1 0.2

4000 1.0 0.3 0.3 0.4 0.3 0.25000 1.5 0.5 0.5 0.5 0.4 0.3

FAN PRESSURES

TP = SP + VP

TP fan = TP outlet - TP inlet

SP fan = SP outlet - SP inlet - VP inlet

VP = Velocity Pressure

TP = Total Pressure

SP = Static Pressure

ALTITUDE AND TEMPERATURE CORRECTION FACTORS (Multiply Factor by SP at Conditions)

Air Temp. (°F.)

Altitude (feet)

0 1000 2000 3000 4000 5000 6000 7000 8000 9000 10000

0 .87 .91 .94 .98 1.01 1.05 1.09 1.13 1.17 1.22 1.26 50 .96 1.00 1.04 1.08 1.11 1.15 1.20 1.25 1.30 1.34 1.3970 1.00 1.04 1.08 1.12 1.16 1.20 1.25 1.30 1.35 1.40 1.45

100 1.07 1.10 1.14 1.19 1.23 1.28 1.33 1.38 1.43 1.48 1.54150 1.15 1.20 1.24 1.29 1.33 1.38 1.44 1.50 1.55 1.61 1.67200 1.25 1.29 1.34 1.40 1.45 1.51 1.56 1.63 1.69 1.75 1.81250 1.34 1.39 1.45 1.50 1.56 1.62 1.68 1.74 1.81 1.88 1.94300 1.43 1.49 1.55 1.61 1.67 1.74 1.79 1.86 1.93 2.00 2.07350 1.53 1.59 1.62 1.72 1.78 1.85 1.91 1.99 2.07 2.14 2.22400 1.62 1.69 1.75 1.82 1.89 1.96 2.03 2.11 2.19 2.27 2.35450 1.72 1.79 1.86 1.93 2.00 2.08 2.15 2.24 2.32 2.41 2.49500 1.81 1.88 1.96 2.03 2.11 2.19 2.26 2.35 2.44 2.53 2.62550 1.91 1.98 2.06 2.14 2.22 2.30 2.39 2.48 2.58 2.67 2.77600 2.00 2.08 2.16 2.24 2.33 2.42 2.50 2.60 2.70 2.80 2.90650 2.10 2.18 2.26 2.35 2.44 2.54 2.62 2.73 2.84 2.94 3.05700 2.19 2.27 2.36 2.46 2.55 2.65 2.74 2.85 2.94 3.01 3.18750 2.28 2.37 2.47 2.56 2.66 2.76 2.81 2.96 3.08 3.19 3.31800 2.38 2.48 2.57 2.66 2.76 2.86 2.98 3.09 3.21 3.33 3.45850 2.47 2.57 2.67 2.77 2.87 2.96 3.09 3.21 3.33 3.46 3.58900 2.56 2.66 2.76 2.87 2.97 3.07 3.20 3.33 3.46 3.58 3.71950 2.66 2.77 2.87 2.98 3.09 3.19 3.33 3.46 3.59 3.72 3.86

1000 2.76 2.87 2.98 3.09 3.20 3.31 3.45 3.59 3.73 3.86 4.00

WEIGHTS OF MATERIALS, MEAN VALUES

Material Density lbs./ft.3




Air .0749 Cinders 43 Gravel, loose, piled 120 Salt, gran, and piled 48 Aluminum 165 Clay, loose, dry 63 Grit blast dust 160 Saltpeter 80Aluminum chips 48 moist 110 Gypsum, compressed 152 Sand, dry, loose 99Antimony 414 Coal, anthracite 98 loose 70 Sand, wet 110Asbestos 153 anthracite, piled 54 Iron, gray cast 442 Sandstone 144Asbestos, loose 64 bituminous 85 Iron ore, loose 150 Sandstone, crushed 82Ashes, coal, dry 40 bituminous, piled 47 Lead 710 Sawdust 7-15Ashes, wood, dry 47 Coffee 48 Lead oxide (red) 567 Shale, riprap 105Bakelite, Laminated 86 Coke 75 Leather 56 Shavings, planer 7-15

wood filler 85 Coke, piled 28 Lime 53-64 Slag, Iron 172asbestos filler 118 Coke, dry, crushed 15 Limestone 163 Slag, granulated 60crushed 43 Concrete, cinder 97 Lucite 74 Slate 172

Baking powder 56 stone 142 Magnesia 214 Soda ash 74Bauxite, dry, crushed 43 Copper 556 Magnesium 109 Soda ash, granulated 30Borax 109 Copper ore, crushed 190 Manganese ore, crushed 259 Sodium carbonate 91Borax, dry, crushed 75 Copper oxide 190 Marble, crushed 95 Sodium nitrate 141Brass 530 Cork 15 Mica 183 Sodium sulphate 167Brass chips 163 Corn meal 40 Monel metal 556 Starch 95Brick, masonry 118 Corundum, alundum 247 Natural gas 0.04475 granulated 35Bronze 509 Cotton, baled 93 Nickel 547 Steel 487Bronze, phosphor 554 loose 30 Nylon 70 Sucrose 100Calcium, carbonate 177 Dolomite 181 Paper 58 Sugar, bulk 55Calcium chloride 134 Duralumin 175 Strawboard or Sulphur 126Calcium sulphate 185 Earth, dry and loose 76 newspaper 33-44 Sulphur, crushed 50Carbide, dry, crushed 50 Earth, moist & loose 78 Paraffin 56 Talc 170Carborundum 195 Emery 250 Peat, dry 30 Tar, bituminous 69Carborundum, loose 140 Feldspar 160 Phosphate, ground 75 Tile 113Caustic soda 88 Feldspar, crushed 88 Porcelain 150 Tin 457Celluloid 90 Ferrous, grind dust 125 Potash 60 Tobacco 16Cellulose 94 Flour, compressed barreled 47 Quartz 165 Water 62.4Cement, loose 94 loose 28 Quartz, ground 84 Zinc 443Cereals, bulk barley, corn 37 Fullers earth, dry 30 Resin 67 Zinc oxide 350

oats 26 Glass, crown 160 Rubber, India 58 rye, wheat 48 Glass, flint 215 compound 115

Chalk 142 pyrex 140 hard 75 Charcoal, hardwood 34 ground 90 hard sponge 30

softwood 23 Granite 165 tire reclaim, solid 74 broken 12 loose, piled 96 tire reclaim, shred 27

Graphite 132

Form 812 MJN

MISCELLANEOUS CONVERSION FACTORS

Pressure Area 1 Pa = 1 N/m2

1 Pa = 10 dy/cm2 1 psi. = 0.0703 kg/cm2 1 lb./ft.2 = 4.884 kg/m2

1 in.2 = 6.4516 cm2 1 ft.2 = 0.0929 m2 1 yd.2 = 0.8361 m2

1 mi.2 = 2.5899 km2 Length Volume

1 mil. = 0.0254 mm 1 in. = 2.54 cm 1 ft. = 0.3048 m

1 mi. = 1.6093 km 1 nau. mi. = 1.1516 mi.

1 in.3 = 16.3871 cm3 1 ft.3 = 0.0283 m3 1 ft.3 = 7.48 gal. 1 ft.3 = 28.316 l.

1 yd.3 = 0.7646 m3 1 oz. = 29.57 ml. 1 gal. = 3.785 l.

1 gal. U.S. = 0.833 Imp. gal.

Energy

1 Btu = 777.97 ft.-lb. 1 HP = 2545 Btu/Hr.

1 HP = 1.014 metric HP 1 HP = 0.0761 boiler HP

1 KW = 3414 Btu/Hr. 1 Ton = 12000 Btu/Hr.

Metric Prefixes

deci = x 0.1 centi = x 0.01 mili = x 0.001

micro = x 0.000001 deca = x 10.0

hecto = x 100.00 kilo = x 1000.00

Mass

1 lb. = 453.5924 g.

U. S. INCH

Fraction Decimal MM

1/16 0.06250 1.588

1/8 0.12500 3.175

3/16 0.18750 4.763

1/4 0.25000 6.3 50

5/16 0.31250 7.938

3/8 0.37500 9.525

7/16 0.4375 11.113

1/2 0.5000 12.700

9/16 0.56250 14.288

5/8 0.62500 15.875

11/16 0.68750 17.463

3/4 0.75000 19.050

13/16 0.81250 20.638

7/8 0.87500 22.225

15/16 0.93750 23.813

1 1.00000 25.400

METAL SHEET AND PLATE DATA

Mild Steel, Stainless T-1, INX

Aluminum

Gauge Thickness Weight

(lbs./ft.2 ) Gauge

Weight (lbs./ft.2 )

1" 1.0 40.8 .250 3.50 3/4" .75 30.6 .190 2.65 5/8" .625 25.5 .160 2.24 1/2" .50 20.4 .125 1.75 3/8" .375 15.3 .100 1.40 1/4" .250 10.2 .080 1.12

7 (3/16") .1875 7.5 10 .1345 5.625 12 .1046 4.375 14 .0747 3.125 16 .0598 2.50 18 .0478 2.0

SHAFTING DATA (Mild Steel, Stainless)

Diameter (in.)

Weight (lbs./ft.)

5/8 1.04 1 2.67

1 3/16 3.77 1 7/16 5.52 1 11/16 7.60 1 15/16 10.02 2 3/16 12.78 2 7/16 15.87 2 11/16 19.29 2 15/16 23.04 3 3/16 27.13 3 7/16 31.55 3 15/16 41.40 4 7/16 52.58 4 15/16 65.10 5 7/16 78.95

6 96.13

°C = (°F - 32) ÷ 1.8 °K = °C + 273.15 °F = (°C x 1.8) + 32 °R = °F + 459.67

TEMPERATURE CONVERSION

ELECTRICAL FORMULAS

ENGINEERING LETTER GThe New York Blower Company ●7660 Quincy Street, Willowbrook, Illinois 60527-5530

GLOSSARYThe following terms are common to the fields of air movement, general ventilation, industrial process, and pollution control. Thedefinitions contained in this glossary provide brief descriptions of the terms as generally used in these fields. In many cases, amore thorough discussion of these terms can be found in the text of the appropriate Engineering Letter.

ABSOLUTE TEMPERATURE degrees Rankine, whereabsolute 0°R. = -459.7°F.; density corrections for temperatureare based on the percentage rate of change in degrees Rankine:degrees Kelvin where absolute 0°K = -273.1°C.

Density (temp.) =460°F. +70°F.Density (std.) x ( 460°F. + °F. (non-standard)

standard standard))

ABSORPTION the process of one substance entering into theinner structure of another.

ACCELERATION LOSS the energy required to induce air tomove at the entry to a system.

ACFM actual cubic feet per minute; the quantity or volumeof a gas flowing at any point in a system. Fans are rated andselected on the basis of ACFM, as a fan handles the same volumeof airregardless of density.

.075ACFM = actual density x SCFM

ACTUATOR mechanical device attached to a damper to moveits blades. May be manual, electric, pneumatic, or hydraulic.

ADSORPTION adhesion of a thin film of liquid or gases to thesurface of a solid substance.

AF fan wheel design with airfoil-shaped blades.

AIR CONDITIONING treating air to meet the requirements of aconditioned space by controlling its temperature, humidity,cleanliness, and distribution.

AIR CURTAIN mechanical air-moving device designed tolimit the influx of unwanted air at a building opening.

AIR-HANDLING UNIT factory-made encased assemblyconsisting of a fan or fans and other equipment to circulate,clean, heat, cool, humidify, dehumidify, or mix air.

AIR VELOCITY rate of speed of an airstream, expressed inFPM.

ALTITUDE the height above sea level of a given location.Density corrections for altitude are made using the followingformula where Z is the feet above sea level.

Density (Alt) = Density (Std) x [1 - (6.73 x 10 - 6) Z] 5.258

AMBIENT immediate surroundings or vicinity. AMCA AirMovement and Control Association.

ANEMOMETER a device which reads air velocity such as awind vane. In fan applications, it is usually a spinning-vane-type instrument used to read low velocities at registers or grills.

ANNEAL the process of relieving stress and brittleness inmetals by heating.

ANODIZE an electrolytic action of affixing a protectivecoating or film, usually applied to aluminum.

ANSI American National Standards Institute.

API American Petroleum Institute.

APPURTENANCES accessories added to a fan for the purposesof control, isolation, safety, static pressure regain, wear, etc.

ARI Air Conditioning and Refrigeration Institute.

ASHRAE American Society of Heating, Refrigeration, andAir Conditioning Engineers.

ASME American Society of Mechanical Engineers.

ASPECT RATIO the ratio of the width to the length.

ASTM American Society for Testing and Materials

ATMOSPHERIC PRESSURE one atmosphere is approximately14.7 PSI; 408" water gauge. Airflow is the result of a differencein pressure (above or below atmospheric) between two points.

ATTENUATION absorption of sound pressure. Attenuationreduces the amplitude only of a sound wave while leaving thefrequency unchanged.

AXIAL FAN fan where the airflow through the impeller ispredominantly parallel to the axis of rotation. The impeller iscontained in a cylindrical housing.

AXIAL FLOW in-line air movement parallel to the fan ormotor shaft.

BABBITT METAL an alloy containing tin, copper, andantimony; commonly used for lining bearings.

BACKDRAFT DAMPER damper used in a system to relieveair pressure in one direction and to prevent airflow in theopposite direction.

BALANCING the process of adding (or removing) weight on arotor in order to move the center of gravity toward the axis ofrotation.

BARCOL NUMBER a standard measure of FRP surface hardness.

BAROMETRIC PRESSURE a measurement of the pressure ofthe atmosphere; standard is 29.92" Hg.

BEARING LOSSES the power losses resulting from friction inthe main bearings.

BERNOULLI’S THEOREM the principle that the total energyper unit of mass in the streamline flow of a moving fluid isconstant, being the sum of the potential energy, the kineticenergy, and the energy due to pressure. In terms of airmovement, the theorem states that the static pressure plusvelocity pressure as measured at a point upstream in the directionof airflow is equal to the static pressure plus velocity pressureas measured at a point downstream in the direction of airflowplus the friction and dynamic losses between the points.

BI fan wheel design with backwardly-inclined blades.

BILLET a section of semi-finished metal or non-ferrous alloy.

BLADE LINERS pieces of material added over the wheelblades to reduce abrasion of the blades.

BLADE-PASS FREQUENCY the tone generated by the bladespassing a fixed object.

BLAST AREA the fan outlet area less the projected area of thecut-off.

BOILER HORSEPOWER the capability to evaporate 34.5pounds of water per hour into dry steam at 2 12°F. at sea level;33,500 BTU/Hr.

BRAKE HORSEPOWER [BHP] mechanical energy consumedat a rate of 33,000 ft. lbs. per minute; a consumption rating, ascompared to the production rating of horsepower itself.

BREAKDOWN TORQUE maximum torque a motor willproduce without a sudden decrease in speed. Often referred toas pull-out torque or maximum torque.

BRINELL NUMBER a standard measure of metal surfacehardness; metals with Brinell hardness ratings of 250 or moreare generally considered abrasion-resistant.

BTU British Thermal Unit; heat required to raise thetemperature of 1 pound of water by 1 °F. The Btu/hr. required toraise the temperature of a volume of standard air a specificnumber of degrees is calculated by the formula:

Btu/hr = Temp. Rise x CFM x 1.085

CAPACITOR START MOTOR type of single-phase inductionmotor with a capacitor connected in series with the startingwinding. High-starting and breakdown torque, medium startingcurrent. Used in hard-starting applications; compressors,pumps, etc.

CAPTURE VELOCITY air velocity necessary to overcomeopposing air currents or natural flow and cause contaminatedair, fumes, or material to flow in a desired direction.

CATALYST the final ingredient that triggers the chemicalreaction known as curing, which converts liquid resin to a solid.

CELSIUS a thermometric scale in which water boils at 100°and freezes at 0°, same as centigrade:

°C = .5556 x [°F. - 32°]

CENTRIFUGAL FAN a fan design in which air is dischargedperpendicular to the wheel’s rotational axis.

CFM cubic feet per minute; the volume of flow for a givenfan or system.

COATINGS specialty coverings, typically referred to as paints,with varying degrees of resistance to atmospheric or chemicalcorrosion.

COEFFICIENT OF CONDUCTIVITY the rate of heat transferthrough a material, expressed in Btu transmitted per hourthrough one square foot of surface per degree difference intemperature across the material. Figures are usually expressedfor basic materials, such as wood or insulation; per inch ofthickness, and called by the symbol “K”.

COMPRESSIBILITY a factor used by fan manufacturers tocorrect performance ratings in higher pressure ranges toaccount for the fact that air is a compressible gas that does notfollow the perfect gas laws.

COMPRESSION a phenomenon related to positive pressure.When air is forced into a system it is compressed and becomesmore dense. Depending on the volume or weight of airrequired down stream in the positive pressure portion of thesystem, the volume of air at the inlet of a fan may have to beadjusted by the ratio of absolute pressure at the entrance of thefan versus the design requirements in the system.

CONVEYING VELOCITY the air velocity required in a ductsystem to maintain entrainment of a specific material.

CORROSION the deterioration of a material by chemical orelectrochemical reaction resulting from exposure to weathering,moisture, chemical, or other agents in the environment inwhich it is placed.

CRP Certified Ratings Program.

CSA Canadian Standards Association. Sets safety standards formotors and other electrical equipment used in Canada.

CURVE, FAN PERFORMANCE a graphic representation ofstatic or total pressure and fan BHP requirements over anairflow volume range at a stated inlet density and fan speed.

CURVE, SYSTEM a graphic representation of the pressureversus flow characteristics of a given system and density.

DAMPER an accessory to be installed at the fan inlet or outletfor air-volume modulation.

dbA sound-pressure level corrected to the “A” weighingnetwork.

DECIBEL the logarithmic ratio between some known referenceand some quantity of electric or acoustic signal power.

DENSITY the measure of unit mass equal to its weight dividedby its volume (lbs./ft.3); standard air is .075 lbs./ft.3 .

DEW POINT the temperature at which condensation begins toform as air is cooled.

DFT dry-film thickness usually expressed in thousandths ofan inch (mils).

DIFFERENTIAL PRESSURE the difference of static pressuresat the fan outlet and inlet; also see FAN CAPACITY.

DILUTION VENTILATING the mixing of contaminated airwith uncontaminated supply air for the purpose of attainingacceptable working or living conditions.

DIRECTIVITY FACTOR the number representative of theradiation characteristics of a sound source.

DRY-BULB TEMPERATURE the combined temperature of awater vapor and air mixture.

DUST air suspension of particles [aerosol] of any solidmaterial, usually with a particle size smaller than 100micrometers.

DUST COLLECTOR an air-cleaning device used to removeheavy-particulate loadings from exhaust systems prior todischarge.

DWDI double-width, double-inlet fans, Arrangement 3.

DYNAMIC BALANCE the mechanical balancing of a rotatingpart or assembly in motion.

DYNAMIC INSERTION LOSS a reduction of airborne noiselevels affected by the installation of an acoustical silencer.

DYNE a unit of force equal to that which would accelerate onegram by one centimeter per second.

EFFICIENCY, MECHANICAL TOTAL the ratio of fan outputto the power applied to the fan. Can be helpful in selecting fansize, type, or manufacturer for the same application:

TP x CFMME = 6356 x BHP

EFFICIENCY, STATIC the ratio of fan output less the kineticenergy [outlet-velocity pressure] leaving the fan to the powerapplied to the fan:

SP x CFMSE = 6356 X BHP

ELEVATION the distance of the subject site above or belowsea level.

END REFLECTION a known value of sound radiated backinto a duct or opening.

ENTHALPY the heat content per unit mass of a substance.

ENTRY LOSS the loss in pressure caused by air flowing into asystem; normally expressed in fractions of velocity pressure.

EQUIVALENT DUCT DIAMETER for rectangular duct withsides a and b is:

D = (4ab/π)0.5

EVASE a diffuser at the fan outlet which gradually increases inarea to decrease velocity and to convert kinetic energy to staticpressure [regain.]

FAHRENHEIT a thermometric scale in which water boils at212° and freezes at 32°.

°F = (1.8 x °C) + 32°

FAN a power-driven machine which moves a continuousvolume of air by converting rotational mechanical energy to anincrease in the total pressure of the moving air.

FAN CAPACITY performance requirement for which a fan isselected to meet specific system calculations given in terms ofACFM at the fan inlet.

FAN CLASS operating limits at which a fan must be physicallycapable of operating safely.

FAN LAWS theoretical constant relationships between CFM,RPM, SP, and BHP for a given fan used in a given fixedsystem:

CFM varies as RPMSP varies as (RPM)2

BHP varies as (RPM)3

FC fan wheel design using forward-curved blades.

FINITE ELEMENT ANALYSIS (FEA) computerized analyticaltechnique used to divide a rotating body into many segments todetermine the stress of each segment and therefore thecomplete body.

FLASHING sheet-metal strip placed at the junction of inter-secting exterior building surfaces to make the joint water-tight.

FOOT-POUND (Ft . - Lb. ) torque rating or requirement;equivalent to the force required to move a one-pound weightone foot in distance, equal to 12 in.-lb.

FORCED DRAFT how air is provided in a process, such as acombustion process; when air is blown or forced into a process,it is known as a “forced draft” system. Also see induced draft.

FPM feet per minute; commonly defines air velocity (todetermine velocity pressure or suitability for material-conveying), shaft/bearing speeds (used to determine lubricationrequirements) and wheel tip speeds.

FRAME SIZE a set of physical dimensions of motors asestablished by National Electrical Manufacturers Association(NEMA) for interchangeability between manufacturers.Dimensions include; shaft diameter, shaft height, and motormounting foot print.

FREE FIELD the surroundings of a specific equipmentlocation in which no obstructions or reverberant surfaces exist todistort or amplify sound waves.

FREQUENCY any cyclic event whether vibration, alternatingcurrent, or rotational speed. Usually expressed in cycles persecond (cps) or just “cycles.”

FRICTION LOSS resistance to air flow through any duct orfitting, given in terms of static pressure.

FRP abbreviation for fiberglass-reinforced-plastic.

FULL-LOAD SPEED the speed at which the rated horsepoweris developed. This speed is less than synchronous speed andvaries with motor type and manufacturer.

FULL-LOAD TORQUE the torque required to produce therated horsepower at full-load speed.

FUMES airborne particles, usually less than 1 micrometer insize, formed by condensation of vapors, sublimation, distillation,or chemical reaction.

GALVANIZING the process of coating or plating with a zinc-rich solution; can be a hot-dip process, cold spray, or brushapplication.

GAS STREAM the specific airstream composition within anyfan or system.

GASES formless fluids which tend to occupy an entire spaceuniformly at ordinary temperatures and pressures.

GAUGE (GAGE) metal manufacturers’ standard measure ofthickness for sheet stock; some examples for steel are:

Gauge Thickness(Inches)

Weight of Steel(Lbs./Ft. 2)

7 .1793 7.5010 .1345 5.62512 .1046 4.37514 .0747 3.12516 .0598 2.50

GAUGE PRESSURE the pressure differential betweenatmospheric and that measured in the system.

GEL COAT a special resin system, sometimes includingpigment, but without glass-reinforcing, that is applied to themold before applying the FRP.

GROUND MOTOR a short circuit between any point in themotor’s electrical circuit and its connection to the ground.

HEAT EXCHANGER a device such as a coil or radiator whichis used to transfer heat between two physically separatedfluids.

HEPA FILTER high-efficiency particulate air filterscommonly called absolute filters.

HERTZ frequency measured in cycles per second.

Hg symbol for mercury. Pressure is often measured in inchesof mercury: (1" Hg. = 13.64" WG)

HORSEPOWER (as applied to motors) is an index of theamount of the work the machine can perform in a period oftime. 1 HP equals 33,000 ft. lbs. of work per minute, alsoequal to 0.746 kilowatts. Horsepower can be calculated by:

Torque (ft. lbs.) x RPMHP = 5250

HOUSING the casing or shroud of a centrifugal fan.

HVAC heating, ventilating, and air conditioning.

IMPELLER another term for fan “wheel.” The rotating portionof the fan designed to increase the energy level of the gasstream.

IMPELLER DIAMETER the maximum diameter measuredover the impeller blades.

IMPINGEMENT striking or impacting; such as materialimpingement on a fan wheel.

INCH OF WATER unit of pressure equal to the pressureexerted by a column of water one inch high at a standarddensity (27.73" water = 1 PSI).

INCH-POUND torque equal to one-twelfth foot pound.

INCLINED MANOMETER a metering device used to obtainaccurate pressure measurements.

INDUCED DRAFT how air is provided in a process, such as acombustion process; where air is drawn or pulled through aprocess. Also see forced draft.

INDUCTION the production of an electric current in a conductorin a changing magnetic field.

INERTIA tendency of an object to remain in the state it is in;see WR2 .

INLET-VANE DAMPER round multiblade damper mountedto the inlet of a fan to vary the airflow.

INSTABILITY the point of operation at which a fan or systemwill “hunt” or pulse; common in FC fans and some other fantypes where the point of operation is left of the peak of thestatic-pressure curve.

INTERFERENCE FIT specified interference between matingparts requiring either a press fit or a shrink fit.

KELVIN see Absolute Temperature.

KILOPASCAL Kpa; metric pressure unit; one inch watergauge is 0.24836 Kpa.

KILOWATT Kw; measure of power equal to 1.34 horsepower.

L-10 BEARING LIFE the theoretical number of hours afterwhich 90% of the bearings subjected to a given set of conditionswill still be in operation; also known as B-10.

LAMINAR FLOW gas or fluid in parallel layers with somesliding motion between the layers.

LAMINATE the total structure of the FRP part. For nybcorrosion-resistant products it consists of a resin-rich surfaceand a thickness of glass-reinforced resin as required forstructural strength.

LITHIUM a soft element in the alkali metal group commonlyused as a lubricant base.

LOGARITHM a mathematical term used as a basis of thedecimal system. A logarithm is the exponent of 10 whichproduces a given number. For instance the log of 100 is 2 since:

log10 100 = 2 10 2= 100

LOUVER a device comprised of multiple blades which, whenmounted in an opening, permits the flow of air but inhibits theentrance of undesirable elements.

LOWER EXPLOSIVE LIMIT the lowest percentage of anelement in otherwise standard air that will explode whenexposed to a spark.

MACH NUMBER a fraction of the speed of sound; used in fanengineering where air moving at a mach number of 0.9, or 9/10the speed of sound, begins to deviate from the fan laws.

MAKE-UP AIR a ventilating term which refers to thereplacement of air lost because of exhaust air requirements.

MANOMETER instrument for measuring pressure, u-shaped,and partially filled with liquid, either water, light oil, ormercury.

MAXIMUM CONTINUOUS RATING the point at which thefan is expected to operate.

MICROBAR a unit of pressure equal to one-millionth of anatmospheric pressure; 0.0000 146 PSI.

MICRON a unit of measure equal to one-millionth of a meter,commonly used in dust collection and material-handlingapplications to designate particle size.

MIL a unit of measure equal to 25 microns or one-thousandthof an inch.

MIXED-FLOW FAN a fan where the airflow is primarilyaxial and is changed by the blade shape to induce a small radialflow at the discharge.

MOLECULAR WEIGHT the weight of a molecule expressedon a scale in which the carbon isotope weighs exactly 12.0;represents the sum of the weights of all the atoms in a molecule.As air is a gas mixture, it does not have a true molecularweight but an apparent molecular weight determined by thepercentages of the molecular weights of each gas in acomposition.

NACE National Association of Corrosion Engineers.

NATURAL FREQUENCY the frequency at which a componentor system resonates.

NEC National Electrical Code.

NEMA the National Electrical Manufacturers Association;the trade association establishing standards of dimensions,ratings, enclosures, insulation, and other design criteria forelectric motors.

NOISE CRITERIA a way for an architect to specify themaximum permissible sound-power level in each of the eightoctave bands. NC curves give, in a graphical form, maximumpermissible intensity per octave band.

OCTAVE BANDS ranges of frequencies. These octave bandsare identified by their center frequencies (63, 125, 250, etc.).

OHM a measure of electrical resistance. A wire in which onevolt produces a current of one ampere has a resistance of oneOhm.

OPPOSED-BLADE DAMPER a type of damper whereadjacent blades rotate in the opposite direction.

OSHA Occupational Safety and Health Administration.

OSI ounces per square inch; a unit of pressure equal to one-sixteenth PSI or 1.733 inches of water.

PARALLEL-BLADE DAMPER a type of damper where theblades rotate in the same direction.

PARALLEL FANS two or more fans which draw air from acommon source and exhaust into a common duct or plenum. Aparallel fan arrangement is generally used to meet volumerequirements beyond that of single fans. Two identical fans inparallel will effectively deliver twice the rated flow of any oneof the fans at the same static pressure.

PERMANENT SPLIT CAPACITOR MOTOR very low startingtorque. Performance and applications similar to shaded polebut more efficient, with lower line current and higher horse-power capabilities.

pH a symbol as part of a logarithmic designation to indicateacidity or alkalinity on a scale from 0 to 14; pH7 is taken asneutral, 6 to 0 increasingly acid, 8 to 14 increasingly alkaline.

PHENOLIC a thermosetting resin system used for coatings andadhesives.

PIEZOMETER RING a device consisting of a number ofpressure taps connected to a common manifold to measurepressure.

PITCH DIAMETER the mean diameter or point at which V-beltsride within a sheave. This dimension is necessary for accuratedrive calculations.

PITOT TUBE a metering device consisting of a double-walledtube with a short right-angle bend; the periphery of the tube hasseveral holes through which static pressure is measured; thebent end of the tube has a hole through which total pressure ismeasured when pointed upstream in a moving gas stream.

PLENUM a chamber or enclosure within an air-handlingsystem in which two or more branches converge or wheresystem components such as fans, coils, filters, or dampers arelocated.

POINT OF OPERATION the intersection of a fan’s staticpressure curve and the system curve to which the fan is beingapplied; may be designated as velocity pressure divided bystatic pressure or by a given CFM and SP.

POLES the number of magnetic poles established inside anelectric motor by the placement and connection of thewindings.

POLYESTER a large group of thermosetting plastics whichexhibit a high degree of corrosion-resistance over a widespectrum of corrosive agents.

PSI pounds per square inch measured in gauge pressure, notincluding atmospheric.

PSIG pounds per square inch measured in gauge pressure, notincluding atmospheric.

PSYCHROMETRIC CHART a graphic depiction of therelationship between pressure, density, humidity, temperature,and enthalpy for any gas-vapor mixture, used extensively incomfort ventilation.

PULL-OUT TORQUE breakdown torque.

PURE TONE a sound that is characterized by a very uniformwave pattern. Such a sound might be created by a tuning fork.

PVC polyvinyl chloride; a synthetic thermoplastic polymer.

QUADRANT commonly the damper control plate.

RADIAL BLADE fan wheel design with blades positioned instraight radial direction from the hub.

RANDOM NOISE a sound that has an average amplitude andconstantly changing frequency.

RANKINE see Absolute Temperature.

RAREFICATION a phenomenon related to negative pressure.When air is drawn through resistance into a fan inlet, the air isstretched out, or rarefied, and becomes less dense than at theentry to the system. While negligible at low pressures andvolumes, high pressure fan selection must be based on rarefiedinlet density.

RELATIVE HUMIDITY the ratio of existing water vapor tothat of saturated air at the same dry-bulb temperature.

RESIN an organic polymer in liquid form which, when reactedwith the proper catalyst, becomes solid.

REYNOLDS NUMBER a mathematical factor used to expressthe relation between velocity, viscosity, density, and dimensionsin a system of flow; used to define fan proportionality.

ROCKWELL HARDNESS a standard measure of a metal’ssurface hardness. Also see Brinell Number.

ROTOR the rotating part of most AC motors.

RPM revolutions per minute.

RT fan wheel design with radial-tip blades.

RTP reinforced thermoset plastic. Also see FRP.

SATURATED AIR air containing the maximum amount ofwater vapor for a given temperature and pressure.

SCFM standard cubic feet per minute; a volume of air at0.075 lbs./ft. 3 density; used as an equivalent weight.

SCROLL the general shape of a centrifugal fan housing; theformed piece to which housing sides are welded.

SENSIBLE HEAT any portion of heat which effects a change ina substance’s temperature but does not alter that substance’sstate.

SERIES FANS a combination of fans connected such that theoutlet of one fan exhausts into the inlet of another. Fansconnected in this manner are capable of higher pressures than asingle fan and are used to meet pressure requirements greaterthan single fans.

SERVICE FACTOR the number by which the horsepowerrating is multiplied to determine the maximum safe load that amotor may be expected to carry continuously.

SHADED-POLE MOTOR a special type of single-phaseinduction motor. Low starting torque, low cost. Usually usedon direct-drive fans.

SHAFT SEAL a device to limit gas leakage between the shaftand fan housing.

SHUNT-WOUND MOTOR a DC motor in which the fieldcircuit and armature circuit are connected in parallel.

SI UNITS Systeme International d’Unites, InternationalSystem of Units; any one of the units of measure in theinternational meter-kilogram-second system.

SLIP the percentage difference between synchronous andoperating speeds.

SOUND produced by the vibration of matter. The vibrationcauses sound waves to spread through the surroundingmedium.

SOUND-POWER LEVEL acoustic power radiating from asound source. Expressed in watts or in decibels.

SOUND-PRESSURE LEVEL the acoustic pressure at a pointin space where the microphone or listener’s ear is situated.Expressed in units of pressure or in decibels.

SP static pressure; pressure as measured in all directionswithin an air-handling system, not including the force orpressure of air movement.

SPECIFIC GRAVITY the ratio of the weight or mass of agiven volume of any substance to that of an equal volume ofsome other substance taken as a standard. The ratio of thedensity of any gas to the density of dry air at the sametemperature and pressure is the specific gravity of the gas.

SPECIFIC HEAT the ratio of the quantity of heat required toraise a certain volume one degree to that required to raise anequal volume of water one degree.

SPI Society of the Plastics Industry.

SPLIT-PHASE MOTOR the most common type of single-phase induction motor. Moderate starting torque, high startingcurrent, high breakdown torque. Used on easy-startingequipment, such as belt-drive fans.

SQUIRREL-CAGE WINDING a permanently short-circuitedwinding, usually uninsulated and chiefly used in inductionmotors, having its conductors uniformly distributed around theperiphery of the machine and joined by continuous end rings.

SRC Spark-Resistant Construction; AMCA standard ofguidelines for general methods of fan construction whenhandling potentially explosive or flammable particles, fumes,or vapors.

SSPC Steel Structures Painting Council.

STANDARD AIR DENSITY 0.0750 lbs./ft.3, correspondsapproximately to dry air at 70°F. and 29.92 in. Hg.

STARTING TORQUE the torque produced by a motor as itbegins to turn from a standstill and accelerate. Sometimescalled locked rotor torque.

STATIC BALANCE the mechanical balance of a rotating part orassembly by adding weights to counter-balance gravitationalrotating of the part without power driving it.

STATIC PRESSURE the static pressure for which a fan is to beselected based on system calculations;

fan SP = SP outlet - SP inlet - VP inlet

STATOR the stationary parts of a magnetic circuit withassociated windings.

SURGE LIMIT that point near the peak of the pressure curvewhich corresponds to the minimum flow at which the fan canbe operated without instability.

SWSI Single-Width Single-Inlet Centrifugal Fans.

SYNCHRONOUS SPEED rated motor speed expressed inRPM. Synchronous speed = 120 x frequency divided bynumber of poles.

SYSTEM a series of ducts, conduits, elbows, filters, diffusers,etc., designed to guide the flow of air, gas, or vapor to and fromone or more locations. A fan provides the energy necessary toovercome the system’s resistance to flow and causes air or gasto flow through the system.

SYSTEM CURVE graphic presentation of the pressure versusvolume flow rate characteristics of a particular system.

SYSTEM EFFECT the effect on the performance of a fanresulting from the difference between the fan inlet and outletconnections to the actual system, and the standardizedconnections used in laboratory tests to obtain fan-performanceratings.

TACHOMETER an instrument which measures the speed ofrotation; usually in RPM.

TENSILE STRENGTH the maximum stress a material canwithstand before it breaks; expressed in pounds per squareinch.

TEST BLOCK an operating point above and beyond themaximum specified continuous rating demonstrating the fanmargin to the customer.

THRESHOLD LIMIT VALUES TLV; the values for airbornetoxic materials which are to be used as guides in the control ofhealth hazards and represent time weighted concentrations towhich nearly all workers may be exposed 8 hours per day overextended periods of time without adverse effects (OSHA).

TIP SPEED fan wheel velocity at a point corresponding to theoutside diameter of the wheel blades; normally expressed infeet per minute (circumference times RPM).

TORQUE a force which produces, or tends to produce, rotation;commonly measured in ft.-lbs. or in.-lbs. A force of one poundapplied to the handle of a crank, the center of which is displacedone foot from the center of the shaft, produces a torque of oneft.-lb. on the shaft if the force is provided perpendicular to, notalong, the crank. Torque can be calculated by:

HP x 5250Torque (ft. lbs.) = RPM

TP total pressure; the sum of velocity pressure plus staticpressure.

TUBEAXIAL FAN axial fan without guide vanes.

For m 507 DJ K

TUBULAR CENTRIFUGAL FAN fan with a centrifugalimpeller within a cylindrical housing discharging the gas in anaxial direction.

TURBULENT FLOW airflow in which true velocities at agiven point vary erratically in speed and direction.

UNBALANCE the condition of a rotor in which its rotationresults in centrifugal force being applied to the rotors support-ing bearings.

UNIFORM FLOW airflow in which velocities between anytwo given points remain fairly constant.

UNIT HEATER factory-assembled unit designed to heat andcirculate air. Types include steam, hot water, or gas fired.

UTILITY SET centrifugal fan designed as a packaged unit,ready to run.

VANEAXIAL FAN axial fan with either inlet or dischargeguide vanes or both. Includes fixed-pitch, adjustable-pitch, andvariable-pitch impellers.

VENA CONTRACTA the smallest flow area for flow through asharp-edged orifice.

VENTILATION supplying and removing air by natural ormechanical means to and from any space.

VIBRATION alternating mechanical motion of an elasticsystem, components of which are amplitude, frequency, andphase.

VINYL ESTER a significant variation of polyester providingincreased corrosion-resistance, strength, and flexibility, henceits suitability to the fabrication of FRP fan wheels.VISCOSITY the characteristic of all fluids to resist flow.

VOLT a unit of electrical potential or pressure. 110 or 220volts are normally found in the U.S.VP velocity pressure; the pressure or force of air in motion.The common equation based on standard air is:

VelocityVP = ( 4005 )2

VP/SP velocity pressure divided by static pressure; a singlenumber reference used to define a fan’s point of operation.Each system curve has a unique VP/SP value.

WATT a unit of power. In electrical terms, the product ofvoltage and amperage. 746 watts are equal to one horsepower.

WET-BULB DEPRESSION the difference between the dry-bulb and wet-bulb temperatures at the same location.

WET-BULB TEMPERATURE temperature at which air isbrought to saturation by evaporating a liquid into the air at thesame temperature.

WG water gauge; see Inch of Water.

WR2 the unit designation of fan wheel rotational inertia in lb.-ft.2 , also known as WK2 .

YIELD STRENGTH maximum stress to which a ductilematerial can be subjected before it physically distorts.

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