nyb fan handbook 2007

5/27/2018 NYB Fan Handbook 2007

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EN G IN EER IN G LETTER SThe New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521 -553

NUMBER

1

23

45

678

91011

1213

14

1516

1718

192021

2223

2425EG

SUBJECT

System Calculation

Fan Laws and System CurvesUnderstanding Fan Performance Curves

Temperature and Altitude Affect Fan SelectionFan Performance - The System Effect

Increasing Fan PerformanceField Testing of Fan SystemsProper Selection of Pressure Blowers

Pneumatic ConveyingFans and Blowers for Combustion ProcessSelection Criteria for Fan Dampers

Fan AcousticsFan Balance and Vibration

Stainless Steel Specifications for Fan Equipment

Practical Limits of Spark-Resistant ConstructionCorrosion-Resistant Coatings for Fan EquipmentCoating Surface Preparation Specifications

Corrosion Resistance of FRP Fans

Design and Construction ofnyb FRP FansAccessories and Construction Modifications for FRP FansSurface Veil for FRP Fans

Integral Motors for Centrifugal FansElectric Motor Codes and Standards

Fundamentals of SteamIndustrial Steam Heating SystemsMiscellaneous Engineering DataEngineering Letter Glossary

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ENGINEERING LETTERThe New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60527-553

S Y S T E M C A L C U L A T I O N

INTRODUCTION

A fan system is any combination of ductwork, hoods,filters,louvers, collectors, etc., that relies upon a fan to produce

airflow.When the air moves past each of these components,resistance is created which must be considered in system

calculations. It is also important to remember that fans are ratedindependently of a system and that fan performance will vary

depending upon the accuracy of the system calculations. ThisEngineering Letter will explain some of the basic fundamentals

of system design and calculation.

SYSTEM DESIGN

The purpose of the system will dictate the design criteria to be

used. Generally they will fall into one of the following two

catagories:

Velocity is typically the primary consideration in dust

collection, dilute pneumatic conveying, fume removal, andcontaminant applications. In these applications, a capture

velocity is required to redirect the flow of airborne materialsinto the duct system. In addition, a minimum conveying velocity

is necessary to maintain the flow of the materials within thesystem.

Given these velocity requirements, system components can beselected to maintain the appropriate air volume andrequired

velocity 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 cubic

feet per minute (ACFM), the mass of air required must beconverted from standard cubic feet per minute (SCFM) to

ACFM.

The velocity through a system can be determined once

ACFM is known. The relationship between velocity and airfis defined by the equation:

Q = VA

where: Q = ACFM

V = velocity in lineal feet per minute

A = cross-sectional area in square feet

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

System design is really a matter of defining the required work

terms of volume or velocity and then sizing and selecting necessary system components to accomplish that work. course, this must be done within the economic and sp

constraints of the installation.

DETERMINING SYSTEM RESISTANCE

System resistance is the sum of the resistance through ecomponent within the system. The system depicted in Figur

may appear complex, but dealing with each componseparately provides an orderly process for determining

overall resistance.

HOOD LOSS

To determine hood or entrance losses, resistance calculatimust be made for both the acceleration loss and the entry lo

Since the air or atmosphere surrounding the hood mustaccelerated from a state of rest, energy will be required to

the air in motion. This energy is equal to the velocity pressurthe entrance to the duct. Assuming the hood in this exam

empties into a 7" diameter duct, the required 1165 ACresults in a velocity of 4363 FPM:

V = Q A where: Q = 1165 C

Figure 1 Typical System


3/112

(3.5 in. radius)2 x 3.1416A =

144 in.2 /ft.2 = .267 ft.2

therefore: V = 1165 CFM .267 ft.2

= 4363 FPM

The 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 in

Figure 3. The relative efficiencies are expressed as losses inpercentage of the duct velocity pressure. Consequently, the

lowest percentage is actually the most efficient.

OutletVelocity

VelocityPressure

OutletVelocity

VelocityPressure

OutletVelocity

VelocityPressure

800 .040 2800 .489 4600 1.32

1000 .063 3000 .560 4800 1.441200 .090 3200 .638 5000 1.56

1400 .122 3400 .721 5200 1.691600 .160 3600 .808 5400 1.82

1800 .202 3800 .900 5600 1.952000 .250 4000 .998 5800 2.10

2200 .302 4200 1.10 6000 2.242400 .360 4400 1.21 6200 2.40

2600 .422Figure 2 Acceleration Loss

Figure 3 - Entrance Loss Percentage

The hood in this example is most similar to item 2 in Figure

Therefore, the entry loss from atmosphere into the hood is

times the entering air velocity pressure at 1000 feet per minuor:

1000 2Entry Loss = .90 x

4005 ) = .06" W.G.

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

Figure 3.

The total hood loss in the example is the acceleration loss adto 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

determined by using the equivalent length method. This run

duct includes 62' of 7" diameter duct and one 4 piece 90 elbof R/D = 2. According to Figure 4, the elbow has a loss equa12 diameters of 7" duct, or 7'. Thus, the total equivalent len

of 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' ofdiameter duct handling 1165 CFM. The loss for this run can

determined as:

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

Therefore, the total resistance of the hood branch to the junctis:

Branch Loss = 1.25" + 2.76" = 4.01" W.G.

SECONDARY BRANCH

A secondary branch is calculated in the same manner as main branch. For example, a grinder hood handling 880 CF

through a 6" pipe results in a velocity of 4500 FPM, which ha1.26" VP.

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

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

Page 2


4/112

The duct branch from the grinder hood to the junction consists

of 27' of 6" pipe and (2) 4 piece 90 elbows of R/D = 2. With an

equivalent 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 equal

during operation or the system will automatically equalize byadjusting the flow different than the design point. If the

variation in resistance between any two converging branches

exceeds 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 2045

ACFM (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 results

in:4630 2

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

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

CFM in Upstream

Main CFM in Branch

Loss in Main

in % of Main V.P.

1 .202 .17

3 .15

4 .145 .13

6 .12

7 .118 .10

9 .10

10 .10CORRECTION FACTORS FOR OTHER THAN 45 TEE.

Tee Angle 45 Loss X Factor

0 0

15 0.130 0.5

45 1.0

60 1.7

75 2.590 3.4

Figure 5

Chart II on page 4 indicates a resistance of 3.3" for every 100

9" diameter duct handling 2045 CFM. According to Figure 4

two elbows are equal to another 18' of duct, so the toequivalent 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, expressed as negative pressure. Also that only the branch w

the 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,

30 of duct and the one elbow will have a loss equivalent to following:

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

The pressure drop across the dust collector, like coils or filtmust be obtained from the manufacturer of the devAssuming a 2.0" loss for this example, the resistance at the

outlet is:

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

FAN SELECTION

At this point enough information is known about the system

begin fan selection. Because fans are rated independent osystem, their ratings include one VP to account for accelerati

Since the system resistance calculations also consiacceleration, fan static pressure can be accurately determined

follows:

Fan SP = SP outlet - SP inlet - VP inlet

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

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

For this example, a fan should be selected for 2045 ACFM8.5" SP and have an outlet velocity of at least 4500 FPM

prevent material settling. This presumes a standard airstredensity of .075 lbs./ft.3 If the density were other than standa

the system-resistance calculations would have been the same the resulting fan SP would have been corrected. Refer

Engineering Letter 4 for density correction procedures.

This example also assumes that the fan inlet and ou

connections are aerodynamically designed. Fans are sensitivabrupt changes in airflow directly adjacent to the fan inle

outlet. The effects of abrupt changes and other system effeproblems are discussed in Engineering Letter 5.

CONCLUSION

It is the responsibility of the system designer to ensure that th

are adequate air flows and velocities in the system and that selection of duct components and fan equipment has b

optimized. While computer programs do the bulk of syscalculations today, this Engineering Letter should help

provide a common set of methods and terminology to assisthat effort.

Page 3


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6/112

ENGINEERING LETTER 2The New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521-553

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 a

given system. With a basic understanding of these rules, theperformance of a fan can be quickly calculated for various

conditions.

SYSTEM REQUIREMENTS

The three fundamental rules governing fan performance arecommonly called the fan laws. These rules are only valid

within a fixed system with no change in the aerodynamics orairflow characteristics of the system. For the purpose of this

discussion, 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, filter

media, etc.) and any other air molecules. Energy is required toovercome this friction, or resistance. The faster the air moves

the 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 the

pressure that results in air velocity is described as velocitypressure (VP). The portion necessary to overcome friction in the

air 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 flow

g = force of gravity

h = pressure causing flow

As can be seen from the equation, the pressure necessary to

cause flow is proportional to the square of the velocity. In asystem, this means that SP will vary as the square of the change

in velocity or volume expressed in cubic feet per minute (CFM).This makes it possible to predict all possible combinations of SP

at 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 is

desired 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 ratin

would result in a plotted curve as shown in Figure 1.Regardless of fan type, fan size, or volume of flow through a

system, the relationship of CFM to SP will not change unless system itself is altered in some way. SP always varies as the

square of the change in CFM. The only exception to this rule found in a laminar flow characteristic where VP is of far great

importance than SP. Such circumstances are not typical of fansystems.

FAN LAWS

In air movement systems, it is the fan wheel that does the wo

In a sense, the fan wheel acts like a shovel. As it revolves

discharges the same volume of air with each revolutiWorking within a fixed system, a fan will discharge the savolume of air regardless of air density, (disregarding the effe

of compression at high pressures).

If the fan RPM is increased, the fan will discharge a greavolume of air in exact proportion to the change in speed. Thi

the 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.


7/112

As shown earlier, in a system, the SP varies as the square of

the 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). As

the fan speed changes, this relative point of operation remainsunchanged as long as the system remains unchanged. Thus, the

fan 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 be

considered independently.FAN CURVE AND SYSTEM CURVE

As stated previously, a system curve can be plotted to show all

possible combinations of SP and CFM for a given fixedsystem. Any fan used on that system must operate somewhere

on that system curve.

Fan performance is determined by laboratory testing and is

presented graphically in the form of fan curves. Unless it isphysically altered in some way, a fan must operate somewhere

on 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 somewhere

on 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 and

the fan SP CFM curve.

If the fan speed is increased or decreased, the point

operation will move up or down the existing system curThis is shown in Figure 4.

The following are examples of how the fan curve can be used

calculate changes to flow and pressure requirements.

Example 1: A fan has been selected to deliver 35,530 CFM a

8" SP. The fan runs at 1230 RPM and requires 61.0 BHP.

After installation, it is desired to increase the output 20%.what RPM must the fan run? What SP will be develop

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 CFM2 1 / 4 "SP and to run at 877 RPM, requiring 5.20 BHP.

After installation, it is found that the system only deliv

9,150 CFM at 2 1/2" SP and uses 4.70 BHP. This indicates original calculations were in error, or that the system was

installed according to plan. What fan RPM and BHP willnecessary to develop the desired 10,300 CFM? What

should 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 nomodified fan. Adding or deleting system components such

dampers, or incurring density changes, will create completnew system curves. Changing fan accessories such as in

boxes, evases, or inlet dampers will alter the fans performa

curve from standard. These variables must be considered befthe fan laws can be applied.

During the process of system design, the fan laws can

helpful in determining alternate performance criteria ordeveloping a minimum/maximum range. If safety factors

applied to system calculations, it should be recognized tha10% factor on volume will result in an increase in horsepow

of 33% according to the third fan law. An evaluation shouldmade weighing the necessity of the safety factor versus

cost penalty incurred.

F o r m 6 0 7 G

Figure 3

Figure 4


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ENGINEERING LETTER The New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521 -55

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 S

INTRODUCTION

One of the most important documents customers request from

fan 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 the

performance 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 the

Air Movement and Control Associations Standard 210 to assureuniform and accurate readings. Data points are collected at a

given RPM while the flow is slowly modulated from full closedto full open. The information gathered is then used to develop

computer 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 and

pressure calculations. This curve is constructed by plotting aseries of static pressure points versus specific flow rates at a

given test speed. While the static pressure curve depicts a fansperformance at a given speed, it can be used to determine thefans pressure capability at any volume.

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

1. CFM varies as RPM



To locate a fans point of operation, first locate the requiredstatic pressure on the SP scale at the left of the curve. Then draw

a horizontal line to the right, to the point of intersection with theSP curve. Next, draw a vertical line from the point of operation

to the CFM scale on the bottom to determine the fans flowcapability for that SP at the given speed.

As shown in Figure 1, the performance for this fan is 8750 CFM

and 12" SP at 1750 RPM.

Figure 1 - Static Pressure Curve

Assuming this same fan was intended to operate at 1200 RPM,

fan laws can be applied to obtain performance at this lower spee

1. CFM varies as RPM

CFM (new) RPM (new)

CFM (old) =

RPM (old)

Therefore:

1200CFM (new) =

1750 (8750) = 6000 CFM


SP (new) RPM (new)

P (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 cbe established. An accurate BHP rating is necessary to prope

size the motor or to determine the operating efficiency of one as compared to another. Performance curves contain a B

curve from which the BHP rating can be determined for speccapacities. To determine BHP at a specific point of operation

horizontal line is drawn to the right from the point intersection of the vertical CFM line and the BHP curve.


9/112

Figure 2 - Performance Curve

As shown in Figure 2, the fan operating at 8750 CFM and 12" SP

at 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 fans capabilities within a

given system must be provided. The fan laws apply equally toany system; therefore, CFM and SP variations within the system

are predictable. This enables system lines to be superimposed onperformance curves to simplify performance calculations. The

system line is nothing more than the sum of all possible CFMand SP combinations within the given system. Any combination

of fan and system must operate somewhere along that system line.

Because a fan must operate somewhere along its SP curve and

since the system has a known system line, their intersection isthe point of operation. If the fan speed is changed, the point of

operation must move up or down the system line. If the system ischanged in some way, the point of operation must move up or

down 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 for

8750 CFM and 12" SP operating at 1750 RPM. Included inFigure 3 are a number of different system lines. If the system

does not operate properly upon start-up, measurements can betaken and compared against the available performance curve.

Figure 3 - Performance Curve with System Lines

Lets assume that a tachometer reading indicates the fanrunning at 1200 RPM instead of 1750 RPM, because of mista

in motor speed or drive selection, and an airflow check indicaonly 6000 CFM. These readings confirm that the system w

calculated correctly and that the fan speed must be corrected1750 RPM as originally specified to achieve the desired 87

CFM. If the tachometer reading indicates the proper speed bthe airflow reading is down, additional system resistance beyo

that originally calculated is indicated. This additional resistacould be caused by partially closed louvers/dampers, change

duct sizing from the original design, system effect losses, or jan error in the system-resistance calculations. The deficiency

usually be corrected by either altering the system or increasthe fan speed.

Often, performance curves for one speed must be used

determine the performance of a fan for use on systems requirmore air or higher pressures. A performance curve such

Figure 4 can be used to determine fan performance beyond thescale shown by using the fan laws to obtain a reference point

operation on the system line. This can be accomplished applying some suitable factor to the required CFM and

square of that factor to the required SP.

For example, the performance curve shown in Figure 4 canused to determine fan performance requirements for a syst

calculated at 15,000 CFM and 23.5" SP, even though that pois beyond the curve. By determining a suitable refere

capacity using the fan laws, that falls within the curve data, performance requirements can be obtained at the curve sp

and then projected up to the system requirements using the laws once again.

The required 15,000 CFM and 23.5" SP is on the same syst

line as 10,000 CFM at 10.4" SP which intersects the fans curve drawn for 1750 RPM and has a corresponding BHP

33.0 at 1750 RPM. Therefore:

15000RPM (new) =

10000 (1750) = 2625 RPM

15000BHP (new) = ( 10000 )

3(33.0) = 111 BHP

Page 2


10/112Page 3

FAN PERFORMANCE CHARACTERISTICS

The performance characteristics of a fan can be determined

from 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 the

air 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 of

these general fan types. Further, these performance characteristicsmay vary from one manufacturer to the next depending upon

the particular design. The characteristics described in this letterare based on nybfan equipment.

The performance curves presented in Figures 1 through 4 are

typical of fans with radial-blade wheels. The SP curve issmooth and stable from wide open to closed off. The BHP

curve 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 static

pressure peak. This sharp dip (shaded area) indicatesunpredictable flow characteristics. Though not technically

accurate, this region is often referred to as the the stall region.It indicates that at these combinations of pressure and relatively

low volumes, the airflow characteristics across the wheel bladeschange or break away so that the fan performance point is no

longer 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. This

indicates 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 resistance

will 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, the

majority of fans with backwardly-inclined wheels exhibit an SP

curve 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 smoother

depression to the left of the static pressure peak. This indicatesthat the overall fan design is such that internal flow

characteristics 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 1,000s

Figure 4 Typical Radial-Blade Fan Performance Curve

Figure 5 Typical FC Fan Performance Curve

CFM

CFM

Figure 6 Typical BI Fan Performance Curve

Static

Pressure

Static

Pressure

Brake

Horsepow

BrakeHorsepow

Brake

Horsepow


11/112

The BHP curve for all backwardly-inclined fans is the major

difference between them and all other fan types. As shown in

Figures 6 and 7, the BHP curve for backwardly-inclined fansreaches a peak and then drops off as the fans volume increases.With this non-overloading BHP characteristic, it is possible to

establish a maximum BHP for a given fan speed and select amotor that can not be overloaded despite any changes or errors in

system 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 on

the SP curve. Although stability or performance may not be aproblem, a point of operation down to the far right on the SP

curve should be avoided. Selecting a fan that operates far downto the right, eliminates the flexibility to compensate for future

system changes. Also, the fan is less efficient in this area ascompared to a larger fan operating at a slower speed. Figure 7

shows 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 fans

BHP curve. The radial tip is often more efficient than radial fansand typically best applied in high-pressure applications. As a result

of its efficiency and dust-handling capabilities, the radial-tip fancan also be applied to lower pressure material conveying

systems.

The term axial fan is used to describe various propeller,vaneaxial, tubeaxial, and duct fans. The performance curves of

these fans are characterized by the ability to deliver largevolumes of air in relatively low pressure applications. As can be

seen in Figure 9, the axial flow fan is distinguished by a

drooping 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 the

curve. Depending upon the severity, axial fans are normally onlyselected to the right of this area.

CONCLUSION

A good working knowledge of performance curves is necessary

to understand the performance characteristics and capabilities ofdifferent fan equipment. Use of performance curves in the

selection of fan types and sizing will assure stable and efficientoperation as well as future flexibility.

F o r m 6 0 7 G

CFM

Figure 7 Typical AcoustaFoil Fan Performance Curve

CFM

Figure 8 Typical Radial-Tip Fan Performance Curve

CFM

Figure 9 Typical Axial Fan Performance Curve


12/112

ENGINEERING LETTER The New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521 -5

TEMPERATURE AND ALTITUDE AFFECT FAN SELECTION

INTRODUCTION

Fan performance changes with the density of the gas beinghandled. Therefore, all fans are cataloged at a standard condition

defined as: 70F. air, at sea level, with a gas density of .075lb./ft.

3at a barometric pressure of 29.92" Hg. At any other

condition, the fans horsepower requirement and its ability todevelop pressure will vary. Therefore, when the density of the

gas stream is other than the standard .075 lb./ft.3, correction

factors must be applied to the catalog ratings in order to select

the 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 too

pliable at temperatures other than 70F. Temperature deratefactors must be applied to the fans catalog maximum safe

speed to ensure against overspeed situations.

HOW TO CALCULATE ACTUAL FAN PERFORMANCE

AT OTHER THAN 70 DEGREES FAHRENHEIT

As illustrated in Figure 1, a fan wheel is similar to a shovel. In

a given system, it will move the same volume of air regardlessof the airs weight. If a fan moves 1000 CFM at 70F., it will

also move 1000 CFM at 600F.

However, air at 600F. weighs half as much as it does at 70F.Therefore, the fan requires just half the horsepower. (See

Figure 2.) Likewise, since the air weighs half as much, it willcreate only half the static and velocity pressures. The reduction

in 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, 70F. air, delive12,400 CFM against 6" SP (static pressure) requiring 14.6

BHP (brake horsepower). If the system and fan RPM are notchanged, but the inlet airstream temperature is increased to

600F., how will the fan perform?

The fan will still deliver 12,400 CFM, but s ince the air

600F. weighs half as much as the air at 70F., static pressand horsepower will be cut in half. The fan will generate only

SP and require only 7.3 BHP.

A typical fan specification based on hot operating conditions

illustrated in Example 2.

Example 2. Required: 11,000 CFM and 6" SP at 600F. (T

means the actual, measurable static pressure of the fan600F. will be 6 inches of water.)

The fans catalog performance tables are based on 70F. ai.075 density. The specified SP must be corrected by the ra

of the standard density to operating density. Since densities inversely proportional to absolute temperature (degrees F

460):

460 + 600 10606 ( 60 + 70 ) = 6 ( 530 ) = 12

The fan must be selected from the rating tables for 11,0CFM at 12" SP. The BHP obtained from the table should

multiplied by the ratio of operating density to standard densin order to obtain the BHP at 600F. If the rating table show

30.0 BHP, the operating BHP would be 30.0 (530/1060) = 1BHP.

In most hot systems, the fan is required to handle cold until the system reaches temperature. A good example is

oven exhaust systems.

Figure 1 - A fan wheel is like a shovel.

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


13/112Page 2

If Example 2 were such a case, the fan would require 30.0 BHP

when operating at 70F., and 15.0 BHP when the oven hadwarmed to 600F. Very often a damper is furnished with the fan

so that, during the warming-up period, the fan can be damperedto reduce the horsepower. Without the damper, a 30 HP motor

would be needed.

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

either:

11,000 CFM and 6" SP at 600F., or

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

Table 1 gives correction factors used to convert from a non-standard density to a standard density of 70F. air. These factors

are merely the ratios of absolute temperatures. Multiply theactual static pressure by the specific temperature/altitude factor

so standard catalog rating tables can be used. Divide the brakehorsepower from the catalog rating table by the

temperature/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 1.58

60 0.98 400 1.62

70 1.00 450 1.72

80 1.02 500 1.81100 1.06 550 1.91

120 1.09 600 2.00

140 1.13 650 2.09

160 1.17 700 2.19

180 1.21 750 2.28200 1.25 800 2.38

225 1.29 900 2.56

250 1.34 1000 2.76

Table 2 - Corrections for Altitude

AltitudeFeet Above

Sea Level

FactorAltitude

Feet Above

Sea Level

Factor

0 1.00 5000 1.20

500 1.02 5500 1.22

1000 1.04 6000 1.25

1500 1.06 6500 1.27

2000 1.08 7000 1.30

2500 1.10 7500 1.32

3000 1.12 8000 1.35

3500 1.14 8500 1.37

4000 1.16 9000 1.40

4500 1.18 10000 1.45

HOW TO CALCULATE ACTUAL FAN PERFORMANCAT OTHER THAN SEA LEVEL

A fan operating at an altitude above sea level is similar to a operating at air temperatures higher than 70F.; it handlesless dense than standard. Table 2 gives the ratio of standard

density at sea level to densities of 70F. air at other altitudes.

Example 3. Required: 5800 CFM at 6" SP at 5000 ft. altitu70F. air at sea level weighs 1.20 times as much as 70F. air

5000 Ft. Therefore, at sea level, the SP is 1.2 x 6 = 7.20" SThe fan would need to be selected for 5800 CFM at 7.2" SP

70F. .075 density.

When both heat and altitude are combined, the density of the

is modified by each, independently, so that the correction factcan be multiplied together.

Example 4.Required: 5800 CFM at 6" SP at 5000 ft. altitud

600F. Air at 70F. at sea level weighs 2.00 x 1.20 = 2.40 timas much as 600F. air at 5000 ft. altitude. At sea level and 70

SP = 2.40 x 6 = 14.4" SP. Select a fan for 5800 CFM at 14SP. Divide the brake horsepower in the rating table by 2.40

obtain horsepower at 600F. and 5000 ft. If the fan is to stcold, it will still be at 5000 ft. altitude. Therefore, to get

cold horsepower requirement, divide by 1.20, the altitufactor only.

DENSITY CHANGES FROM OTHER THAN HEAT ANALTITUDE

Fan densities may vary from standard for other reasons than h

and altitude. Moisture, gas, or mixtures of gases (other than are a few possibilities. In these cases, it is necessary to obt

the actual density of the airstream gas by some other referenmaterial. A similar factor, as shown in Table 1, is then crea

using the standard density of air .075 lb. per cubic foot dividedthe new density.

.075 lb./ft.3

Factor =special gas density

ACFM-SCFM DEFINITION

The terms ACFM and SCFM are often used in design work an

cannot be used interchangeably.

SCFM is Standard Cubic Feet per Minute corrected to standdensity conditions. To determine the SCFM of the volume uin Example 2, which was 11,000 CFM at 600F., we wo

multiply the CFM by the density ratios.

.03711000 x

.075= 5500 SCFM

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

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

density.

The terms ACFM and SCFM are often used in system des

work where both quantities need to be known. It should

remembered, however, that since a fan handles the same voluof air at any density, ACFM should be used when specifyand selecting a fan.


14/112Page 3

FAN SAFE SPEED AND TEMPERATURE

Whenever a fan is used to move air at temperatures substantially

above or below 70F., 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 determined

by calculations or actual tests. Safe speed depends entirely uponthe wheel and shaft assemblys ability to withstand the

centrifugal forces created by its own weight. Highertemperatures can affect the wheel and shaft assemblys ability to

withstand these forces and therefore must be considered.

Most metals become weaker at higher temperatures. Thisweakness is measurable in terms of yield and creep strength. It

can be translated into formulas that accurately determine the safespeed of a wheel and shaft assembly in relation to its tested

maximum speed at standard conditions. Manufacturers providesafe speed reductions in their catalogs based on the alloy that

was used to manufacture the wheel and/or shaft.

Some metals withstand heat better than others. Certain grades of

stainless steel can be substituted to increase temperature limits.On the other hand,fan wheels constructed of aluminum should

never be operated above 200F.

For information regarding fiberglass reinforced plastic fan

equipment, 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.M i ld S t ee l A lum inum

304L 316L 347

70 1.0 1.0 1.0 1.0 1.0

200 .97 .97 .88 .95 .95300 .95 -- .82 .92 .93

400 .94 -- .78 .89 .90

500 .93 -- .75 .86 .90

600 .92 -- .73 .84 .90

800 .80 -- -- .79 .86

1000 -- -- -- .75 .83

The limiting temperature on any fan is the highest temperaturethat any component of the fan assembly will reach during any

operating cycle. A fan in a process oven application may handleair several hundred degrees above the highest temperature the

oven reaches, especially during start-up. On such applications, a

temperature 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 are

located on the inlet side of the fan. In all cases, the fan shouldremain in operation until the air is cooled to 180F. or less to

prevent 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 bearing

designs, as well as high temperature lubricants, are available toextend the operating range to higher temperatures.

Arrangement 4 centrifugal fans, where the fan wheel is moun

on the motor shaft, should not be used above 180F., unlspecial provisions are made (i.e., a shaft cooler or heat shield

keep heat radiated from the housing from increasing mobearing and winding temperatures.

When fan bearings are located outside of the airstream, as

Arrangement 1, 8, and 9 centrifugal fans, higher airstretemperatures are possible. Table 4 lists some typical maxim

recommended operating temperatures for fans using ball

roller bearings.

A conventional fan using standard bearings and standard lubriccan normally be operated to a maximum of approximat

300F. With the addition of a shaft cooler (Figure 3), ttemperature limitation can be extended to 650F. The sh

cooler has the effect of absorbing and dissipating heat from shaft while circulating air over the inboard bearing.

Table 4 - Maximum Fan Inlet Temperatures

Arrangement 1 and 8 (Overhung Wheel)

Standard ConstructionWith Shaft Cooler

With Shaft Cooler and Heat GapWith Shaft Cooler, Heat Gap,

Stainless Wheel, and Alloy Shaft

300F.650F.

800F.

1000F.

Arrangement 3 (Wheel Suspended Between Bearings)

Standard Construction 200F

Arrangement 4 (Wheel on Motor Shaft)

Standard Construction 180F

Enclosed Bearing Fans (Axial Fans)

Arrangement 4 105F

Arrangement 9 120F

ith Special V -Belts with 2.0 S.F. 200FArrangement 9 Duct Fan

With Heat-Fan Construction 600F

Plenum Fans

Arrangement 3 105F

Arrangement 4 105F

Figure 3 Shaft Color


15/112Page 4

With the addition of a heat gap (Figure 4) the temperature

limitation can be extended to 800F. 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 and

shaft 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 much

more critical.

When the fan shaft is heated to the point that it expands more

than the structure to which it is attached, one expansion bearingand one fixed bearing should be furnished. The fixed bearing is

located on the drive end of the fan while the floating bearing islocated next to the fan. This arrangement, however, is not

critical and may vary by manufacturer.

When the fan is handling air below 70F., there is the possibility

of other problems. Below -30 to -50F., ordinary steel is toobrittle. Aluminum wheels or wheels of steel containing at least

5% nickel must be used, and shafts must be made of nickel-

bearing steel. In addition, lubricants become stiff, or even solid

these low -temperature applications. Exact operating conditioshould be given to the fan manufacturer to relay to the bearin

supplier for proper selection.

CALCULATING HOT RESISTANCE FOR SYSTEM

Figure 5 shows a system that operates at the same temperat

throughout. If the inlet temperature is known, the fan may

selected from the fan capacity tables and the rated horsepower static pressure corrected by the temperature correction facfrom Table 1. However, what happens to the system that the

was operating against? If a fixed system, which originally wcalculated for standard air, was subjected to the same temperat

increase as the fan, then system static pressure will change andidentical to the fan static pressure change. The result is that

fan and system operate together the flow will remain unchang(See Figure 6.) Unfortunately, this example assumes that

entire system is being subjected to the same temperature chanwhich is not always the case.

Figure 4 Heat Gap between fan andbearin .

Figure 6 Fan-system curve relationship

with fan and system at the same temperature.

Fi ure 5 A s stem with the same tem erature throu hout.

Figure 4 Heat Gap betweenfan and bearing.


16/112

Page 5

Figure 7 shows a system in which different temperatures are

involved. The fan will not handle the same volume of air whenoperating hot as it does when cold. If the burner is on, the fan

will handle 1430 ACFM against an actual static pressure of 1.2inches. This is arrived at by adding the filter, burner, and nozzle

resistance, neglecting for the sake of simplicity any externalresistance from additional ductwork. The fan would be selected

from the capacity tables on the basis of 1430 CFM at 1.72inches static pressure (300F. correction factor times 1.2

inches).

If the burner is turned off while the fan continues to operate atthe same RPM, it is necessary to determine the system

characteristic curve and plot its intersection with the fan todetermine how much air the fan would move and at what static

pressure. To accomplish this we must assume an arbitrarycapacity, such as 1000 CFM at 70F. The filter louver resistance

would be the same, cold or hot, at .3 inches 70F. The burnerresistance would remain unchanged with temperature since it must

be assumed that air expansion takes place after the hi

velocity section of the burner. The nozzles will vary resistance directly as the density changes and inversely as t

square of the flow. The nozzle would then have a resistance coat 1000 CFM of:

1000.5 x ( 1430 )

2x 1.43 = .35

Summing these resistances yields the cold resistance at 10CFM of 1 .05"SP. This new system point and correspond

curve are then plotted against a fan curve at standard conditions

such that the resulting intersection will be the final operat

point of the cold system. Using an actual fan as an example,the resulting flow would be 1220 CFM at 1.5 inches st

pressure. (See Figure 8.)

Figure 7 - A system with different temperatures.

Figure 8 - Fan-system curve relationship with

fan at different temperatures.


17/112

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 70F. to 600F. Obviously, the heater

will 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 70F. against 3 inches of static pressure at

70F. One possible selection is a fan with a 27-inch wheel

diameter, Class I design utilizing a 71/ 2HP motor.

The alternative fan, pulling air from the heater, would specified to handle 20,000 ACFM at 600F. against 3" SP600F. It would be selected from the capacity tables for 20,0

CFM at 6" SP. One suitable choice is a fan with a 3 61/ 2-inwheel diameter, Class II design utilizing a 15 HP motor. (No

26 HP, from the tables, at 70F., divided by temperat

correction factor, is 13 HP at 600F.) This example illustrawhy it is usually more economical to locate the fan at

coolest part of the system. In this case, the push fan micost half as much as the pull fan.

Figure 9 - The importance of fan location.

F o rm 6 0 7 G A W


18/112

ENGINEERING LETTER The New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521-5

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 T

INTRODUCTION

Fans are typically tested and rated in prescribed test configurationsdefined by the Air Movement and Control Association. This is

done to ensure standardized procedures and ratings so thatsystem designers can make realistic choices among various

manufacturers. Beyond the routine system resistance calculations,the location of some common components and their proximity to

the fan inlet or outlet can create additional immeasurable lossescommonly called System Efect. These losses, if not eliminated

or minimized, will necessitate fan speed and horsepowerincreases to compensate for the performance deficiencies. This

Letter will outline some of the common causes for thesedeficiencies and provide useful guidelines for more efficient and

predictable air-handling systems.

SYSTEM DESIGN

The 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 air

through an opening in the wall of a building, or as involved as amulti-zoned system with varying flows and densities. The

calculations for determining the performance requirements arediscussed in Engineering Letter 1. The effects of the system

design 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 the

appropriate fan. However, in many cases the effects of therelationship between the system components and the fan are not

considered in the calculation or selection process. For example,the resistance of a given size elbow at a given flow can be easily

determined using the equivalent length calculation method.However, if that elbow is located at the fan inlet or outlet, further

immeasurable losses will be imposed in addition to the simpleloss through the elbow itself. Most importantly, these losses

cannot be measured or even detected with field instrumentsbecause they are, in fact, a destruction of the fan performance

characteristics.

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 fan

equipment 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 field

performance problems.

The four most common causes of system-induced performance

deficiencies:

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. should be drawn into the fan inlet with an evenly distribu

velocity profile. As shown in Figure 1, this allows all portionsthe fan wheel to handle an equal air load.

If the air is not drawn into the fan inlet evenly, performandeficiencies will result from the combined effects of turbulen

and uneven air distribution. This is illustrated in Figure 2, whan elbow is installed directly on the fan inlet.

Figure 1 Even Air Loading

Figure 2 Uneven Air Loading


19/112

When the system attempts to change the direction of flow, the

air hugs the outside of the inlet elbow entering the fan. Thiscauses uneven, turbulent airflow into the fan. Another common

cause of non-uniform flow into the fan inlet is a poorly designedinlet box, such as the one shown in Figure 3. It is important to

remember 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 a

vortex-type inlet damper.

The direction air is flowing when it enters the fan wheel is very

important. In order to produce its rated capacity, the fan workson the air by changing its direction and accelerating its velocity. If

the air is spinning in the same direction as the wheel rotation, thefan capacity will be diminished. If the air is spinning in the

opposite direction of the wheel rotation, the brake horsepowerand noise of the fan will increase. The static pressure of the fan

may 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 connections

or components that can contribute to pre-spin. Also, spinningoften occurs in combination with eccentric flow such as the case

with the inlet box shown in Figure 4.

Pre-spinning flow can result from any number of comm

situations. Two elbows in close proximity to one another force the air to make consecutive turns in perpendicular pla

to form a corkscrew effect. As shown in Figure 5, air convergtangentially into the main duct or plenum can create an obvi

spinning effect.

Pre-spinning flow can also be induced by such common

cleaning devices as a venturi scrubber or a cyclone as seen

Figure 6. In these cases, it is often the very function of the cleaning 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


20/112

CORRECTING BAD INLET CONNECTIONS

The ideal fan inlet connection creates neither eccentric norspinning flow. Where an inlet duct is required, the best

connection is a long straight duct with straightening vanes.However, it is usually necessary to adapt the system to the

available 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 represent

some 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 fan

static-pressure curve. Figure 7 illustrates the recommended rangefor backwardly-inclined fans. A deleterious system effect could

move the point of operation to the left on the pressure curve.This would force the fan to operate at an unstable point. The same

situation can occur with any of the basic fan types that exhibitunstable flow characteristics as discussed in Engineering Letter

3. 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/or

spinning flow. The egg-crate straightener, such as the one shownin Figure 6, can be used in the available space to stop pre-spin

and improve fan inlet conditions.

Most of the inlet connections illustrated, with or without

corrective devices, can produce losses in performance. Theselosses would be difficult, if not impossible, to predict. Even the inlet

box 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 speed shown in the fans rating table at the required volume anpressure 21% greater than originally calculated:

(110% 100%)2 = 1.21

Of course the fans speed should never be increased beyond cataloged maximum safe speed!

It is important to note that the increased resistance will notobserved on the system. The pressure increase is only for

purpose of selecting the fan to compensate for the losassociated with the particular system effect.

The fan laws cannot be applied selectively, only simultaneou

According to the fan laws, if the fan speed is increased 10% fogiven system, the flow through the system will increase 10%, system resistance will increase 21%, and the fan BHP w

increase 33%. This represents an obvious waste of energy duean often avoidable system-related deficiency. In most cases, su

a change would require the purchase of a larger motor as well anew drive. If the fan is a direct-connected arrangement, limited

one fixed motor speed, the solution becomes even mexpensive. These considerations and horsepower penalties apply

all the major causes of system-induced performadeficiencies.

If the available space dictates the need for a turn into the

inlet, a standardized inlet-box design, with predictable losshould be used whenever possible.

DISCHARGE DUCTWORK

The connection made to a fan outlet can affect fan performan

An outlet duct ranging in length from 21/2 to 6 fan whdiameters, depending on velocity, is necessary to allow the fan

develop its full rated pressure. If the outlet duct is omitcompletely, a static pressure loss equal to one half the ou

velocity pressure will result. The system resistance calculatshould include this loss as additional required static pressure.

StaticPressure

BrakeHorsepower

CFM

Figure 7 - Static Pressure Curve for

Backwardly-Inclined Fan

Figure 8 Turning Vanes

Page 3


21/112

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 is

thrown to the outside of the scroll. Figure 9 shows a typicalvelocity profile.

In a duct with a uniform cross-section, the average velocity will

be 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, the

average velocity pressure at the fan outlet will be higher than theaverage downstream. Since total pressure will be virtually the

same, 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 should

follow the same direction as the wheel rotation. Turns made inthe opposite direction of wheel rotation (such as those shown in

Figure 10) can have a pressure drop beyond normal systemcalculations. Usually the drop is between .5 to 1.5 fan outlet

velocity pressures.

INLET OR OUTLET OBSTRUCTIONS

System obstructions can be as obvious as the cone-shaped stack

cap which can have a pressure drop as high as one velocitypressure, or as subtle as the installation of a large fan sheave

directly 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 a

plenum 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 could

result in additional losses beyond the normal system calculation.These losses will increase as the velocity increases or as the

distance between the obstruction and the fan inlet decreases.

CONCLUSION

AMCA Publication 201 - Fans and Systems, presents an in-depth

discussion of system effect and provides methods for estimatinglosses associated with many common situations.

If system effect situations cannot be avoided, their impactperformance should be estimated and added to the calcula

system resistance prior to sizing or selecting the fan. Ignor

the system effect could lead to difficult field performaproblems later. It could be that the installed fan does not hthe necessary speed reserve, or the motor is not of suffic

brake horsepower. The cost of correcting such a fiperformance problem could escalate rapidly.

System designers need to carefully consider the system eff

values presented in AMCA Publication 201. By accuratdefining the true performance requirements of fans in instal

systems, field performance problems can be reduced significantl

F o r m 6 0 7 G

Figure 10 - Poor Fan Outlet Connections

Figure 11 - Plenum System

Figure 9 - Velocity Profile at Fan Outlet


22/112

ENGINEERING LETTERThe New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521-55

I N C R E A S I N G F A N P E R F O R M A N C E

INTRODUCTION

Industrial processes and plant-ventilation systems often needmore air than originally designed. Increased productionrequirements, process changes, and facility renovations are a

few of the major reasons. Additionally, the lack of adequate

maintenance over time can negatively impact system airflows.This letter discusses several procedures that can increase airflow.

CHECK THE FANS MECHANICAL CONDITION

Often airflow can be increased by adhering to proper fan

maintenance 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 in

large fans, a sixteenth of an inch of build up can reduceperformance.

Check fan rotation. See Figure 2. Centrifugal fans will move

some 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 FigureComponents may be out of position due to routine cleaning

painting or the wheel could have shifted during shipment. backward inclined fans, the relation of wheel to inlet cone

very critical. Even a quarter of an inch can have a maimpact. The fans installation and maintenance literature sho

the proper positioning of the wheel to the inlet cone (dimension) or inlet plate.

INSPECT THE SYSTEM

The design and maintenance of the system plays a large role

achieving the overall desired performance. Visual inspectioften reveal some easily rectified problems that

significantly impair performance.

Check for clogged filters or coils. If the system has not beproperly maintained, clogged filters or obstructed coils w

reduce airflow. The greater the obstruction, the greater the lin airflow.

Eliminate System leaks. Any leaks in the ductwork wcontribute to reduced performance, especially leaks arou

plenum bulkheads that can lead to recirculation of air. Wflexible connectors are a common source of leaks and sho

be inspected regularl y.

Verify that dampers are installed correctly and operatproperly. If the damper linkage is out of adjustment, damper may not be opening completely, thereby reduc

performance. If inlet dampers are used, make sure they installed so that the air is pre-spun in the same direction

wheel rotation. See Figure 4. If the air distribution systemploys 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


23/112

For all dampers, make sure there is sufficient clearance for the

blades to open and close completely without hitting theductwork or other system components. Last, for systems with

either 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 through

the 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 in

Figure 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 speeding

up the fan. While airflow is increased by speeding up the fan, sotoo are static pressure, noise, and power requirements. Figure 6

presents this graphically. Therefore, while increasing the fansspeed is an easy procedure with low first cost, the additional

operating expense over time makes it the most costly solution.See Engineering Letter 2 - Fan Laws and System Curves, for

additional information.

When increasing fan speed, it is necessary to check maximum safe speed of the fan and make sure the moto

capable of the horsepower required to run the fan at the nspeed.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 most costly alternative. However, on a life-cycle-cost ba

considering operating and maintenance expense, it can be least expensive, as compared to increasing the speed of

existing f an.

Sometimes a second fan may be added, either in series

parallel with the original, although it may be more cost effectto simply upgrade the system with a new fan capable of

required airflow and pressure.

Adding another fan in series will increase the airflow because

the additional pressure. The operating point of the new systmoves further out/up the system curve. Where duct size

adequate to handle the desired amount of air but the existing doesnt provide sufficient pressure, a second fan in series m

be the best solution. However, make sure the ductwork handle the increase in pressure.

Adding another fan in parallel with the first will increase airfldue to the combined capacities. Because capacities are be

combined instead of pressures, a greater increase in airflow wresult for a given system. However, system pressures will a

increase and caution is required to avoid the unstable operatarea of the combined fan system.

CONCLUSION

When more air is required it is important to investigate

system on a step-by-step basis, considering the least expenpossibilities first. For existing systems that seem to have

performance, fan and system maintenance is the place to stOften, simply improving the efficiency of existing compone

will suffice. For systems that require greater airflow andpressure, increased fan speed is generally the first alternati

However, when large increases in performance are requirthere may be no alternative but to purchase a larger fan.

F o r m 6 0 7 G

Figure 4 Inlet Damper/Fan Wheel Rotation

Figure 5 Fan Inlet Connections

Figure 6 The effects on brake horsepower, static pressure an

loudness when fan speed is increased.


24/112

FIELD TESTING OF FAN SYSTEMS

INTRODUCTION

A fan system may require field testing when the system isthought to be malfunctioning, needs modification or requires

balancing 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 list

of 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 f ield 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 be70F. and 29.92 inches of mercury).

For greater convenience, a more compact Magnehepressure gauge may be used with a Pitot tube as a substitfor the manometer mentioned earlier. These gaugillustrated in Figure 4, are available in a variety of pressranges.

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

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

4. An accurate temperature probe is used to meas

temperature at each test location where volume or stapressure readings are taken.

Sometimes there are no accessible test duct locations suitafor use with the Pitot tube. In this case, the air volume candetermined at the system entrance or exit, or through a grillecoil by using an anemometer or velometer. This methhowever, is not as accurate and readings should only be takenexperienced service personnel familiar with this type of testi

PERFORMING A PITOT TUBE/MANOMETER TEST

1. Make a sketch of the system as a record and as a guideselecting locations for taking test readings. Often this w

call attention to poor system-design features. Includimensions, such as duct diameters or areas, duct lengmotor size, motor speed and sheave diameters on bdrive fans.

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

60527-55

7


25/112Page 2

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 test

points within each traverse is shown on the field test sheet

included 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 which

the 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 gases

or 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


26/112

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 testlocation (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 andcompute the average of the square roots. Then:

CFMinlet

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

(Avg. of Sum ofVPs) x ( Density CFM )DensityCFM test

Density Inlet

The above calculation gives air volume in actual cubic feetper 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 Density4. Determine the fan static pressure (SP) by the following

formula:

SPfan

= SPoutlet

- SPinlet

- VPinlet

Where: VP inlet =( CFM inlet )

2

xDensity

inlet1096 x inlet area in sq. ft.

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

SP standard= SP actual ( Actual Density )Standard Density5. Fan motor horsepower may be determined in several ways.

The best is to read the volts and amperes supplied to the

motor 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 3

746

Page 3

This method requires power factor and motor efficiency dwhich may be difficult to obtain.

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

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

b. Draw a dotted line through one-half no-load amps, a

zero HP, and nameplate amps at nameplate HP (points

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,

e. Determine running HP by plotting running amps.

Multiply fan horsepower by the K density correction fato determine HP at standard conditions.

6. Locate volume, static pressure and horsepower operformance curve drawn at the fan RPM. Curves cangenerated using manufacturers fan-selection softwarspecific densities, temperature and altitude.

The test plot values will probably not fall exactly on curve. If the fan system has been designed and instaproperly, the difference should be small, reflecting accuracy. If the difference is great, the system shouldanalyzed as described in the next section. Figure 8 showtypical fan curve and field test points which fall on the cur

Figure 7 Amperes versus Horsepower

Figure 8 Typical Fan Curve and Field Test Points


27/112

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 readings

ninety 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


28/112

ENGINEERING LETTER The New York Blower Company 7660 Quincy Street, Willowbrook, Illinois 60521-55

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 S

INTRODUCTIONIn general terms, a pressure blower provides relatively high

pressure at low volume when compared to other types ofcentrifugal fans. For purposes of this letter, fans with volumes

to 10,000 CFM with pressures to 80" WG are consideredpressure blowers. Typical applications require constant pressure

throughout the systems 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 Blowers research has resulted in a unique wheeldesign that is not truly radial. The blades are slightly canted

backward and dual tapered from the hub to the blade tip. SeeFigure 1. This design provides better efficiencies and, as a

result, significantly lower noise levels. The volume-pressurecharacteristics remain the same as radial-blade wheels.

POINT OF OPERATIONSince typical pressure-blower applications require a cons

pressure, selections are normally near the flat peak of the spressure curve. See Figure 2. Because of the flat nature of

pressure-blower curve, a typical question is, what keepsfans performance from fluctuating between different point

the fan curve? The answer lies in the relationship betweenfans performance curve and the system curve.

At a given RPM, the fan can only operate on its performa

curve. The only way to alter this curve is to either increasdecrease the fans speed. Conversely, the system can

operate along one system curve. The only way to changesystem curve is to increase or decrease the resistance thro

the system. Since the two curves can only intersect at point, the actual performance of the fan can occur only at

intersection of the fan curve and the system curve. Thdepicted 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 Curve


29/112Page 2

Considering that pressure blowers are often selected near the

peak of their pressure curve, dampering usually results in anoperation left of the pressure peak. One benefit of radial-blade

wheel design is that it delivers stable performance left of peak.

Radial wheels bring other advantages to pressure blowers. The

radial design delivers greater pressures at a specific RPM thanboth the radial-tip and backwardly-inclined designs. The

inherent strength of the radial wheel allows for the relativelyhigh wheel tip speeds required for the development of high

pressures. Remember,

nyb fan handbook 2007

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