B1. Fan Laws and Fan Control - Robinson

8/14/2019 B1. Fan Laws and Fan Control - Robinson

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Robinson Industries Fansfor You


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The Beginning - 1892


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Robinson IndustriesToday

Heavy Duty Fan Manufacturer

Global Support Capability w/ Intl. Representation

113 Years in Fan Industry(1892 - Present)

Four Locations:

4 Fabrication Facilities

1 Service Center

Corporate HQ

Zelienople, PA, USA

Abilene, TX


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Robinson ProductRange

Custom Built Special Purpose Centrifugal Fans

Industries Served: Power, Cement , Paper, Chemical,

Ceramic, Steel, Aluminum, Mining, Fertilizer

Combustion Air Blowers

Baghouse I.D. Fans

Material Handling Exhausters

Repair/Rebuild/Retrofit


5/74

Gas Density Calculation

Density = f (Temperature, Pressure, MolecularWeight)

One Calculation Method Is:D = 0.075 x (460+70)/(460+T) x {(407-k*FASL+PI) / 407} x

MW/28.96

D = Density, lb./cu.ft. (note: standard air = .075 lbs/ft3)T = Temp at fan inlet, F (note: standard temp = 70 F)

MW = Molecular Weight (note: dry air = 28.96)k = Altitude correction (note: =.013)PI= Fan inlet pressure, in-H2O gage( could be + or -)

(Ref. Page 1 of Robinsons Fan Performance and Design Manual)


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43 oF wb2

60 oF db

0.002 lb H2O/lb da3

1

Example: Tdry = 60 degF; Twet= 43 deg F; what is the

absolute humidity ratio? What is the density?

1. Enter the chart at the bottom at 60 oF dry bulb.

2. Find 43 oF wet bulb along the saturation line and read down

and to the right along 43

o

F until it intersects 60

o

F db.3. Plot this point, then read horizontally to the right to the

read absolute humidity ratio W which is in pounds of

water per pound of dry air (= .002 lb H2O /lb dry air)

4. Read specific volume = 13.14 ft3per lb of dry gas

5. Determine density = (1 + .002)/(13.14) = .076 lbs/ft3


9/74

Fan System Assessment Tool(FSAT)

This tool developed by the US

Department of Defense and the Air Movingand Control Association, International

helps evaluate fan and system efficiency.

It also includes a segment that is helpful indetermining gas density.


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System Resistance

Volume (CFM x 1000)

500450400350300250200150100500

StaticPressure(InH20)

60

50

40

30

20

40

0BHP

ROBINSON INDUSTRIES

FAN : 83" x 14.38" FRD

FOR : Basic Fan Laws I

FAN SPEED : 1180

TEMP. : 70

DENSITY : 0.075

SYSTEM RESISTANCE

150,000 CFM

40 SP

Every system has

resistance to flow.

Most gas systems

can be approximated

by an equivalent

orifice. The flow is

turbulent and the

pressure (P) requiredvaries as a function

of the square of the

volume flow-rate (Q).

P = kQ2 (k is a

system resistance

constant.)

(Ref. Page 16)


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Fan and System Curves

Volume (CFM) x 1000

300250200150100500

Pressure(InH20)

50

40

30

20

10

0

ROBINSON INDUSTRIES, INC.

FAN : 83" x 14.38" FRDFOR : Basic Fan Laws I

FAN SPEED : 1180

TEMP. : 70DENSITY : 0.075

SP

SYSTEM RESISTANCE

150,000 CFM

40 SP

The actual

operating

point will

be the

intersection

of the

system

resistance

curve and

the fancurve.

(Ref. page 17)


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Fan BHP Curve

Volume (CFM) x 1000

30025020015010050

0

BHP

2000

1600

1200

800

400

0

1,257 BHP

150000 CFM


FAN : 83" x 14.38" FRD


FAN SPEED : 1180

TEMP. : 70

DENSITY : 0.075

BHP

This shows thepower required

to drive the fan.For mostcentrifugal fansthe BHPincreases as

CFM increases.A Non-overloading fan

has a BHP curvethat peaks and

then drops off athigher flow.

(Ref. page 17)


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Fan CurvesConsiderations for a Good Operating Point

Volume (CFM) x 1000300250200150100500

Pressure(InH20)

50

40

30

20

10

BHP

2000

1600

1200

800

400

40 SP

1257 BHP

150000 CFM


FAN : 83" x 14.38" FRD


FAN SPEED : 1180

TEMP. : 70

DENSITY : 0.075

Efficiency%

100

80

40

20

0

SP

BHP

Eff.

Efficiency

Stability

Sound

Size


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Lab Test Arrangement


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Laboratory AirPerformance Test

Test Fan


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Laboratory AirPerformance Test

Test Fan

Orifice Plate (Throttling Device)


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Pitot Traverse in Outlet Duct


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Fan Laws

Static Pressure, Velocity Pressure

(Dynamic Pressure), Total Pressure

Density Change

Speed Change

Size Change


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Static, Velocity, and TotalPressure

Pt = Ps + Pv Consider two planes

1. fan inlet

2. fan discharge (after evase)

Fan Total Pressure: Ptdischarge- Ptinlet

Fan Static Pressure:

Psfan= Psdischarge- Ptinlet or,

Psfan = Psdischarge- (Psinlet + Pvinlet) Fan Static Pressure Rise (or, Differential Static

Pressure):

PStatic Rise= PsdischargePsinlet


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Density Change

VolumeCFM2= CFM1

Static Pressure

SP2= SP1x (Density2/Density1)Horsepower

HP2= HP1x (Density2/Density1)

Sound Power

Lw2= Lw1+ 20 log x (Density2/Density1)

(Ref. Page 25)


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Density Change Example

Problem:Find the new performance at 600F and

0.0375 lbs/ft3 given the performance at

70F and 0.075 lbs/ft3:

Given: 880 RPM

Flow = 125,000 cfmDensity = 0.075 lbs/ft3 @ 70F

SP = 20 in-H2O

BHP = 525 hp

SolutionCFM2=CFM1=125,000 cfm; (Constant Volume Machine)

SP2 = (20 in-H2O) x (0.0375/0.075) = 10 in-H2O

HP2 = (525 hp) x (0.0375/0.075) = 262 hp

(Ref. Page 25)


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Density Change Example Curve

Volume (CFM) x 1000

25020015010050

Pressure(InH20)

25

20

15

10

5

0

BHP

1000

800

600

400

200

10

20

262 BHP

525 BHP

125000CFM


FAN : 83" x 14.38" FRD


FAN SPEED : 880

TEMP. : See Below

SP

BHP

T1 = 70 F

D1 = 0.075

T2 = 600 F

D2 = 0.0375


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Speed Change(incompressible flow)

Volume

CFM2= (RPM2/ RPM1) x (CFM1)

Static Pressure

SP2= (RPM2/ RPM1)2x (SP1)

Horsepower

BHP2

= (RPM2/ RPM

1)3x (BHP

1)

Sound Power

Lw2= Lw1+ 50 log (RPM2/ RPM1)(Ref. Page 23)


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Speed Change Example

Question: Customer will change speed from 1180 rpm to 880

rpm. What will be the new performance?

Given:The existing fan has the following performance.

1180 rpm

100,000 acfm

20 in-H2O Static Pressure @ 70 degF

432 bhp

Solution:

acfm2= (100,000) x (880/1180)1= 74,576 acfm

Ps2 = (20.0) x (880/1180)2= 11.1 in-H2O

bhp2= (432) x (880/1180)3= 179 bhp

Note: The new performance is still on the original system resistance curve.The original efficiency is maintained.

(Ref. Page 23 of Robinson Fan Performance and Design Manual)


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Speed Change Example Curve

Volume (CFM)1601208040

Pressure

InH20

30

25

20

15

10

5

0

BHP

1200

1000

800

600

400

200

11.1 SP

20 SP

179 BHP

432 BHP

74576 CFM 100000 CFM


FAN : 64" x 12.94" FRD


FAN SPEED : 1180 & 880

TEMP. : 70

DENSITY : 0.075

SP

BHP

SYSTEM RESISTANCE


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Size Change Fan Law

VolumeCFM2= (Size2/ Size1)

3x (CFM1)

Static Pressure

SP2= (Size2/ Size1)2x (SP1)Horsepower

BHP2= (Size2/ Size1)5x (BHP1)

Sound PowerLw2= Lw1 + 70 log (Size2/ Size1)Note: All fan dimensions change in proportion to the wheel

diameter. Used to size fans from lab prototype tests.

(Ref. Page 24)


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Size Change Example

Fans with Geometric Similarity What is the performance of a 90 inch dia radial

bladed fan, based on a 40 inch dia geometricallysimilar test fan (at the same speed and density)?

Given: 40 inch diameter radial bladed fan

9,000 cfm20 SP @ 70F, 0.075 lb/ft339 BHP

Solution: for 90 inch diameter geometrically similarfan

CFM2 = (90/40)3x (9,000) = 102,516 CFMSP2 = (90/40)2x (20) = 101.3 H2OBHP2 = (90/40)5x (39) = 2249 HP

(Ref. Page 24)


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Repairs and tip-outsA worn out wheel can be not only be refurbished, but also tipped-

out to achieve a slightly higher volume and pressure capability.


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Size Change: Tip-Out or De-Tip

VolumeCFMMod= (Diamod/ Diaorig)

2x (CFMorig)

Static Pressure

SPMod= (Diamod/ Diaorig)2

x (SPorig)Horsepower

BHPMod= (Diamod/ Diaorig)4x (BHPorig)

Sound PowerLwmod> Lw + 70 log (Diamod/Diaorig)



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Tipped-out RB Wheel


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Tipped-out RT Wheel


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Tip-out and De-Tip Notes

Only the wheel diameterchanges.

Normal Tip-Out or De-Tip

is up to 5%Maximum is around 10%for some fan types

Increased Noise

Cannot De-Tip an AirfoilWheel

(Ref. Page 24)


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Tip-Out Example

Problem: Customer ABC has a 100 diameter wheel and they wouldlike to tip it out by 5%, (new wheel diameter is 105). What would bethe new performance?

Given: 100 inch diameter wheel

100,000 CFM

20 H2O @ 70F393 BHP

Solution:

CFMmod= (105/100)2x (100,000) = 110,250 cfm

SPmod= (105/100)2x (20) = 22.05 in. H2O

BHPmod= (105/100)4x (393) = 478 bhp



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Fan Laws Summary

Q2= (Q1)(D2/D1)3(N2/N1)1(1)(Kp2/Kp1)-1

P2= (P1)(D2/D1)2(N2/N1)2(2/1)1(Kp2/Kp1)-1

HP2= (HP1)(D2/D1)5(N2/N1)3(2/1)1(Kp2/Kp1)-1

Lw2= (Lw1)+70log(D2/D1)+50log(N2/N1)+20log(2/1)


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Fan Curves

See reference document titled:

Modified Fan Performance Curve Calculator


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Fan Control

Outlet Damper

Inlet Damper (various types)

Speed Adjustment

ConsiderationsInitial Cost

Energy Cost

Mechanical & Electrical Maintenance &Reliability


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Outlet Damper Control

Volume (CFM) x 1000

30025020015010050

Pressure(InH20)

50

40

30

20

10

0

BHP

2000

1600

1200

800

400

42.5 SP

1000 BHP

105000 CFM


FAN : 83" x 14.38" FRD


FAN SPEED : 1180

TEMP. : 70

DENSITY : 0.075

: .

: .

:

SP

BHP

Turndown can

be achieved bypartially closinga fan outletdamper. In thisexample, thepressure drop

across the outletdamper is 22.5in-H2O (= 42.5-20). Thesystemperformance is20 in-H2O at105,000 CFM.

(Ref. page 19)

Rated at

150,000 acfm

and 40 in-H2O

Pressure loss= 22.5 in-H2O

Inlet Louvered Damper Control


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Inlet Louvered Damper Control

The same

operatingpoint canbeachievedbythrottling

an inletlouvereddamper to20 degreesopen. TheBHP is

reduced to880 due topre-spin.(Ref. pages20-21)

Volume (CFM x 1000)

3002752502252001751501251007550250

50

45

40

35

30

25

20

15

10

5

0

BHP

4000

3600

3200

2800

2400

2000

1600

1200

800

400

0

Robinson Industries, Inc.Fan : 83" x 14.38" FRD SWSI

For : Basic Fan Laws I

Fan Speed : 1180 RPMTemperature : 70F

Dens ity : 0.075 LB/FT

Date : 02/02/1999Quote Number : 123456

RII FO # : 123456

EF :1SE :1TF :1

Note: Louvered damperedcurves are approximate

SP

20 4060

BHP20

4060

Note: InletP(s) Ratio = 0

880 BHP

105000 CFM

20 in. H20


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Louvered Damper

15% leakage (typical).

Parallel blade operation.

Lower efficiency when dampered.Good for dirty airstream.

Synchronized operation forDWDI.

(Ref: Page 68)


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Louvered Inlet Damper


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Radial Inlet Damper Control The same

operating

point can beachieved bythrottling aradial inletdamper to25 degreesopen. TheBHP isreducedfurther to700 due tothe moreeffectivepre-spinthan thelouveredinlet damper

(Ref. pages20-21)

Volume (CFM x 1000)

3002752502252001751501251007550250

50

45

40

35

30

25

20

15

10

5

0

BHP

4000

3600

3200

2800

2400

2000

1600

1200

800

400

0

Robinson Indus tries, Inc.

Fan : 83" x 14.38" FRD SWSIFor : Basic Fan Laws I

Fan Speed : 1180 RPM

Temperature : 70FDens ity : 0.075 LB/FT

Date : 02/03/1999

Quote Number : 123456RII FO # : 123456

EF :1SE :1TF :1

Note: Radial damperedcurves are approximate

SP

20

30 45

BHP

2030

45

Note: InletP(s) Ratio = 0

700 BHP105000 CFM

20 in. H20


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Radial Inlet Damper


43/74

Radial Inlet Damper

10-15% leakage (typical).

Good dampered efficiency

Center-support or cantileverdesign.

Good for open inlet SWSI.

Operating linkage outsideairstream.

(Ref: Page 68)


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Vortex Damper


High efficiency in damperedconditions.

Mechanism outside airstream.

Conical inlet pieces preferred.

Cannot be used without inletbox.

Split design available No need to extend bearing

centers for DWDI.

(Ref: Page 68)


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Variable Inlet Vane (VIV)


Built as part of the inletpiece.

Requires cone-shaped inlet

piece design. High efficiency when

dampered.

No need to extend bearingcenters on DWDI.

Mechanism in airstreamfor clean applications only.

(Ref: Page 68)


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Variable Inlet Vane (VIV) Damper


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Speed Control

Volume (CFM) x 100025020015010050

Pressure(InH

20)

25

20

15

10

5

0

BHP

1000

800

600

400

200

20 SP

450 BHP


FAN : 83" x 14.38" FRD


FAN SPEED : 840

TEMP. : 70

DENSITY : 0.075

SP

BHP

SYSTEM RESISTANCE

105,000 CFM

The sameoperating pointcan be achievedby reducing the

speed from 1180rpm to 840 rpm.The BHP isgreatly reducedto 450 sincepeak efficiency

is maintained.

(Ref. page 22)


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Annual Operating Cost

Outlet Damper: 1000 HP = 313,000 $USLouverd Damper: 880 HP = 275,000 $US

Radial or VIV Damper: 700 HP = 219,000 $US

Variable Speed Control: 450HP = 141,000 $US

Example Calculation (for variable speed control):

(450 HP) x (0.746 kW/HP) x (350 days/yr.) x (24 hrs/day) x (0.05 $/kWh) = $140,994/yr

Note: 1 HP costs about $313/yr


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System Effects


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System Effects

Fan Performancecan be affectedby the

configuration ofductworkupstream and

downstreamof the fan.

(Ref. Page 28)

System Effects Fan Outlet


51/74

System EffectsFan Outlet

(Ref. Page 28)Reprinted from AMCA Publication 201-90


52/74

Effective Duct Length


S t Eff t F t f


53/74

System Effect Factor forOutlet Elbows


System Effects (Pressure Losses)


54/74



CFD Analysis of Non Uniform


55/74

CFD Analysis of Non-UniformFlow in Fan Outlet Elbow

Forced Inlet Vortex


56/74

Forced Inlet Vortex -(Counter Rotating Swirl)


Decreased Fan

Aerodynamic Performance

Increased Fan Brake

Horsepower Requirement

Fan Aerodynamic Surge

/ Pulsation may Result

Non Uniform Flow Induced into


57/74

Non-Uniform Flow Induced intoFan Inlet by a Poorly Designed

Rectangular Inlet Duct


Most Common Cause of

Deficient Fan Performance

Fan Aerodynamic Surge

/ Pulsations may Result

Properly Designed Inlet

Box Provides a Predictable

Inlet Condition and

Maintains Stable Fan

Performance

N U if Fl i t F I l t


58/74

Non-Uniform Flow into Fan InletInduced by 90Round Section

Elbow - No Turning Vanes




59/74



S t Eff t


60/74

System Effects


Non Uniform Flow Induced into Fan


61/74

Non-Uniform Flow Induced into FanInletCFD Analysis of Flow


62/74

Fan Stability


63/74

The Mechanics of Surging

(Ref. Page 36)

Th M h i f S i


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(Ref. Page 36)

B

Desired Operating Point:

When operating volume

is reduced from A to B,

the fan responds bygenerating more pressure.

As a result the system

remains stable with the

flow proceeding from the

fan down the duct.

B


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(Ref. Page 37)

Undesirable Operating point:

When flow is reduced from

B to C, the fan responds by

generating less pressure. Then

the pressure in the downstream

duct is higher than the pressure

at the fan. The result is reverseFlow commonly called

surging .

C



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(Ref. Page 37)

We normally only thinkAbout fan performance

in the first quadrant

i.e. positive flow and

positive pressure.

Flow reversal!

C


68/74

Inlet Damper Control

(Ref. Page 35)

Di h D


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Discharge Damper

(Ref. Page 36)

Eliminates the Helmholtz

resonator effect of a largeduct or plenum,

Blow Off or Recirculation


70/74

Blow-Off or Recirculation

(Ref. Page 35)

Blow-off

RecirculationLoop


71/74

Robinson Surgeless Blower

(Ref. Page 38)

Robinson offers a

blower with special

surgeless characteristic

performance curve. The

peak pressure is at zero

volumetric flow.


72/74

Temperature Rise thru Fan

(Ref. Page 38)


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Sound Considerations

Ref. Page 73

C it N i C


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Community Noise Concerns

Documents

B1. Fan Laws and Fan Control - Robinson