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Subject to change without prior notice. The information inthis document contains general descriptions of the technicaloptions available, which may not apply in all cases. The required technical options should therefore be specified inthe contact.

Industrial FansDelivery Program . Centrifugal Fans . Axial Flow Impulse Fans . Sound Protection

www.tlt-turbo.comTLT-Turbo GmbH

TLT-Turbo GmbH

Business Location ZweibrückenHeadquartersPower, Mining and Tunnel Fans, ServiceGleiwitzstrasse 766482 Zweibrücken/GermanyPhone: +49 6332 808-0

Business Location Bad Hersfeld Manufacture and Logistics, Industrial Fans, ServiceWippershainer Strasse 5136251 Bad Hersfeld/GermanyPhone: +49 6621 7962-0

Business Location Frankenthal Service, Power FansHessheimer Strasse 267227 Frankenthal/GermanyPhone: +49 6233 77081-0

Business Location Oberhausen Service, Power FansHavensteinstrasse 4646045 Oberhausen/GermanyPhone: +49 208 8592-0

Sales Office Vienna, AustriaPower, Mining, Tunnel and Industrial Fans Sales, ServiceKarl-Waldbrunner-Platz 11210 Vienna/AustriaPhone: +43 1713 403010

Sales Office Moscow, RussiaPower, Mining, Tunnel and Industrial Fansul. Novoslabodskaya 31,d.4127055 Moscow/RussiaPhone: +7 459 6611780

Rep. Office Beijing, China TLT-Turbo GmbH Beijing Representative OfficeSales Industrial Fans22 D Building EMajestic Garden No. 6North Sichuan Medium RoadChaoyang District100029 Beijing/P.R. ChinaPhone: +86 10 82842683/84

Business Location Chengdu, China (Joint Venture)Chengdu KK&K Power Fan Co., Ltd.Sales, Service, Engineering, ManufactureNo. 15 Wukexisilu, Wuhou Disctrict, Chengdu 610045Sichuan Province/P.R. ChinaPhone: +86 28 85003500

Business Location Akron, OH USATLT-Turbo Inc.Power, Mining, Tunnel and Industrial FansSales, Service, Manufacture2693 Wingate AvenueAkron, OH 44314/USAPhone: +1 844-858-3267 (844-TLT-Fans)

Business Location TLT ACTOM (Pty) Ltd, South Africa(Joint Venture)Sales, Service, ManufactureMagnet House4 Branch RoadDriehoekGermiston, 1401Phone: +27 11 878-3050www.tlt-actom.co.za

Introduction 4

Field of Application 5

Product lines 7

Fan Designs 8

Control Modes and Characteristic Curves 9

Design and Fabrication 11

Fan Inquiry 23

Explanation of Common Fan Terms and Special Problems 25

Questions Regarding Fan Noise 29

Table of contents.

3

The requirements imposed on Industrial Fanshave noticeably increased over the years. The variety of problems that need to be tackledwhen handling gases requires a comprehensi-ve range of fans to optimize the selection foreach particular application.

Decades of intensive research and operatingexperience gained during this time are the basis for our fan range that provides the besteconomical choice for any application.

Guiding factors for thedevelopment of this range have been:

Low Investment CostLow Operating CostHigh ReliabilityLong LifeHigh Noise Attenuation

Fans from the range have been supplied to thefollowing industries:

Steam Generators and Power StationsCentrifugal and Axial Induced Draft Fans

Vapor FansPrimary Air FansDust Transporting FansBooster FansRecirculation FansHot and Cold Gas FansSecondary Air FansSealing Air Fans

Introduction. Field of Application.

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Cement IndustriesExhaust, Flue Gas and Forced Draft FansCooling Air FansPulverizer FansBy-Pass Fans

Mining IndustriesMine Fans for use above or below ground.Centrifugal and axial flow fans for all air quantities.

Steel and Metallurgical Industries Fans of all types for:

Sintering Systems (Sinter Plants)Pelletizing Systems (Pellet Plants)Direct Reduction SystemsDry and Wet Particulate Removal SystemsSoaking Pits and Walking BeamFurnacesEmergency Air SystemsIndirect Induced Draft Systems (Power Stacks)

Coke Oven PlantsCoke Gas Booster Fans, single and doublestage, made of welded steel plate.

Marine IndustriesForced Draft Fans.

Glass IndustriesCooling Air Fans for Glass TroughsCombustion Air and Exhaust Gas Fans

Centrifugal F. D. fan with inletvane control I and inlet silencer.

Mine ventilation fanVolume flow = 383 m3/sDepression ΔPSyst = 5400 PaSpeed n = 440 1/minSheft power Psh = 2560 kW

54

Chemical IndustriesRoasting Gas FansRecirculation FansCooling Air FansIntermediate Gas FansGas FansFans for Calcining and Drying ProcessesFans tor HCL Regeneration SystemsHigh Pressure Fan SystemsProcess Steam Fans

Product lines.

The fan range of TLT includes:

Single or multiple stage centri-fugal fans with maximum efficiency at pressures up to80,000 Pa. Standard andheavy duty designs are available.

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Fan Designs.

Double width double inletcentrifugal fans for high pressure and large flow volume.Standard and heavy dutydesigns are available.

Axial flow impulse fans withadjustable slotted flaps forhigh pressures at low tipspeeds.

Indirect l.D. fans (powerstacks) are often used for exhaust gases at temperaturesabove 5000C

The above diagram shows asummary of the operating ranges for the various fan designs.

Double width double inlet exhaust gasfan on an electro-melt furnace particulateremoval systemVolume flow V = 126 m3/sTemperature t = 120 °CPressure increase Δpt = 3820 PaSpeed n = 990 1/minShaft power PSh = 595 kW

Double width double inlet emergencyF. D. fan in a utility power plant.Volume flow V = 350 m3/sPressure increase Δpt = 9320 PaSpeed n = 990 1/minShaft power PSh = 4100 kW

Centrifugal Fans

Axial Flow Impulse Fans Inlet Vane Controls

(Action Type Axial Flow Fans) Support Structures Indirect l.D. Fan Systems

Silencers (Power Stacks)

Acoustic Insulation & Lagging

Sound Enclosures Emergency Air Systems º

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76

Fan efficiencies around 90%will reduce operating costs toa minimum level. However,not only are the fan eflicienciesat the maximum operatingpoint or design point of importance but also efficien-cies across the system opera-ting range have great signifi-cance.The most effective type of fancontrol is obtained by variati-on of the fan speed.Since speed control can onlybe achieved with high costdrive systems we commonlyutilize inlet vane control forboth centrifugal and axialflow fans.

Control Modes and Characteristic Curves.

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In the diagrams shown, the100% point (V = 100% and ΔPt = 100%)represents the optimum point.For various reasons the opti-mum point may not always beidentical with the designpoint.

Characteristic curves of a centrifugal fan with speed control Characteristic curves of a centrifugal fan with inlet vane control

Characteristic curves of an axial flow impulse fan with inlet vanecontrol.

Inlet vane control for a gas recirculation fan, largely gas-tight design; inlet diameter D = 2730 mm Ø

Model tests in the laboratoryhave provided the data nee-ded to determine specific per-formance characteristics of ourtans, in particular in view ofthe effects of different controlsystems. These performancecharacteristics enable theplanner to predetermine thespecific behavior of the fan ina system. Precise predictionscan be made regarding theoperation of fans operating assingle units or in parallel.If performance verification oflarge fans is required, testscan be conducted either inthe field or in some cases onour test stand.

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With few exceptions, the lar-ge variety of available fan types nearly always permitsdirect coupling of the fan tothe drive motor. We prefer thisarrangement because systemreliability is optimizied byavoiding interconnectingequipment such as gear bo-xes, belt drives, etc. The basicdesign flexibility of our fanspermitting alterations to or replacement of the impellerenables us to match actualoperating conditions if it isfound during operation thatthey differ from the conditionson which the original designdata are based.

Design and Fabrication.

Axial flow impulse fan with slotted flapadjustment, shown during production.

Volume flow V = 660 m3/sTemperature t = 156 °CPressure increase Δpt = 6520 PaSpeed n = 590 1/minShaft power PSh = 5480 kWDiameter D = 4220 mm ØInlet vane control: Diameter D = 4800 mm Ø

Double width double inlet I. D. fan for wasteheat boiler, fan support of laterally flexiblebase frame design.Volume flow V = 180 m3/sTemperature t = 245 °CPressure increase Δpt = 4420 PaSpeed n = 990 1/min

F.D. fan with inlet silencer in a steel mill.Volume flow V = 77 m3/sTemperature t = 30 °CPressure increase Δpt = 7260 PaSpeed n = 990 1/minShaft power PSh = 650 kWEfliciency η = 84 %Diameter D = 2400 mm Ø

Furthermore, slotted blade tipadjustment on centrifugal fanwheels or slotted flap adjust-ment on axial flow impulsefans are, in many cases, asimple means to meet specificoperating conditions.In the diagrams shown, the100% point (= 100% andΔPt = 100%) represents theoptimum point. For variousreasons the optimum pointmay not always be identicalwith the design point.

High gas temperature or particulate matter en-trained in the gas stream require specific atten-tion in fan selection and design. In such caseswe often recommend emphasizing increasedreliability in lieu of maximized efficiencies.

Axial flow impulse I.D. fan designed for verticalinstallation, shown in the manufacturing stage.

Impeller and shaft of an axial flow impulse fanduring balancing operation.

Axial flow impulse I.D. fan designed for verticalinstallation, shown in the manufacturing stage.

Impeller and shaft of an axial flow impulse fanduring balancing operation.

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Double width double inlet gas fan in electro-metallurgical plant.

Volume flow V = 195 m3/sTemperature t = 230 °CPressure increase Δpt = 3720 PaSpeed n = 740 1/minShaft power PSh = 875 kWMotor power PM = 1100 kW

Double width double inlet sintering gasfan in steel plant.

Volume flow V = 366 m3/sTemperature t = 160 °CPressure increase Δpt = 14200 PaSpeed n =990 1/minEfficiency η = 84 %Shaft power PSh = 5900 kW

Left: Impeller for a single stage F.D. fan Volume flow V = 4.8 m3/sTemperature t = 20 °CPressure increase Δpt = 31400 PaSpeed n = 2980 1/minDiameter D’ = 1250 mm ØImpeller mass mImp = 210 kg

Right: Rotor for a two-stage gas fanVolume flow V = 1.03 m3/sTemperature t = 100 °CPressure increase Δpt = 28500PaSpeed n = 2980 1/minDiameter D’ = 865 mm ØImpeller mass mImp = 140 kg(both impellers)

Rotor for a mine fanDesign volume flow VDes = 41 7 m3/s(Maximum volume flow Vmas = 500 m3/s)Temperature t =20 °CDepression ΔpSyst = 5890 PaSpeed n = 420 1/minImpeller diameter D’lmp = 5280 mmImpeller mass mlmp = 14000kgShaft mass mSh = 9900 kg

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Rotor for a double width double inlet sinteringgas fan.Shaft attachment: Centerplate of impeller isbolted between flanges of a divided shaft,centered on a very small diameter.

Volume flow V = 265 m3/sTemperature t = 160 °CPressure increase Δpt = 16650 PaSpeed n = 990 1/minShaft power PSh = 5230 kWImpeller mass mImp = 11000 kgShaft mass mSh = 11000 kg

Rotor of SWSl (single width single inlet) fluegas fan for a steel mill. Torque transfer: hub shaft with key. Erosion protection:

Bolted wear liners coated with wear resistantwelds.

Volume flow V = 39.5 m3/sTemperature t = 150 °CPressure increase Δpt = 13550 PaSpeed n = 1145 1/minDiameter D = 3030 mm ø

Right, background: Rotor of DWDI (doublewidth double inlet) flue gas fan for acement kiln. Torque transfer: integral hub withbody bound bolts.

Volume flow V = 125 m3/sTemperature t = 350 °CPressure increase Δpt = 6770 PaSpeed n = 990 1/minDiameter D = 3160 mm ø

Left hand side of the picture: Rotor coated withSaekaphen for a two-stage coke gas fan.

Volume flow V = 3.9 m3/sTempersture t = 25 °C Pressureincrease Δpt = 19650 PaSpeed n = 2970 1/mmDiameter D = 1224 mm øRight hand side of the picture: Rubberlined impeller for a flue gas fan behind aventuri scrubber.

Volume flow V = 17.6 m3/IsTemperature t = 72 °CPressure increase Δpt = 9810 PaSpeed n = 1485 1/minDiameter D = 1874 mm ø

Rotor for an axial flow impulse fan (I. D. fan fora utility power station)

Volume flow V = 660 m3/sTemperature t = 156 °CPressure increase Δpt = 6520 PaSpeed n = 590 1/minDiameter D = 4220 mm ØImpeller mass mlmp = 12100 kgShaft mass msh = 5200 kg(Hollow shaft)

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SWSI (single width single inlet) fan, supportedon both sides, during shop assembly

Rotor of IncoloyHousing and inlet box are lead coatedDual fixed bearing system with flexiblesupport structure

Volume flow V = 50 m3/sTemperature t = 30 °CPressure increase Δpt = 5870 PaSpeed n = 1000 1/minDiameter D = 2160 mm ø

Two-stage F. D. fan for a waste gas combustor,the fan system consisting of two fans arrangedin line with one common motor drive.To minimize spare part requirements the rotorsare identical in design (1st stage and 2nd stage).

Volume flow V = 5.1 m3/sTemperature t = 26 °CPressure increase Δpt = 53900 PaSpeed n = 2985 1/minShaft power PSh = 331 kw

Lead coated inlet box of thescrubber fan.

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SWSI (single width single inlet) gas recirculationfan, supported on both sides, installed in a utility power plant.

Volume flow V = 167 m3/sTemperature t =350 °CPressure increase Δpt = 2360 PaSpeed n = 720 1/minEfficiency η = 85,5 %Shaft power PSh = 457 kW

SWSI (single width single inlet) gas recirculati-on fan, supported on both sides, installed inan utility power plant supported by integralbase with vibration isolators.

Volume flow V = 142 m3/sTemperature t = 361 °CPressure increase Δpt = 7850 PaSpeed n = 990 1/minShaft power PSh = 1360 kWDiameter D = 3280 mm ø

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DWDI (double width double inlet) gas recircu-lation tan with concrete tilled integral baseframe and vibration isolators.

Volume flow V = 124 m3/sTemperature t = 300 °CPressure increase Δpt = 9615 PaSpeed n = 1490 1/minShaft power PSh = 1410 kWDiameter D = 2320 mm ø

DWDI (double width double inlet) gas recircu-lation tan during shop assembly.

Volume flow V = 250 m3/sTemperature t = 340 °CPressure increase Δpt = 3470 PaSpeed n = 715 1/minShaft power PSh = 1100 kWMotor power PM = 1300 kWDiameter D` = 3200 mm ø

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SWSI (single width single inlet) raw millfan for the cement industry, supported on oneside (AMCA arrangement 8).

Volume flow V = 114 m3/sTemperature t = 90 °CPressureincreaso Δpt = 5700 PaSpeed n = 745 1/minImpeller diameter DImp = 3350 mm ø

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SWSI (single width single inlet) cement kiln exhaust gas fan, supported on one side(AMCA arrangement 8), fan support of laterallyflexible base frame design.

Volume flow V = 133 m3/sTemperature t = 100 °CPressure increase Δpt = 4600 PaSpeed n = 745 1/min

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Our economical production facilities are equipped withmodern machinery. For example, a numerically control-led flame cutting machine and a metal spinning machi-ne are used for the processing of sheet metal.

NC flame cutting machine, with punch tapes automati-cally produced via electronic data processor.

Balancing machine for fan rotors up to 30000 kg and5000 mm Ø

Metal spinning machine for radii of inlet nozzles, impel-ler side plates and spinning flanges.

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SWSI (single width single inlet) forced draftfan, supported on one side (AMCA arrange-ment 8), with inlet vane control, arranged onan integral supporting structure with vibrationisolators.

SWSI (single width single inlet) sintering gasfan, rotor supported on both sides, with fluiddrive, fan supported by an integral steel frame.

Two-stage coke gas fan with oil lubricated roller bearings sealed with carbon packingglands. Speed: 2950 1/min.

Axial flow impulse fan (induced draft fan) withinlet vane control and possibility of adjustingthe slotted flaps individually during down-time.

DWDI (double width double inlet) fan arran-ged on an integral supporting structure with vibration isolators.

A selection of varios designs from our fan range is shown below.

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As a fan manufacturer TLT works closely with the architect I engineer and/orthe final user to optimize fan selection for each specific application.

The following set of conditions at tan inlet tobe supplied by the customer provides the basisfor fan selection and design:

Fan Inquiry.

Types of Fan Design and Installation

Radial-flow fans are normally of the single-inlettype up to approx. 100 m3/s; in some casesthe double-inlet type is used from 60-70 m3/salready. The actual limit between single- anddouble-inlet type is mainly determined by therelevant case of need, the suitable fan type,the required ratio volume/specific energy andthe speed.

We built single-stage radial-flow fans for aspecific energy of more than 40000 J/kg. It depends on the case of need, temperature,volume/pressure ratio and the possible speedfrom what pressure increase the fan has to beof the doublestage type.

The most favourable installation of the fans isthat on an elevated concrete substructure. Thisresults in shont bearing pedestals and the motor can be placed on a low frame or evendirectly on plates which are embedded in theconcrete. This simple, rugged kind of installationis less susceptible to vibration and thereforebest suited to sustain high imbalance forcesdue to wear or dust caking.

1. Identification of the plant and the process in the system for which the fan is to be used:

2. Volume flow: V or V SC **) in m3/s

3. Temperature: t in °C

4. System-related pressure increase: ΔpSyst in Pa ΔpSyst = Pt4 - Pt1*)

(Pressure distribution: Fan suction side/discharge side)

5. Fan inlet pressure measured against barometric pressure (+/-) P1 in Pa

6. Elevation: h in m above see level

7. Mains frequency (Standard frequency): fMains in Hz

8. Permissible noise level:Sound pressure level at a defined distance: Lallow in dB (A)or sound power level: LW allow in dB (A)

9. Information on the fluid handled:9.1 Type of gas:9.2 Density of gas: ρ or ρ SC **) in kg/m3

(if necessary supply gas analysis with moisture content)9.3 Particulate content of the gas: St or StSC **) in g/m3; mg/m3

9.4 Dust characteristics:S: Probability of build-upV: Probability of erosion9.5 Corrosion:K: Probability of corrosion due to

10. Type of preferred fan (see following sketches and explanation)

11. Additional information:

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Notes:

*) See also section"Pressure Definitions". "The system-related pressure increase ΔpSyst", asdefined by us, is often referred to as"static pressure increase Δpstat“, the dynamic pressurecomponents having been ignored, however.

**) The index "sc" identifies the standard condition (t = 0°C, p =101325 Pa).

p

If the substructure has to be made of steel cor-responding plate cross sections have to beused in case of larqe base-to-centre heights toachieve a sufficient vibration resistance. Thusthe fan weight and the costs are increased accordingly.

Fans which have to be installed on substructuressusceptible to vibrations, as e.g. in the structureor on a building roof have to be vibration-isolated.

For this purpose, the compact type of construc-tion is suitable which is a frameless, selfsuppor-ting structure utilizing the rigidity of the almosttotally enclosed suction boxes and housing.The vibration isolators are installed under the"hard" points such as housing walls.

The installation of the fan on an elevated con-crete base resting on vibration isolators is recommended if higher imbalances are expec-ted.

The vibration isolators reduce the amplitudes ofthe dynamic forces (alternating loads from theimbalance rotating with the fan rotor). The so-called isolation efficiency depends on thedistance of the exciter frequency (=fan speed)and the natural frequency of the spring masssystem of vibration isolator - machinemass.Normally, theisolationefficiency is above 90%.For speed-controlled fans it is therefore absolu-tely necessary to indicate the lowest required(reasonable) speed, as then accordingly "soft"springs have to be used.

Elevated concrete base without frame below the motor -rugged, simple installation.

Elevated concrete base with low frame.

Vibration-isolated installation with elevated concrete blockon vibration dampers.

Compact type of construction, selfsupporting structure, placed directly on vibration dampers.

Some installation examples:

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In cases where no specific data relative to gasvelocities, desired duct cross sections or specialinstallation requirements such as for mine fansare given we will determine the fan based onthe assumption that the cross sections A1 andA4 are equal and the total pressure increasebetween these terminal points represents the system-related pressure loss ΔpSyst stated by thecustomer.

A1 = A4

ΔpSyst = pt4 - pt1

With A1=A4 and disregarding compressibilityit follows that Pd4 = Pd1 and therefore:

ΔpSyst = (ps4 - pd4)-(ps1 - pd1) ΔpS

The dynamic pressure Pd is understood to bebased on the average gas velocity in a crosssection.

Example: Dynamic pressure at terminal point 4:

Pressure losses pV caused by fan componentsbetween the cross sections A1 and A4, for example inlet box, louver, inlet vane control, diffuser will be considered by us when sizingthe fan.

Explanations.Explanation of Common Fan Terms and Special Problems.

The following definitions of fan terminologymay facilitate communication between the fanmanufacturer and customer.

p PressureΔp Pressure Difference or IncreasepV Pressure Loss, ResistanceV Volume Flow at Inlet ConditionA Cross Sectional AreaI Lengthm Mass Flowc Gas Velocityρ Densityf Compressibility FactorY Specific Delivery WorkPSh Shaft Powerη EfficiencyT Absolute Temperature (Kelvin)χ Adiabatic Exponent

Indexes:

t totals staticd dynamic1, 2, 3, 4, markings of the cross sections

concerned (terminal points)

Pressure Definitions

The energy transmitted through the fan impeller to the volumeflow is needed to overcome the system resistances.These resistances can comprise the following:

Friction lossesBack pressure from pressurized systemsVelocity changes at system inlet and outlet and within the systemDraft forces due to density differentalsGeodetic head differences which are mostly negligible.

The sum of the above mentioned resistances as far as they occur in the system concerned, represent the total system resi-stance. According to Bernoulli this total resistance is to be un-derstood as total pressure. This pressure pt (analogy: total ener-gy) comprises static pressure ps (analogy:potential energy) andthe dynamic pressure pd (analogy: kinetic energy).

pt = ps + pd

(This definition is in accordance with VDMA-standard 24161)

To move the design volume flow the fan must generate - withinthe specified fan terminal points - a pressure increase equivalentto the total resistance of the system. This equivalent pressure in-crease is defined by us as system-related pressure increaseΔpSyst.

ΔpSyst = pt4 - pt1

If the cross sections 1 and 4 (Figure D- 1) represent the terminalpoints of the fan the system-related pressure increase ΔpSyst willthen be the difference of the total pressures at the terminal points1 and 4. This system-related pressure increase which is in factthe pressure increase required by the customer is often incorrectlystill refered to as static pressure increase Δpstat unfortunatelyignoring the dynamic pressure difference existing in most cases.The probable cause of disregarding this dynamic pressure in-crease lies in the generally used method of measuring the staticpressures in ducts by means of holes drilled in the duct wallsperpendicular to the direction of flow. In such cases the dyna-mic pressure differential at the terminal points has to be addedto the result of the static pressure measurement to obtain the correct system-related pressure increase.

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The design pressure Δpt, the parameter determi-ning fan size, is the sum of the system-relatedpressure increase and the pressure losses of thefan components.

Δpt = ΔpSyst + pv = p13 - p12

Our characteristic curves show the design pressure Δpt, as this pressure differential repre-sents a parameter defined by tests for a specificfan type at given operating conditions, where-as the system-related pressure increase ΔpSyst

varies with the losses pV which arise depending on the fan components used.

In determining the fan design pressure Δpt

other losses are considered in addition to theabove mentioned pressure losses pV if specialdesign conditions are specified involving inletand outlet pressure losses, e. g. pressure lossesin the case of silencers and turning bends, outlet pressure losses in the case of mine fansetc.

Fan Power and Efficiency

The fan design pressure Δpt of our fans is equalto the total pressure increase between the crosssections A2 and A3.With this design pressure, and the design volume flow V at fan inlet conditions, thepower requirements at the fan shaft PSh and theefficiency η will be determined, optimum inletconditions being a prerequisite.

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p

kpd = . c2

2

k4pd4 = . c42 = ( ) 2 . f

22k1

2VA4

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ºpSh = = V.Δpt . f m.Yη η

V in m3/sΔpt in Pa = N/m2

f < 1η < 1m in kg/sY in J/kg = Nm/kgPSh in W = NM/s

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Meaning of Symbols

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Examples.

Influences of Temperature and Density

A noticeable temperature rise will occur acrossfans with high pressure increase, in particularwhen the fan operates in a throttled conditionand at low efficiency. The adiabatic temperature increase ∅tad is:

Δtad = T2 . [(p3/ps) -1]

The real temperature increase is:

Δt =

Operating a fan at a temperature significantlybelow design temperature will cause the shaftpower requirement to increase as a function ofdensity or as a ratio of the absolute temperatures.Should a different gas with higher densitybe handled the shaft power requirement will increase with the ratios of the densities.

Since such operating conditions often occur atstart-up the louvers or inlet vanes need to beclosed in these cases.The pressure increase of the fan will also riseas a function of lower temperatures or highergas densities. This must be considered whendesigning flues and ducts, expansion joints,etc.

Examples of Various Fan Arrangements in a System with Corresponding Pressure Distribution Figures D-2 to D-4

x-1

x

≈ΔtAD Δptη 1250 . η

=PSh operationTdesign

Tcold

. PSh design

=kalternativ gas

kdesign. PSh design

=Δpt operationTdesign

Tcold

. Δpt design

=kalternativ gas

kdesign. Δpt design

Influence of Mass and Mass Inertia Erosion,corrosion and system-related contaminationcausing build-up can posibly lead to imbalan-ces due to uneven mass distribution. For suchcases impellers with large masses are advan-tageous because the shifting of the gravity center caused by the imbalance is smaller.

For the start-up time of a fan the determiningfactor is the inertia of the rotating mass. This start-up time is of importance relative totemperature increases of the electrical drive system causing limitations of the number of system startups.

Some installation examples:

Figure D-1: Marking of the Cross Sections and Reference Planes.

Figure D-2: Open Inlet Fan (e.g. Forced Draft Fan)

Figure D-3: Fan Connected to Ducts at Fan Inlet and Disscharge (e.g. Induced draft Fan)

Figure D-4: Exhaust Fan (e.g. Mine Fan)

Δpt Total Pressure Increase Betwee Section 2 and 3

ΔpSyst System Resistance-Related TotalPressure Increase

ΔpSyst.s System Resistance-Related StaticPressure Increase

pd Dynamic Pressure (Velocity Pressure)Difference

Δpd Dynamic Pressure Difference (Velocity Pressure

pv Pressure (Loss(es)pv in Pressure Loss at Fan InletpvBox Inlet Box PressurepvDif Diffuser Pressure LosspvOut Outlet Pressure LossA Cross Section

Indexes:

1, 2, 3, 4, Marking of Cross Sections Concerned (Terminal Points)

Behavior of

Total PressureStatic PressureDynamic Pressure (Velocity Pressure)

Δpt = ΔpSyst + ∑pv

ΔpSyst = ΔpSyst.s + Δpd

Meaning of Symbols used in Equations:

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Questions Regarding Fan Noise.

1. First Fundamentals

With progressing industrialization man is facedwith increasing environmental problems. Noiseemitted by fans belongs in this category. Thefollowing will give guidance in the problemarea of noise emitted by fans as well as theflow in the connecting flues and ducts.

The sound [“Schall“]1) perceived by the humanear is the result of oscillation of particles of anelastic medium in the frequency range of about16 to 16,000 hertz (Hz). One hertz is one oscillation per second. Depending upon themedium in which the sound travels we distinguish air sound, body sound, and watersound [“Luftschall, Körperschall, Wasser-schall“].

A pitched tone [“Ton“] is defined as sound oscillating as a sinusoidal function (compressi-on and depression). With increasing amplitudesound will be perceived as being louder andwith increasing frequency it will be perceivedas being higher. The tone in Figure 2 (soundpressure P2) is perceived as higher and gene-rally louder than the tone in Figure 1 (soundpressure P1).

For additional details see paragraph 2. Aclang [“Klang“] is created by the harmo-nic interaction of several tones.

Noise [“Geräusch“, “Rauschen“] is defined asstatistical sound pressure distribution across theperceivable frequency range. A noise annoy-ing the human ear is called an “excessive noise“ [“Lärm“].

Discharge silencer for two centrifugal forceddraft fans, designed as absorption type silencer, level reduction by 15 dB. Fan data:

Volume flow V = 2 x 62 m3/sTemperature t = 50 °CPressure Increase Δpt = 8120 PaSpeed n = 1490 1/min

level = lgof soundparameter

in B

effective valueof sound parameterreference valueof sound parameter

Double width double inlet centrifugal forceddraft fan with disc silencers and cover for noisetreatment of the fan inlet noise (shown duringshop assembly).

2. Human Noise Perception

Sound pressure is exactly measurable with instruments. The physiological effect on humansis much more difficult to determine. The humanear, for example, will perceive two tones ofequal sound pressure yet different frequency asunequally loud.

Numerous tests were made on listeners in orderto compare the loudness of tones at differentfrequencies and different sound pressures withthose of a 1000 Hz tone. In particular the objective was to identify the sound pressure px(measured in dB)1)at a frequency of 1000 Hzat which the sound pressure pn (measured indB) and the frequency fm (measured in Hz)would evoke the same perception in the listenerwith respect to loudness. As a result of thesetests, curves of constant loudness (stated inphone) were identified over the frequency ran-ge. By definition sound pressure level and loud-ness coincide in terms of figures at 1000 Hz.Graphs in Figure 3 show these curves of equalloudness.

Because the shape of the curves changes withfrequency as well as sound pressure one wasfaced with the problem of designing a handymeasuring instrument for an objective measu-ring of the loudness of sound. This was the impetus behind the search for a different eva-luation system. An additional reason lies in thefact that the phon curves can only be used toevaluate single tones. There is, however, a difference between the human ear's perceptionof single tones and its perception of noise.

1)The sound pressure is often referred to assound pressure level, see specifics underparagraph 3 "Fundamentals of Acoustics".

2)The reason for the special nuisance crea-ted by a single tone is its information con-tent (Example: Tones produced by sirens,warning and mating calls in the animalworld).

The solution, which takes these factors into account and which has been internationally accepted, is found in the socalled “A“ soundevaluation curve. The curve represents an approximation of the phon curve in midrangeof the sound pressure level. To give considerati-on to the fact that single tones are perceivedas being more annoying2) than broad bandnoise, a higher reduction is imposed on singletones in addition to the total noise level require-ments, for example such a typical single tone isthe “blade passing tone“ [“Schaufelton“] ofafan whose frequency is calculated with thenumber of blades and the fan speed expressedin Hz. This blade passing tone and its integermultiples (harmonics) form the so-called “bladepassing frequencies“ [“DrehkIang“].

3.Fundamentals of Acoustics(Definitions)

Units of Sound Parameters

In acoustics it is common to work with levels,i.e. it is common not to use the original para-meters with their corresponding units, but Ioga-rithmical parameter ratios using the logarithmto the base 10, the corresponding units beingbe (B) or decibel (dB).

º

level = 10 lgof soundparameter

in dBeffective valuereference value

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Discharge silencer designed as a resonantsilencer (λ/4-silencer or interference silencer)for two induced draft fans in a power plant(volume flow V = 2 x 660 m3/s, pressure in-crease Δpt = 6520 Pa), insertion loss = 33 dBat the frequencies of 118/236 Hz (blade pas-sing frequency and first harmonic).

Baffles (shown at center right) and baffle walls(shown below) of the above silencer installati-on after approximately 11,000 operatinghours.

As can be seen in Figure 4, the numerical values for LA are significantly below the L valu-es at low frequencies and have a much smallerimpact at higher frequencies.

Measuring Surface Sound PressureLevel L and LA

[“Meßflächen-Schalldruckpegel“ L und LA]

The-measuring surface sound pressure level L (= the sound pressure level at the envelopingmeasuring surface) is defined as the energeticmean1) of multiple sound level measurementsover the measuring surface S with eliminationofextraneous noise and room effects (reflections).

LA is the “A“ evaluated measuring surfacesound pressure level.The measuring surface S is an assumed areaencompassing the sound source at a defineddistance (mostly one meter). This envelopingsurface comprises simplified surfaces such asspherical, cylindrical and square surfaces generally following the shape of the sound producing equipment.

As the same units are applied to all sound parameter levels it is important to properlyidentify the type of the sound parameter levelreferred to, that means to distinguish, for exam-ple, between sound pressure level and soundpower level.

Sound Pressure Level L[“Schalldruckpegel“ L]

The sound pressure level L (most commonly alsocalled sound level) quantifies the sound pressu-re measured at a specific point.

By definition:

with p = effective value of sound pres sure atmeasuring point in Nim2

p0 = 2x10-5 N/m2

= 20 μ Pa

= 2 · 10-4 μ bar

(reference sound pressure, the audible thresholdfor 1000 Hz pitch)

Evaluated Sound Pressure Level LA

(= Sound Pressure Level Evaluated AccordingtoEvaluation Curve“A“) [“Bewerteter Schalldruck-pegel“ LA]. The evaluated sound pressure levelLA - expressed in dB (A) - is obtained by adding at the various frequencies a ΔL from theevaluation curve “A“ (see Figure 4) to the mea-sured sound pressure level L at the correspon-ding frequencies. The evaluation curve and theevaluation procedure are defined in DIN standard 45633, sheet 1.

L=10 lg = 20 lg in dBp2

p02

pp0

Baffles of the absorptive discharge silencer ofa forced draft fan (after approximately 11,000operating hours);

Close up photograph of the baffle wall~afterapproximately 11,000 operatinc hours.

Three-stage absorptive silencer for ambient airinlet to a forced draft fan Volume flow

V = 433 m3/s, pressure increaseΔpt = 8250 Pa) in a power station.

Design point V = 60%Attenuation to sound pressure level 70dB (A).

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_ _

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The individual center frequencies of the octave band are at:31.5 Hz63 Hz125 Hz250 Hz500 Hz

Sound Intensity Level LI

[“Schall-lntensitatsspegel“ LI]

At this point mention should be made of the so-called sound intensity I [“Schall-Intensitat“ I]which is the sound power relative to the reference area of 1m2.

I = in

With this definition an analogy to electrical-technology can be made:The sound intensity is proportional to the squareof the sound pressure. The definition of the corresponding sound intensity level LI is as follows:

LI = 10 lg in dB

with I0 = 10-12 watts/m2

(reference sound intensity)

4. Sound Analysis

The “total sound level“ or “sum sound level“ ofnoise is derived from the logarithmic additionof a multiple of single sound levels at differentfrequencies (Figure 5). In order to perform noise measurements, the audible frequency range has been divided into 10 octave bands.

The width of the octave is identified such thatthe ratio of the upper limiting frequency of thespectrum f0 to the lower limiting frequency fU is2:1.

Octave: = 2

The corresponding ratio for the „terz“ is:

„Terz“: = 3√2

Three „terz“ together make up an octave:Center frequencies are determined by:

fm = √

Octave: fm = √2 . fu =

„Terz“: fm = 6√2 . fu =

Sound Power Level LW

[“Schall-Leistungspegel“ Lw]

The value of the total sound power radiatingfrom a sound source is given by the soundpower level LW.

W = gas-borne acoustical power emitted as air sound in watts

W0 = 10-12 watts (reference sound power at audible threshold at 1000Hz)

Evaluated Sound Power Level LWA

[“Bewerteter Schall-Leistungspegel“ LWA] When an evaluation, similar to the one descri-bed in the example, of the sound pressure levelis onducted, using the evaluation curve “A“, the evaluated sound power level LWA will beobtained from the sound power level LW.

Relationship Between Sound Pressure Leveland Sound Power Level

The sound power W is not measured directlybut is calculated using the measured soundpressure p, sound particle velocity ν (molecularmovement velocity), [“Schallschnelle“ ν] andthe measuring surface S:

using

is follows that

Assuming that

the proportional relationship obtained is:

1) To calculate the energetic mean of all soundlevel measurements taken over the envelopesurface (taking into account the time interval oftesting) the following formula is to be used:

If the difference between the individual soundlevels is smaller than 6 dB the formula belowcan be used as an approximation representingthe arithmetical mean:

Any components that protrude beyond the surface but contribute little to the emission canbe neglected. Sound reflecting boundaries,such as floors and walls, are not incorporatedwithin the measuring surface. The measurementpoints shall be sufficient in number and evenlydistributed over the enveloping surface. Thenumber depends on the size of the sound source and the uniformity of the sound field.

Because of the logarithmical parameter ratiosused in acoustics the measuring surface in m2,will be related to a reference area to definethe measuring surface level LS [“Meßflächen-maß“ LS] as the characteristic parameter:

In terms of expressing theabove equation in acoustic level parameters the followingimportant equations can beobtained:

The sound power level LW canbe approximated by the sumof the measuring surfacesound pressure level L and themeasuring surface level LS.From the above relationship itcan be deduced that with agiven sound power level aspherical or a semisphericalsound dispersion (ideal sounddispersion) the sound pressurelevel will diminish by 6 dB forevery doubling of the distancefrom the sound source.

Through absorption of thesound in the air and on theground this value will increaseand through reflection of thesound by obstructions it willbe reduced. Furthermore, we-ather conditions can causeeither an increase or decreaseof the sound pressure level reduction.

L=10 lg ( . 10 lg )1n

∑i=n

i=1

0.1L-

Ls = 10 lg in dBS

S0

L≈ .1n

∑i=n

i=1Li-

- - - -

_

S = Measuring surface in m2

S0 = 1 m2 (reference area)

LW = 10 lg in dBWW0

W = p . v . S

v =pk .c

W = . Sp2

k .c

k = air densityc = air sound velocity

= constantc = constant

W ≈ p2 . S

k

LW ≈ L+10 lg = L + Ls in dBS

S0 WS

wattsm2

II0

-

-

- - -LWA ≈ LA+10 lg =LA+ Ls in dBS

S0

f0fU

WS

f0fU

fU . f0f0√2

f06√2

1,000 Hz2,000 Hz4,000 Hz8,000 Hz16,000 Hz

W = p . v . S

= I

I = p . v

I = = v2 . k . c

I ≈ p2

N = OutputU = VoltageI = AperageR = Resistance

N = U . 1

I =

N = = I2 . R

N ≈ U2

v =pk .cp2

k .c

Acoustics

The sound intensity is proportianal to the square of the sound pressure.

Electrotechnics

UR

U2

R

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In practical application, the first and last octavebands mostly play a secondary role. Commercially available sound measurement instruments to measure sound levels in dB anddB (A) are equipped with adjustable octaveand “terz“ filters to conduct frequency analyses.If the octave band analysis proves inadequatethe more selective “terz“ analysis should be employed, the octave band width being devi-ded into 3 “terz“ band widths.In the case of single tones or noises extendingover one “terz“ band only, the “terz“ band andthe octave band analyses will give the same figures.

The example in Figure 5 shows an octaveband and “terz“ band analysis. For a more selective analysis of a noise spectrum, narrowerband filters can be employed to further dividethe noise spectrum (search tone analyzer).

5.Addition of Levels

To determine the total level LtOt, partial levels L(sound pressure level or sound power level) willbe added in accordance with the followingequation:

When adding sound pressure levels it must beconsidered that all individual sound pressure le-vels have one common reference point. In thespecific case where n sound sources haveequal sound power W1, the total level Lw10can be determined by the following:

For a number of sound sources the level increa-se can also be determined using the diagramin Figure 6.In the special case of two single sound sourceswith different levels, the total level is obtainedby adding the difference of the individual le-vels to the higher level.

With reference to the diagram in Figure 7 itcan be seen that for a level difference of morethan 10 dB practically no level increase will result. For the special case of two sound sourceswith equal levels (level difference 0) a level in-crease of 3 dB will result (see also Figure 6).The case where two superimposed single tonesof equal sound pressure p1, equal frequencyf1, and equal phase n are to be added requires special consideration. Deviating fromthe above described summation a total level 6 dB higher than the sound pressure level ofthe single tone will be obtained:

If a difference in phase of 1 80 degrees exists(n1 = 00; n2 = 180°) or λ/2 interference re-sults, the tones eliminate each other (see Figure8).

Ltot=10 lg 10

Ltot= LW1 + 10 lg n

∑i=n

i=1

0.1Li

Ltot = 10 lg (2. )2 = 20 lg 2 .

= L1 + 20 lg 2

p1

p0

p1

p0

These two occurances are ofpractical importance in thecase where two fans are ope-rating at almost equal speedin a common duct system. Inthis case sound wave super-position results in periodicsound level variations calledbeats. The beat frequency isdetermined by the differencein the operating speeds of thetwo fans.

6. Noise Developement in Fans

The operating noise of a fancomprises various sound com-ponents. In boundary zonesof confined fast moving gasflow, eddy currents occur asthe result of the influence ofthe viscosity of the gas. Onfans these eddy currents occurat the blade discharge edges.The resulting noise caused bythe rotating impeller is referredto as “eddy current noise“ andis considered the “primary noi-se“. Superimposed on this noi-se is the “self-noise of highlyturbulent flow“ in the fan hou-sing and ducts. “Eddy currentnoise“ and “selfnoise“ [“Wir-belgeräusch und Strömungs-rauschen“] display a broadband frequency spectrum. Thesound power increasing ap-proximately with the 5th to 7thpower of the impeller tipspeed. In addition to broad

band noise, “pulsation noise“ [“Pulsationsge-räusch“] occurs at different frequencies causedby periodical pressure Oscillations of the medium due to the relative movement betweenthe impeller and a stationary fixture exposed tothe flow. “Pulsation noise“ will occur when theflow in the closed environment of the impeller isdisturbed by obstructions with protruding edges(cut off in centrifugal fans and stationary guidevanes in axial flow fans). For fans such distur-bing noise is also referred to as “blade pas-sing tone“ [“Schaufelton“] or “blade passingfrequencies“ [“Drehklang“], where the main disturbing frequency (base frequency) is theproduct of blade number times revolutions persecond. Integer multiples of the base frequencycan also occur as harmonics. The occurrenceof the “blade passing frequencies“ (base fre-quency + harmonics) can, depending on thetype and intensity of the disturbance, cause asignificant increase of the sound power in individual frequency ranges.

7. Sound Pressure- and Sound Power Level of Fans

The sound pressure level Lofthefans can be pre-determined using fan tip speed, fan impel-ler diameter, and certain constants. Dependingon fan type and performance data, averageevaluated sound pressure levels LA between90 and 110 dB (A) are common (these valuesare usually measured at a distance of one meter from the fan and at an angle of 45 de-grees to the flow direction). As an approximati-on the “A“-soundpower levels can be pre-deter-mined according to the following equation:

LWA = K + 10 lg + 20 lg in dB(A)

whereby:

p = Total pressure difference in : barp0= 100 : barV = Volume in m3/hrV0 = 1 m3/hrK = 11 dB (A) for centrifugal fans withcurved, backward inclinedK = 16 dB (A) for axial flow fans

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VºV0

pp0

The total sound power W produced by the fanor the respective sound power level Lw areused as the basis for determining the soundpropagation of fan noise.

Relative to noise radiation to the surroundings itis necessary to differentiate between

the primary sound power radiating withiagainst the gas stream through the fan out-let/inlet area (“gas-borne sound“)and the secondary sound power radiatingfrom the fan components (“body sound“)being excited by the sound energy of the gasstream.

Primarily emitted sound powers are WS andWD.

LWS: Level of the sound power radiating gainstthe gas stream through the inlet area.

LWD: Level of the sound power radiating withthe gas stream through the outlet area.

Secondary emitted sound powers are:WG, WU, WSL and WDL.

LWG: The sound power WG transmitted to hou-sing walls evokes structureborne sound (bodysound) that radiates in the form of air sound tothe surroundings. The respective sound powerlevel is LWG.

LWU: The structure-borne sound (body sound) ofthe housing is transmitted through sound con-duction to fixed components of the housing(especially supports) from where it radiates inthe form of air sound. The respective soundpower level is LWG.

LWSL: The sound powers WS

LWDL: WD radiating

as gas-borne sound through the inlet and outletarea of the fan evokes structure-borne sound(body sound) in the duct system which is notconnected to the fan mechanically, but by ex-pansion joints, and is therefore acoustically se-parated. This body sound in turn radiates tothe surroundings in the form of air sound.

Respective sound power levels are LWSL and LWDL.

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To determine these individual sound powers atthe measuring surface at a distance of one meter from the fan (defined in section 3) an approximation can be made to calculate thesound power, emitted in the form of air sound,by means of the following equation:

LWi = Li + 10 lg

where, for the respective individual componentsunder consideration:

i = S, SL, D, DL, G, U

With the example of a forced draft fan (withand without sound protection) Figures 10.1through 10.4 graphically depict the varioussound components, described above.

For primary and secondary sound sourcessound emissions are symbolized by arrows ofdifferent color and corresponding sound powerlevels are symbolized by arrows of differentlengths.

8. Sound Protection

The noise generated by the fan can be reduced through the installation of sound enclo-sures or acoustical insulation and lagging, onthe one hand (“sound insulation“)[“Schalldäm-mung“] and silencers on the other (“sound atte-nuation“) [“Schalldämpfung“].

When a silencer [“Schalldämpfer“] is installed,the sound propagation in the duct system is reduced without essential influence on the gasstream. (Values LWS and LWD are reduced byconverting sound energy to thermal energy.)

The installation of acoustic insulation and lag-ging [“Schallisolierung“] or sound enclosures[“Schallhauben“] provides extensive protectionto the area surrounding the fan from the propagation of air sound caused by the structure-borne sound (body sound) of the fancomponents excited by the sound energy of thegas stream. (Values LWG, LWSL, LWDL are reducedby the reflection of sound energy back to thenoise source and in addition partially by conversion to thermal energy. Depending onindividual requirements, sound attenuation in

fans can be achieved by untuned absorptionsilencers or resonant silencers tuned to certainfrequencies (these resonant silencers are alsoknown as interference, chamber or λ/4 silencers).

For both types of silencers, baffles [“Kulissen“]are arranged within a housing, parallel to thedirection of flow.

The attenuation principle chosen (friction or reflection with interference) determines the design of the baffles.

In the case of absorption silencers the spacebetween the baffle walls, built of perforatedplate, is filled with sound absorbing mineralwool.

The molecules in the gas stream excitedto produce sound oscillations are impededbythe mineral wool packing in the baffles suchthat the sound energy penetrating the perforati-ons is converted to thermal energy by the frictionof the molecules.

The absorption silencer isused for reducing the noise level of a wide band soundspectrum. Continued troublefree operation of this silencercan only be achieved if it isused in a relatively clean envi-ronment. In a dust laden atmosphere the dust will blockthe perforations in the bafflewalls, thus reducing effective-ness.

In dust laden air or gas stre-ams resonant silencers (λ/4silencers, interference si-lencers or chamber silencers)are used. This type of silencer,however, has only limited ef-fectiveness in reducing broadband noises. According to itsattenuation principle, this silencer primarily reduces protruding single tones.

By adding sound absorbingmineral wool mats to the bafflechamber plates a certainbroad band attenuation is obtained in addition to thesingle tone attenuation (see Figure 9).

The effectiveness of the singletone attenuation is explainedby the principle of reflectionand interference. The most im-portant dimension when desi-gning a resonant silencer isthe depth of the chamber t;this dimension must be appro-ximately 1/4 of the wavelength of the pitch to be atte-nuated (t = λ/4) to cause thefollowing action: At a distance λ/4 from theouter baffle wall the soundwave hits the solid back wallof the chamber, is reflectedand travels another distanceof λ/4 back to the soundsource where the reflectedsound wave, traveling 2 xλ/4 or λ/2 relative to thenext following sound wave,arrives with a 180 degreesphase displacement and thuscauses interference (tone elimi-nation).

Axial flow impulse induced draft fan in a powerplantwith heat-sound insulation and laggingand discharge silencer.Volume flow V = 660 m3/sTemperature t = 156 °CPressure increase Δpt = 6520 PaSpeed n = 590 1/minShaft power Pw = 5480 kW

Axial flow induced draft fan (axial flow impulsefan) with heat-sound insulation and lagging anddischarge silencer for reducing the sound levelemitted from the stack outlet.

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SS0

Sound Radiation of a Forced Draft Fan withoutSound Protection (Figure 10.1.)

LW Total sound power generated by fan

LWD Gas-borne sound power radiating in flow direction through the discharge area

LWDL Sound power radiating from the discharge duct as air sound due to LWD and body sound transmission

LWS Gas-borne sound power radiatingagainst flow direction through the inletarea.

LWSL Sound power radiating from the inlet duct as air sound due to LWS and body sound transmission (Figure 10.2.and 10.3.)

LWG Sound power radiating from the fan housing as air sound due to body sound excitation through the sound energy in the gas stream

LWU Sound power radiating as air sound from the support structure due to bodysound conduction from the fan housing

LWM Sound radiation by attached or neighboring machinery (for instancefan motor drive)

Sound Protection of a Forced Draft Fan by Means of Inlet Silencer and Acoustic Insulation and Lagging (Figure 10.2.)

Attenuation of the sound powerLWS through an inlet silencer

Insulation of the sound powerLWG, LWDL, LWSL through acoustic nsulation and lagging. Since no reduction of the gasborne sound energyoccurs inside the system the sound willradiate at full level from all surfaceswhere there are gaps in the acoustic insulation and lagging.

The sound power LWD radiates into the duct system.

Sound Radiation and Sound Protection.

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p

Sound Protection of a Forced Draft Fan byMeans of Inlet and Discharge Silencers andAcoustic Insulation and Lagging (Figure 10.3.)

In addition to the protection shown in Figure10.2. the air-borne sound power LWD radiatingthrough the discharge area is reduced by adischarge silencer.

Sound Protection of a Forced Draft Fan by Means of Sound Enclosure with Integrated Inlet Silencer (Figure 10.4.)

The sound powers LWG LWU radiating from the fan as air sound as well as the sound power LWM radiating from the motor are insulated by the enclosure.

The silencer integrated in the sound enclosure attenuates the sound power LWS radiating from the inlet area of the fan to the allowable value.

The sound power LWDL radiating to the environment from the discharge duct can be reduced either by acoustic insulation and lagging (in this case, however, the sound power LWDradiates into the ductsystem) or through the addition of a discharge silencer.

For the design of sound enclosures appropriateventilation must be assured to remove fan andmain drive generated heat so that the maxi-mum permissible temperature within the soundenclosure can be maintained. If necessary, for-ced ventilation has to be used (for example forhot gas fans).

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p

Sound attenuation: (through absorption)

Sound insulation:Sound energy penetrates through porous wallsand is converted to thermal energy through vis-cosity friction. Sound energy strikes non-porouswalls and is reflected.

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