TLT Industrial Fans

Industrial Fans

Delivery program▲ Centrifugal Fans▲ Axial Flow Impulse Fans▲ Sound Protection

TLT-Turbo GmbH

Mine ventilation fan

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Introduction

Field of Application

Product lines

Fan Designs

Control Modes and Characteristic Curves

Design and Fabrication

Fan Inquiry

Explanation of Common Fan Terms and Special Problems

Questions Regarding Fan Noise

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Table of contents

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The requirements imposed on IndustrialFans have noticeably increased over theyears. The variety of problems that needto be tackled when handling gasesrequires a comprehensive range of fansto optimize the selection for each parti-cular application.

Decades of intensive research and oper-ating experience gained during this timeare the basis for our fan range that pro-vides the best economical choice forany application. Guiding factors for thedevelopment of this range have been:

• Low Investment Cost

• Low Operating Cost

• High Reliability

• Long Life

• High Noise Attenuation

Fans from the range have been supplied tothe following industries:

Steam Generators and Power Stations

Centrifugal and Axial Induced Draft FansForced Draft Fans for all pressures

Introduction Field of Application

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

Centrifugal F. D. fan with inlet vane con-trol I and inlet silencer. ▼

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

Mining IndustriesMine Fans for use above or belowground. Centrifugal and axial flow fansfor all air quantities.

Steel and Metallurgical IndustriesFans of all types for:Sintering Systems (Sinter Plants)Pelletizing Systems (Pellet Plants)Direct Reduction SystemsDry and Wet Particulate RemovalSystemsSoaking Pits and Walking BeamFurnacesEmergency Air SystemsIndirect Induced Draft Systems (PowerStacks)

Coke Oven PlantsCoke Gas Booster Fans, single and dou-ble stage, made of welded steel plate.

Marine Industries Forced Draft Fans.

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

Mine ventilation fanVolume flow = 383 m3/sDepression ∆pSyst = 5400 Pa

Speed n = 440 1/minSheft power Psh = 2560 kW

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Product lines

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

Centrifugal FansAxial Flow Impulse Fans Inlet Vane Controls(Action Type Axial Flow Fans) Support Structures

Indirect l.D. Fan SystemsSilencers (Power Stacks)Acoustic Insulation & LaggingSound Enclosures Emergency Air Systems

Double width double inlet exhaust gasfan on an electro-melt furnace particu-late removal systemVolume flow = 126 m3/sTemperature t = 120 °CPressure increase ∆pt = 3820 Pa

Speed n = 990 1/minShaft power PSh = 595 kW

Double width double inlet emergency

F. D. fan in a utility power plant.

Volume flow = 350 m3/s

Pressure increase ∆pt

= 9320 Pa

Speed n = 990 1/min

Shaft power PSh

= 4100 kW

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Fan Designs

The above diagram shows a summaryof the operating ranges for the variousfan designs.

Indirect l.D. fans (power stacks) areoften used for exhaust gases at tempe-ratures above 5000C

Single or multiple stage centrifugal fanswith maximum efficiency at pressuresup to 80,000 Pa. Standard and heavyduty designs are available.

Double width double inlet centrifugalfans for high pressure and large flowvolume. Standard and heavy dutydesigns are available.

Axial flow impulse fans with adjustableslotted flaps for high pressures at low tipspeeds.

The fan range of TLT includes:

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Control Modes and Characteristic Curves

Fan efficiencies around 90% will reduceoperating costs to a minimum level.However, not only are the fan eflicienciesat the maximum operating point ordesign point of importance but also effi-ciencies across the system operatingrange have great significance.

The most effective type of fan control isobtained by variation of the fan speed.Since speed control can only be achie-ved with high cost drive systems wecommonly utilize inlet vane control forboth centrifugal and axial flow fans. Inthe diagrams shown, the 100% point( = 100% and ∆Pt = 100%) represents

the optimum point. For various reasonsthe optimum point may not always beidentical with the design point.

Characteristic curves of a centrifugalfan with speed control

Characteristic curves of a centrifugalfan with inlet vane control

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Model tests in the laboratory have provi-ded the data needed to determine spe-cific performance characteristics of ourtans, in particular in view of the effects ofdifferent control systems. These perfor-mance characteristics enable the plan-ner to predetermine the specific beha-vior of the fan in a system. Precise pre-dictions can be made regarding the ope-ration of fans operating as single units orin parallel.

If performance verification of large fansis required, tests can be conducted eit-her in the field or in some cases on ourtest stand.

Characteristic curves of an axial flowimpulse fan with inlet vane control.

Inlet vane control for a gas recirculationfan, largely gas-tight design; inlet dia-meter D = 2730 mm Ø

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Design and Fabrication

With few exceptions, the large variety ofavailable fan types nearly always permitsdirect coupling of the fan to the drivemotor. We prefer this arrangementbecause system reliability is optimiziedby avoiding interconnecting equipmentsuch as gear boxes, belt drives, etc. Thebasic design flexibility of our fans per-mitting alterations to or replacement ofthe impeller enables us to match actualoperating conditions if it is found duringoperation that they differ from the condi-tions on which the original design dataare based.Furthermore, slotted blade tip adjust-ment on centrifugal fan wheels or slottedflap adjustment on axial flow impulsefans are, in many cases, a simple meansto meet specific operating conditions.

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

Volume flow = 660 m3/sTemperature t = 156 °CPressure increase ∆pt = 6520 Pa


Diameter D = 4220 mm Ø

Inlet vane control:Diameter D = 4800 mm Ø

Double width double inlet I. D. fan forwaste heat boiler, fan support of lateral-ly flexible base frame design.Volume flow = 180 m3/sTemperature t = 245 °CPressure increase ∆pt = 4420 Pa

Speed n = 990 1/min

F.D. fan with inlet silencer in a steel mill.Volume flow = 77 m3/sTemperature t = 30 °CPressure increase ∆pt = 7260 Pa


Efliciency η = 84 %Diameter D = 2400 mm Ø

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High gas temperature or particulatematter entrained in the gas streamrequire specific attention in fan selec-tion and design. In such cases we oftenrecommend emphasizing increased reli-ability in lieu of maximized efficiencies.

Axial flow impulse I.D. fan designed forvertical installation, shown in the manu-facturing stage.

Impeller and shaft of an axial flowimpulse fan during balancing operation.

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Double width double inlet gas fan inelectro-metallurgical plant.Volume flow = 195 m3/sTemperature t = 230 °CPressure increase ∆pt = 3720 Pa


Motor power PM = 1100 kW

Double width double inlet sintering gasfan in steel plant.Volume flow = 366 m3/sTemperature t = 160 °C Pressure increase ∆pt = 14200 Pa

Speed n = 990 1/minEfficiency η = 84 %Shaft power PSh = 5900 kW


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Left: Impeller for a single stage F.D. fanVolume flow = 4.8 m3/sTemperature t = 20 °CPressure increase ∆pt = 31400 Pa

Speed n = 2980 1/minDiameter D’ = 1250 mm ØImpeller mass mImp = 210 kg

Right: Rotor for a two-stage gas fanVolume flow = 1.03 m3/sTemperature t = 100 °CPressure increase ∆pt = 28500Pa

Speed n = 2980 1/minDiameter D’ = 865 mm ØImpeller mass mImp = 140 kg

(both impellers)

Rotor for a mine fanDesign volume flow Des = 41 7 m3/s

(Maximum volume flow mas = 500 m3/s)

Temperature t = 20 °CDepression ∆pSyst = 5890 Pa

Speed n = 420 1/minImpeller diameter D’lmp = 5280 mm

Impeller mass mlmp = 14000kg

Shaft mass mSh = 9900 kg

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Rotor for a double width double inletsintering gas fan.Shaft attachment: Centerplate of impel-ler is bolted between flanges of a divi-ded shaft, centered on a very small dia-meter.Volume flow = 265 m3/sTemperature t = 160 °CPressure increase ∆pt = 16650 Pa


Impeller mass mImp = 11000 kg

Shaft mass mSh = 11000 kg

Rotor for an axial flow impulse fan(I. D. fan for a utility power station)Volume flow = 660 m3/sTemperature t = 156 °CPressure increase ∆pt = 6520 Pa

Speed n = 590 1/minDiameter D = 4220 mm ØImpeller mass mlmp = 12100 kg

Shaft mass msh = 5200 kg

(Hollow shaft)

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Below and right, foreground: Rotor ofSWSl (single width single inlet) flue gasfan for a steel mill. Torque transfer: hubshaft with key. Erosion protection:Bolted wear liners coated with wearresistant welds.Volume flow = 39.5 m3/sTemperature t = 150 °CPressure increase ∆pt = 13550 Pa

Speed n = 1145 1/minDiameter D = 3030 mm ø

Right, background: Rotor of DWDI (dou-ble width double inlet) flue gas fan for acement kiln. Torque transfer: integralhub with body bound bolts.



Left hand side of the picture: Rotor coa-ted with Saekaphen for a two-stagecoke gas fan.Volume flow = 3.9 m3/sTempersture t = 25 °C Pressureincrease ∆pt = 19650 Pa

Speed n = 2970 1/mm Diameter D = 1224 mm ø

Right hand side of the picture: Rubberlined impeller for a flue gas fan behind aventuri scrubber.Volume flow = 17.6 m3/IsTemperature t = 72 °C Pressure increase ∆pt = 9810 Pa


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Lead coated inlet box of the scrubberfan.

Two-stage F. D. fan for a waste gascombustor, the fan system consisting oftwo fans arranged in line with one com-mon motor drive.

To minimize spare part requirements therotors are identical in design (1st stageand 2nd stage).Volume flow = 5.1 m3/sTemperature t = 26 °CPressure increase ∆pt = 53900 Pa

Speed n = 2985 1/minShaft power PSh = 331 kw

SWSI (single width single inlet) fan, sup-ported on both sides, during shopassembly• Rotor of Incoloy• Housing and inlet box are lead coated• Dual fixed bearing system with flexible

support structure



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SWSI (single width single inlet) gas re-circulation fan, supported on bothsides, installed in a utility power plant.


Speed n = 720 1/minEfficiency η = 85,5 %Shaft power PSh = 457 kW

SWSI (single width single inlet) gas re-circulation fan, supported on bothsides, installed in an utility power plantsupported by integral base with vibra-tion isolators.Volume flow = 142 m3/sTemperature t = 361 °CPressure increase ∆pt = 7850 Pa


Diameter D = 3280 mm ø

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DWDI (double width double inlet) gas re-circulation tan with concrete tilled inte-gral base frame and vibration isolators.Volume flow = 124 m3/sTemperature t = 300 °CPressure increase ∆pt = 9615 Pa


Diameter D = 2320 mm ø

DWDI (double width double inlet) gasrecirculation tan during shop assembly.Volume flow = 250 m3/sTemperature t = 340 °CPressure increase ∆pt = 3470 Pa



Diameter D` = 3200 mm ø

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SWSI (single width single inlet) raw millfan for the cement industry, supportedon one side (AMCA arrangement 8).Volume flow = 114 m3/sTemperature t = 90 °CPressureincreaso ∆pt = 5700 Pa

Speed n = 745 1/minImpeller diameter DImp = 3350 mm ø

SWSI (single width single inlet) cementkiln exhaust gas fan, supported on oneside (AMCA arrangement 8), fan supportof laterally flexible base frame design.Volume flow = 133 m3/sTemperature t = 100 °CPressure increase ∆pt = 4600 Pa

Speed n = 745 1/min

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Our economical production facilities areequipped with modern machinery. Forexample, a numerically controlled flamecutting machine and a metal spinningmachine are used for the processing ofsheet metal.

Balancing machine for fan rotors up to30000 kg and 5000 mm Ø

Metal spinning machine for radii of inletnozzles, impeller side plates and spin-ning flanges.

NC flame cutting machine, with punchtapes automatically produced via elec-tronic data processor.

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A selection of varios designs from ourfan range is shown below.

SWSI (single width single inlet) forceddraft fan, supported on one side (AMCAarrangement 8), with inlet vane control,arranged on an integral supportingstructure with vibration isolators.

DWDI (double width double inlet) fan ar-ranged on an integral supporting struc-ture with vibration isolators.

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SWSI (single width single inlet) sinteringgas fan, rotor supported on both sides,with fluid drive, fan supported by anintegral steel frame.

Two-stage coke gas fan with oil lubrica-ted roller bearings sealed with carbonpacking glands. Speed: 2950 1/min.

Axial flow impulse fan (induced draftfan) with inlet vane control and possibi-lity of adjusting the slotted flaps indivi-dually during down-time.

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

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

3. Temperature: t in °C

4. System-related pressure increase: ∆pSyst in Pa

∆pSyst = Pt4 - Pt1*)

(Pressure distribution: Fan suction side / discharge side)

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

6. Elevation: h in m abovesee 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: or 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/m

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9.4 Dust characteristics:S: Probability of build-upV: Probability of erosion

9.5 Corrosion:K: Probability of corrosion due to

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

11. Additional information:

Fan Inquiry:

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As a fan manufacturer TLT works closelywith the architect I engineer and/or thefinal user to optimize fan selection foreach specific application.

The following set of conditions at taninlet to be supplied by the customer pro-vides the basis for fan selection anddesign:

Types of Fan Design and Installation

Radial-flow fans are normally of the sin-

gle-inlet type up to approx. 100 m3/s;

in some cases the double-inlet type is

used from 60-70 m3/s already. The actu-

al limit between single- and double-inlettype is mainly determined by the relevantcase of need, the suitable fan type, therequired ratio volume/specific energyand the speed.

We built single-stage radial-flow fans fora specific energy of more than 40000J/kg. It depends on the case of need,temperature, volume/pressure ratio andthe possible speed from what pressureincrease the fan has to be of the double-stage type.

The most favourable installation of thefans is that on an elevated concrete sub-structure. This results in shont bearingpedestals and the motor can be placedon a low frame or even directly on plateswhich are embedded in the concrete.This simple, rugged kind of installation isless susceptible to vibration and therefo-re best suited to sustain high imbalanceforces due to wear or dust caking.

If the substructure has to be made ofsteel corresponding plate cross sec-tions have to be used in case of larqebase-to-centre heights to achieve a suf-ficient vibration resistance. Thus the fanweight and the costs are increasedaccordingly.

Fans which have to be installed on sub-structures susceptible to vibrations, ase.g. in the structure or on a building roofhave to be vibration-isolated.

Notes:*) See also section"Pressure Definitions"."The system-related pressure increase∆pSyst", as defined by us, is often referred

to as"static pressure increase ∆pstat“, the

dynamic pressure components havingbeen ignored, however.

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

δ δ

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

Vibration-isolated installation with elevated concre-te block on vibration dampers.

Some installation examples:

Elevated concrete base with low frame. Compact type of construction, selfsupportingstruc-ture, placed directly on vibration dampers.

For this purpose, the compact type of con-struction is suitable which is a frameless, self-supporting structure utilizing the rigidity of thealmost totally enclosed suction boxes andhousing. The vibration isolators are installedunder the "hard" points such as housingwalls.The installation of the fan on an elevated con-crete base resting on vibration

isolators is recommended if higher imbalan-ces are expected.

The vibration isolators reduce the amplitudesof the dynamic forces (alternating loads fromthe imbalance rotating with the fan rotor). Theso-called isolation efficiency depends on thedistance of the exciter frequency (= fan speed)and the natural frequency of thespring mass system of vibration isolator -

machine mass. Normally, the isolation efficien-cy is above 90%. For speed-controlled fans itis therefore absolutely necessary to indicatethe lowest required (reasonable) speed, asthen accordingly "soft" springs have to beused.

Explanation of Common Fan Terms and Special

Problems

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The following definitions of fan termino-logy may facilitate communication bet-ween the fan manufacturer and custo-mer.

Meaning of Symbols:p Pressure∆p Pressure Difference or IncreasepV Pressure Loss, Resistance

Volume Flow at Inlet ConditionA Cross Sectional AreaI Length

Mass Flowc Gas Velocity

Densityf Compressibility FactorY Specific Delivery WorkPSh Shaft Power

η EfficiencyT Absolute Temperature (Kelvin)χ Adiabatic Exponent

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

sections concerned (terminal points)

Pressure DefinitionsThe energy transmitted through the fanimpeller to the volume flow is needed toovercome the system resistances.These resistances can comprise the fol-lowing:• Friction losses• Back pressure from pressurized

systems• Velocity changes at system inlet and

outlet and within the system• Draft forces due to density differentals• Geodetic head differences which are

mostly negligible.The sum of the above mentioned re-sistances as far as they occur in thesystem concerned, represent the totalsystem resistance. According toBernoulli this totalresistance is to be understood as totalpressure. This pressure pt (analogy: total

energy) comprises static pressure ps

(analogy:potential energy) and the dyna-

mic pressure pd (analogy: kinetic ener-

gy).

pt = ps + pd

(This definition is in accordance withVDMA-standard 24161)

To move the design volume flow the fanmust generate - within the specified fanterminal points - a pressure increaseequivalent to the total resistance of thesystem. This equivalent pressure increa-se is defined by us as system-relatedpressure increase ∆pSyst.

∆pSyst = pt4 - pt1

If the cross sections 1 and 4 (Figure D-1) represent the terminal points of thefan the system-related pressure increa-se ∆pSyst will then be the difference of

the total pressures at the terminalpoints 1 and 4. This system-relatedpressure increase which is in fact thepressure increase required by thecustomer is often incorrectly still refe-red to as static pressure increase∆pstat unfortunately ignoring the dynamic

pressure difference existing in mostcases. The probable cause of disregar-ding this dynamic pressure increase liesin the generally used method of measu-ring the static pressures in ducts bymeans of holes drilled in the duct wallsperpendicular to the direction of flow. Insuch cases the dynamic pressure diffe-rential at the terminal points has to beadded to the result of the static pressu-re measurement to obtain the correctsystem-related pressure increase.In cases where no specific data relativeto gas velocities, desired duct crosssections or special installation require-ments such as for mine fans are givenwe will determine the fan based on theassumption that the cross sections A1

and A4 are equal and the total pressure

increase between these terminal pointsrepresents the system-related pressureloss ∆pSyst stated by the customer.

A1 = A4

∆pSyst = pt4 - pt1

With A1 = A4 and disregarding compres-

sibility it follows that Pd4 = Pd1 and there-

fore:

The dynamic pressure Pd is under-

stood to be based on the average gasvelocity in a cross section.

Example: Dynamic pressure at terminalpoint 4:

Pressure losses pV caused by fan

com-ponents between the cross sec-tions A1 and A4, for example inlet box,

louver, inlet vane control, diffuser will beconsidered by us when sizing the fan.

The design pressure ∆pt, the parameter

determining fan size, is the sum of thesystem-related pressure increase andthe pressure losses of the fan compo-nents.

Our characteristic curves show thedesign pressure ∆pt, as this pressure

differential represents a parameter defi-ned by tests for a specific fan type atgiven operating conditions, whereas thesystem-related pressure increase ∆pSyst

varies with the losses pV which arise

depending on the fan componentsused.In determining the fan design pressure∆pt other losses are considered in addi-

tion to the above mentioned pressurelosses pV if special design conditions

are specified involving inlet and outletpressure losses, e. g. pressure losses inthe case of silencers and turning bends,outlet pressure losses in the case ofmine fans etc.

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2

δ

.

_2

δ

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. _2

δ

1 __2

A4

2 . 2V°

_pd = c2

pd4 = c = ( ) f

∆pt = ∆pSyst + pV = pt3 - pt2

∆pSyst = (ps4 + pd4) - (ps1 + pd1) ∆ps

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Fan Power and EfficiencyThe fan design pressure ∆pt of our fans

is equal to the total pressure increasebetween the cross sections A2 and A3.

With this design pressure, and thedesign volume flow at fan inlet condi-tions, the power requirements at the fanshaft PSh and the efficiency η will bedetermined, optimum inlet conditionsbeing a prerequisite.

in m3/s

∆pt in Pa = N/m2

f < 1η < 1

in kg/sY in J/kg = Nm/kgPSh in W = Nm/s

Influences of Temperature and DensityA noticeable temperature rise willoccur across fans with high pressureincrease, in particular when the fan ope-rates in a throttled condition and at lowefficiency. The adiabatic temperatureincrease ∆tad is:

The real temperature increase is:

∆pt in Pa

∆t in °Cη < 1

Operating a fan at a temperature signifi-cantly below design temperature willcause the shaft power requirement toincrease as a function of density or as aratio of the absolute temperatures.Should a different gas with higher densi-ty be handled the shaft power require-ment will increase with the ratios of thedensities.

Since such operating conditions oftenoccur at start-up the louvers or inletvanes need to be closed in these cases.The pressure increase of the fan willalso rise as a function of lower tempera-tures or higher gas densities. This mustbe considered when designing flues andducts, expansion joints, etc.

Influence of Mass and Mass InertiaErosion, corrosion and system-relatedcontamination causing build-up canposibly lead to imbalances due to un-even mass distribution. For such casesimpellers with large masses are advan-tageous because the shifting of the gra-vity center caused by the imbalance issmaller.For the start-up time of a fan the deter-mining factor is the inertia of the rotatingmass. This start-up time is of importan-ce relative to temperature increases ofthe electrical drive system causing limi-tations of the number of system start-ups.

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

V°

_______V° .∆pt . fη

_____m° .Yη

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m°

2 .χ−1χ

____

___∆tad

η∆pt_______

1250 η.

Tdesign

Tcold

_____ .

δδ

alternativ gas

design

________.PSh= =

∆tad = T [ (p3/ps) -1]

∆t = ≈

PSh operation = PSh design

= PSh design

∆pt operation = ∆pt design

= ∆pt design

Tdesign

Tcold

_____ .

δδ

alternativ gas

design

________ .

Meaning of Symbols Used in Equations:

∆pt Total Pressure Increase Between Cross Sections 2 and 3

∆pSyst System Resistance-Related Total Pressure Increase

∆pSyst, s System Resistance-Related Static Pressure Increase

pd Dynamic Pressure (Velocity Pressure)

∆pd Dynamic Pressure Difference (Velocity Pressure Difference)

pV Pressure Loss(es)

pV In Pressure Loss at Fan Inlet

pV Box Inlet Box Pressure Loss

pV Dif Diffuser Pressure Loss

pV Out Outlet Pressure Loss

A Cross Section

Indexes:

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

Examples of Various Fan Arrangements in a System with

Corresponding Pressure Distribution (Figures D-2 to D-4)

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Behavior ofTotal Pressure

Static Pressure

Dynamic Pressure

(Velocity Pressure)

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

∆pt = ∆pSyst + ∑ pV

∆pSyst = ∆pSyst, s + ∆pd

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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)

Questions Regarding Fan Noise

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Discharge silencer for two centrifugalforced draft fans, designed as absorp-tion type silencer, level reduction by 15dB. Fan data:

Volume flow = 2 x 62 m3/s

Temperature t = 50 °CPressure Increase ∆pt = 8120 PaSpeed n = 1490 1/min

Below:Double width double inlet centrifugal for-ced draft fan with disc silencers andcover for noise treatment of the fan inletnoise (shown during shop assembly).

1. First FundamentalsWith progressing industrialization man isfaced with increasing environmental pro-blems. Noise emitted by fans belongs inthis category.The following will give guidance in theproblem area of noise emitted by fans aswell as the flow in the connecting fluesand ducts.

The sound [“Schall“]1) perceived by the

human ear is the result of oscillation ofparticles of an elastic medium in thefrequency range of about 16 to 16,000hertz (Hz). One hertz is one oscillationper second. Depending upon themedium in which the sound travels wedistinguish air sound, body sound, andwater sound [“Luftschall, Körperschall,Wasserschall“].A pitched tone [“Ton“] is defined assound oscillating as a sinusoidal func-tion (compression and depression). Withincreasing amplitude sound will be per-ceived as being louder and with increa-sing frequency it will be perceived asbeing higher. The tone in Figure 2 (soundpressure P2) is perceived as higher and

generally louder than the tone in Figure 1(sound pressure P1).

For additional details see paragraph 2.A clang [“Klang“] is created by theharmo-nic interaction of several tones.Noise [“Geräusch“, “Rauschen“] is de-fined as statistical sound pressure distri-bution across the perceivable frequencyrange. A noise annoying the human earis called an “excessive noise“ [“Lärm“].

1)The terms in square brackets [ ] are

the equivalent German words.

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2. Human Noise PerceptionSound pressure is exactly measurablewith instruments. The physiologicaleffect on humans is much more difficultto determine. The human ear, for exam-ple, will perceive two tones of equalsound pressure yet different frequencyas unequally loud.Numerous tests were made on listenersin order to compare the loudness oftones at different frequencies and diffe-rent sound pressures with those of a1000 Hz tone. In particular the objectivewas to identify the sound pressure px

(measured in dB)1)at a frequency of 1000

Hz at which the sound pressure pn (mea-

sured in dB) and the frequency fm (mea-

sured in Hz) would evoke the same per-ception in the listener with respect toloudness. As a result of these tests, cur-ves of constant loudness (stated inphone) were identified over the frequen-cy range. By definition sound pressurelevel and loudness coincide in terms offigures at 1000 Hz. Graphs in Figure 3show these curves of equal loudness.

1)The sound pressure is often referred to as

sound pressure level, see specifics underparagraph 3 "Fundamentals of Acoustics".

2)The reason for the special nuisance created

by a single tone is its information content(Example: Tones produced by sirens, warn-ing and mating calls in the animal world).

Because the shape of the curves chan-ges with frequency as well as soundpressure one was faced with the pro-blem of designing a handy measuringinstrument for an objective measuring ofthe loudness of sound. This was theimpetus behind the search for a differentevaluation system. An additional reasonlies in the fact that the phon curves canonly be used to evaluate single tones.There is, however, a difference betweenthe human ear's perception of singletones and its perception of noise.The solution, which takes these factorsinto account and which has been inter-nationally accepted, is found in the so-called “A“ sound evaluation curve. Thecurve represents an approximation ofthe phon curve in midrange of the soundpressure level. To give consideration tothe fact that single tones are perceived

as being more annoying2)

than broadband noise, a higher reduction is impo-sed on single tones in addition to thetotal noise level requirements, for exam-ple such a typical single tone is the“blade passing tone“ [“Schaufelton“] of afan whose frequency is calculated withthe number of blades and the fan speedexpressed in Hz. This blade passingtone and its integer multiples (harmo-nics) form the so-called “blade passingfrequencies“ [“DrehkIang“].

3.Fundamentals of Acoustics(Definitions)Units of Sound ParametersIn acoustics it is common to work withlevels, i.e. it is common not to use theoriginal parameters with their correspon-ding units, but Iogarithmical parameterratios using the logarithm to the base 10,the corresponding units being be (B) ordecibel (dB).

As the same units are applied to allsound parameter levels it is important toproperly identify the type of the soundparameter level referred to, that meansto distinguish, for example, betweensound pressure level and sound powerlevel.

Sound Pressure Level L[“Schalldruckpegel“ L]The sound pressure level L (most com-monly also called sound level) quanti-fies the sound pressure measured at aspecific point.

By definition:

with p = effective value of sound pressure at measuring point in Nim2

p0 = 2x10-5

N/m2

= 20 µ Pa

= 2 · 10-4

µ bar

(reference sound pressure, the audiblethreshold for 1000 Hz pitch)

Evaluated Sound Pressure Level LA

(= Sound Pressure Level Evaluated Ac-cording to Evaluation Curve“A“)[“Bewerteter Schalldruckpegel“ LA]. The

evaluated sound pressure level LA -

expressed in dB (A) - is obtained byadding at the various frequencies a ∆Lfrom the evaluation curve “A“ (see Fi-gure 4) to the measured sound pressurelevel L at the corresponding frequencies.The evaluation curve and the evaluationprocedure are defined in DIN standard45633, sheet 1.

effective valueof sound parameter

reference valueof sound parameter

level = Ig in Bof soundparameter

effective value

reference valuelevel = 10 Ig in dBof soundparameter

___________________

________________

L = 10 lg = 20 lg in dBp

2

p02

__p

p0

___

30

Baffles of the absorptive discharge silen-cer of a forced draft fan (after approxi-mately 11,000 operating hours); fan per-formance data are shown on the right.Below: Close up photograph of the baffle wall~

after approximately 11,000 operatinchours.

Three-stage absorptive silencer forambient air inlet to a forced draft fan(Volume flow

= 433 m3/s, pressure increase

∆pt = 8250 Pa) in a power station.

Design point = 60% Attenuation to sound pressure level70dB (A).

V°

V°

31

Discharge silencer designed as a reso-

nant silencer (λ/4-silencer or interfer-ence silencer) for two induced draft fansin a power plant (volume flow = 2 x660 m3/s, pressure increase ∆pt = 6520

Pa), insertion loss = 33 dB at the fre-quencies of 118/236 Hz (blade passingfrequency and first harmonic).

Baffles (shown at center right) and bafflewalls (shown below) of the above silen-cer installation after approximately11,000 operating hours.

V°

32

As can be seen in Figure 4, the numeri-cal values for LA are significantly below

the L values at low frequencies andhave a much smaller impact at higherfrequencies.

Measuring Surface Sound PressureLevel L and LA

[“Meßflächen-Schalldruckpegel“ L undLA]

The-measuring surface sound pres-sure level L (= the sound pressure levelat the enveloping measuring surface) is

defined as the energetic mean1)

of multi-ple sound level measurements over themeasuring surface S with eliminationof extraneous noise and room effects(reflections).

LA is the “A“ evaluated measuring sur-

face sound pressure level.The measuring surface S is an assumedarea encompassing the sound source ata defined distance (mostly one meter).This enveloping surface comprises sim-plified surfaces such as spherical, cylin-drical and square surfaces generally fol-lowing the shape of the sound produ-cing equipment.

1)To calculate the energetic mean of all

sound level measurements taken overthe envelope surface (taking intoaccount the time interval of testing) thefollowing formula is to be used:

If the difference between the individualsound levels is smaller than 6 dB theformula below can be used as anapproximation representing the arithme-

tical mean:

Any components that protrude beyondthe surface but contribute little to theemission can be neglected. Soundreflecting boundaries, such as floorsand walls, are not incorporated withinthe measuring surface. The measure-ment points shall be sufficient in num-ber and evenly distributed over theenveloping surface. The numberdepends on the size of the sound sour-ce and the uniformity of the sound field.

Because of the logarithmical parameterratios used in acoustics the measuring

surface in m2, will be related to a refe-

rence area to define the measuringsurface level LS [“Meßflächenmaß“ LS]

as the characteristic parameter:

S = Measuring surface in m2

S0 = 1 m2

(reference area)

Sound Power Level LW

[“Schall-Leistungspegel“ Lw]

The value of the total sound power radi-ating from a sound source is given bythe sound power level LW.

W = gas-borne acoustical poweremitted 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 onedescribed in the example, of the soundpressure level is conducted, using theevaluation curve “A“, the evaluatedsound power level LWA will be obtained

from the sound power level LW.

Relationship Between SoundPressure Level and Sound PowerLevelThe sound power W is not measureddirectly but is calculated using the mea-sured sound pressure p, sound particlevelocity ν (molecular movement veloci-ty), [“Schallschnelle“ ν] and the measu-ring surface S:

W = p · ν · S

using ν =

= air densityc = air sound velocity

it follows that:

Assuming that = constant c = constant

the proportional relationship obtained is:

W ~ p2

· S.

In terms of expressing the above equa-tion in acoustic level parameters the fol-lowing important equations can beobtained:

The sound power level LW can be ap-

proximated by the sum of the measu-ring surface sound pressure level L andthe measuring surface level LS.

From the above relationship it can bededuced that with a given sound powerlevel a spherical or a semisphericalsound dispersion (ideal sound disper-sion) the sound pressure level will dimi-nish by 6 dB for every doubling of thedistance from the sound source.

Through absorption of the sound in theair and on the ground this value willincrease and through reflection of thesound by obstructions it will be redu-ced. Furthermore, weather conditionscan cause either an increase or decrea-se of the sound pressure level reduc-tion.

SS0

LS = 10 Ig in dB

LW = 10 Ig WW0

in dB

_ _

__

_

_

L = 10 lg ( · ∑ 10 )1n_ i = n

i = 1

0.1 Li

_

_

L ≈ · ∑1n_

_Li

i = n

i = 1

_

p· c

____δ

δ

W = · Sp2

· c____δ

δ

LWA ≈ LA +10 Ig = LA + LS in dBS__S0

_ _ _ _

LW ≈ L +10 Ig = L + LS in dBS__S0

_ _ _ _

_

Acoustics Electrotechnics

N = Output

W = p · ν · S U = Voltage

I = Amperage

= I R = Resistance

I = p · ν N = U · I

ν = Ι =

N = = I2

· R

Ι ~p2

N ~U2

33

Sound Intensity Level LI

[“Schall-lntensitatsspegel“ LI]

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

m2.

With this definition an analogy to elec-tricaltechnology can be made: The sound intensity is proportional tothe square of the sound pressure.The definition of the correspondingsound intensity level LI is as follows:

with l0 = 10-12

watts/m2

(reference sound intensity)

4. Sound AnalysisThe “total sound level“ or “sum soundlevel“ of noise is derived from the loga-rithmic addition of a multiple of singlesound levels at different frequencies(Figure 5). In order to perform noisemeasurements, the audible frequencyrange has been divided into 10 octavebands.

The width of the octave is identifiedsuch that the ratio of the upper limitingfrequency of the spectrum f0 to the

lower limiting frequency fU is 2:1.

Octave:

The corresponding ratio for the “terz“is:

“Terz“:

Three “terz“ together make up an octa-ve.

Center frequencies are determined by:

Octave:

“Terz“:

The sound intensity is proportional to thesquare of the sound pressure

The individual center frequencies of theoctave band are at:

31.5 Hz63 Hz125 Hz250 Hz500 Hz

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

16,000 Hz

I = inWS___ watts______

m2

LI = 10 lg in dBI

I0

___

= 2f0

fU

___

f0

fU

___ = 3√¯¯¯¯¯2

fm = √¯¯¯¯ · fU =2f0___

√¯¯¯¯¯2

fm = √¯¯¯¯ · fU =2f0___

√¯¯¯¯¯26

6

fm = √¯¯¯¯¯¯¯¯¯¯¯fU · f0

WS__

I = = ν2

· · c

δ

p· c

____δ

p2

· c____δ

UR

___

U2

R___

34

In practical application, the first and lastoctave bands mostly play a secondaryrole. Commercially available sound mea-surement instruments to measure soundlevels in dB and dB (A) are equipped withadjustable octave and “terz“ filters toconduct frequency analyses. If the octa-ve band analysis proves inadequate themore selective “terz“ analysis should beemployed, the octave band width beingdevided into 3 “terz“ band widths.In the case of single tones or noises ex-tending over one “terz“ band only, the“terz“ band and the octave band analy-ses will give the same figures.The example in Figure 5 shows an octa-ve band and “terz“ band analysis.For a more selective analysis of a noisespectrum, narrower band filters can beemployed to further divide the noisespectrum (search tone analyzer).

5.Addition of LevelsTo determine the total level LtOt, partial

levels L (sound pressure level or soundpower level) will be added in accordancewith the following equation:

When adding sound pressure levels itmust be considered that all individualsound pressure levels have one commonreference point.In the specific case where n sound sour-ces have equal sound power W1, the

total level Lw10 can be determined bythe following:

For a number of sound sources the levelincrease can also be determined usingthe diagram in Figure 6.In the special case of two single soundsources with different levels, the totallevel is obtained by adding the differen-ce of the individual levels to the higherlevel.

With reference to the diagram in Figure 7it can be seen that for a level differenceof more than 10 dB practically no levelincrease will result. For the special caseof two sound sources with equal levels(level difference 0) a level increase of 3dB will result (see also Figure 6).The case where two superimposed sin-gle tones of equal sound pressure p1,

equal frequency f1, and equal phase ϕare to be added requires special consi-deration. Deviating from the above des-cribed summation a total level 6 dB hig-her than the sound pressure level of thesingle tone will be obtained:

If a difference in phase of 1 80 degrees

exists (ϕ1

= 00; ϕ

2= 180°) or

λ/2 interfe-

rence results, the tones eliminate eachother (see Figure 8).These two occurances are of practicalimportance in the case where two fansare

operating at almost equal speed in acommon duct system. In this casesound wave superposition results inperiodic sound level variations calledbeats. The beat frequency is determinedby the difference in the operating speedsof the two fans.

6. Noise Developement in FansThe operating noise of a fan comprisesvarious sound components.In boundary zones of confined fastmoving gas flow, eddy currents occur asthe result of the influence of the viscosi-ty of the gas. On fans these eddy cur-rents occur at the blade dischargeedges. The resulting noise caused by therotating impeller is referred to as “eddycurrent noise“ and is considered the “pri-mary noise“. Superimposed on thisnoise is the “self-noise of highly tur-bulent flow“ in the fan housing andducts. “Eddy current noise“ and “self-noise“ [“Wirbelgeräusch undStrömungsrauschen“] display a broadband frequency spectrum. the soundpower increasing approximately with the5th to 7th power of the impeller tipspeed. In addition to broad band noise,“pulsation noise“ [“Pulsationsgeräusch“]occurs at different frequencies causedby periodical pressure Oscillations of themedium due to the relative movementbetween the impeller and a stationaryfixture exposed to the flow. “Pulsationnoise“ will occur when the flow in the

Ltot = 10 lg ∑ 10 i = n

i = 1

0.1 Li

LW tot = LW1 + 10 lg n

Ltot = 10 lg ( 2· )2= 20 lg 2 ·

p1

p0

__ p1

p0

__

= L1 + 20 lg 2

35

closed environment of the impeller isdisturbed by obstructions with protru-ding edges (cut off in centrifugal fansand stationary guide vanes in axial flowfans). For fans such disturbing noise isalso referred to as “blade passing tone“[“Schaufelton“] or “blade passing fre-quencies“ [“Drehklang“], where the maindisturbing frequency (base frequency) isthe product of blade number times revo-lutions per second. Integer multiples ofthe base frequency can also occur asharmonics. The occurrence of the“blade passing frequencies“ (base fre-quency + harmonics) can, depending onthe type and intensity of the disturban-ce, cause a significant increase of thesound power in individual frequencyranges.

7. Sound Pressure- andSound Power Level of FansThe sound pressure level Lofthefanscan be pre-determined using fan tipspeed, fan impeller diameter, and certainconstants. Depending on fan type andperformance data, average evaluatedsound pressure levels LA between 90

and 110 dB (A) are common (thesevalues are usually measured at a distan-ce of one meter from the fan and at anangle of 45 degrees to the flow direc-tion).

As an approximation the “A“-sound-power levels can be pre-determinedaccording to the following equation:

whereby:p = Total pressure difference in µ barp0 = 100 µ bar

= Volume in m3/hr

= 1 m3/hr

K ≈ 11 dB (A) for centrifugal fans with curved, backward inclined blades.

K ≈ 16 dB(A) for axial flow fans.The total sound power W produced bythe fan or the respective sound powerlevel Lw are used as the basis for deter-

mining the sound propagation of fannoise.

Relative to noise radiation to the sur-roundings it is necessary to differentiatebetween

- the primary sound power radiating withi against the gas stream throughthe fan outlet/inlet area (“gas-borne sound“)

- and the secondary sound power radiating from the fan components (“body sound“) being excited by the sound energy of the gas stream.

Primarily emitted sound powers are WS

and WD.

LWS: Level of the sound power

radiating against the gas stream through the inlet area.

LWD: Level of the sound power

radiating with the gas stream through the outlet area.

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

LWG: The sound power WG transmitted

to housing walls evokes structure-borne sound (body sound) that radiates in the form of air sound to the surroundings. The respective sound power level is LWG.

LWU: The structure-borne sound (body

sound) of the housing is trans-mitted through sound conduction to fixed components of the housing (especially supports) from where it radiates in the form of air sound. The respective sound power level is LWG.

LWSL, The sound powers WS,

LWDL: WD radiating

as gas-borne sound through the inlet and outlet area of the fan evokes structure-borne sound (body sound) in the duct system which is not connected to the fan mechanically, but by expansion joints, and is therefore acoustically separated. This body sound in turn radiates to the surroundings in the form of air sound.

Respective sound power levels are LWSL and LWDL.

To determine these individual soundpowers at the measuring surface at adistance of one meter from the fan (defi-ned in section 3) an approximation canbe made to calculate the sound power,emitted in the form of air sound, bymeans of the following equation:

where, for the respective individualcomponents under consideration:

i = S, SL, D, DL, G, UWith the example of a forced draft fan(with and without sound protection)Figures 10.1 through 10.4 graphicallydepict the various sound components,described above.For primary and secondary sound sour-ces sound emissions are symbolized byarrows of different color and correspon-ding sound power levels are symbolizedby arrows of different lengths.

8. Sound ProtectionThe noise generated by the fan can bereduced through the installation ofsound enclosures or acoustical insula-tion and lagging, on the one hand(“sound insulation“)[“Schalldämmung“]and silencers on the other (“soundattenuation“) [“Schalldämpfung“].When a silencer [“Schalldämpfer“] is in-stalled, the sound propagation in theduct system is reduced without essenti-al influence on the gas stream. (ValuesLWS and LWD are reduced by converting

sound energy to thermal energy.)The installation of acoustic insulationand lagging [“Schallisolierung“] orsound enclosures [“Schallhauben“]provides extensive protection to thearea surrounding the fan from the propa-gation of air sound caused by the struc-ture-borne sound (body sound) of thefan components excited by the soundenergy of the gas stream. (Values LWG, LWSL, LWDL are

LWA = K + 10 lg + 20 lg in dB (A)pp0

____V°V0°

V°

V0°

LWi = Li + 10 lg SS0

36

depth of the chamber t; this dimension

must be approximately 1/4 of the wave

length of the pitch to be attenuated (t =λ/4) to cause the following action:

At a distance λ/4 from the outer bafflewall the sound wave hits the solid back

reduced by the reflection of sound ener-gy back to the noise source and in addi-tion partially by conversion to thermalenergy.

Depending on individual requirements,sound attenuation in fans can be achie-ved by untuned absorption silencers orresonant silencers tuned to certain fre-quencies (these resonant silencers arealso known as interference, chamber orλ/4 silencers).

For both types of silencers, baffles [“Ku-lissen“] are arranged within a housing,parallel to the direction of flow.

The attenuation principle chosen (frictionor reflection with interference) determi-nes the design of the baffles.

In the case of absorption silencers thespace between the baffle walls, built ofperforated plate, is filled with sound ab-sorbing mineral wool.

The molecules in the gas stream excitedto produce sound oscillations are impe-ded by the mineral wool packing in thebaffles such that the sound energy pene-trating the perforations is converted tothermal energy by the friction of themolecules.

The absorption silencer is used for redu-cing the noise level of a wide bandsound spectrum. Continued trouble freeoperation of this silencer can only beachieved if it is used in a relatively cleanenvironment. In a dust laden atmosphe-re the dust will block the perforations inthe baffle walls, thus reducing effective-ness.

In dust laden air or gas streams reso-nant silencers (λ/4 silencers, interfe-rence silencers or chamber silencers)are used. This type of silencer, however,has only limited effectiveness in reducingbroad band noises. According to its atte-nuation principle, this silencer primarilyreduces protruding single tones. Byadding sound absorbing mineral woolmats to the baffle chamber plates a cer-tain broad band attenuation is obtainedin addition to the single tone attenuation(see Figure 9). The effectiveness of thesingle tone attenuation is explained bythe principle of reflection and interferen-ce. The most important dimension whendesigning a resonant silencer is the

wall of the chamber, is reflected andtravels another distance of λ/4 back tothe sound source where the reflectedsound wave, traveling 2 x λ/4 or λ/2relative to the next following soundwave, arrives with a 180 degrees phasedisplacement and thus causes interfe-rence (tone elimination).

37

Axial flow impulse induced draft fan in apower plantwith heat-sound insulationand lagging and discharge silencer.

Volume flow = 660 m3/s

Temperature t = 156 °C Pressure increase ∆pt = 6520 Pa

Speed n = 590 1/min Shaft power PW = 5480 kW

Axial flow induced draft fan (axial flowimpulse fan) with heat-sound insulationand lagging and discharge silencer forreducing the sound level emitted fromthe stack outlet.

V°

▼

▼▼

38

Sound Radiation of a ForcedDraft Fan without SoundProtection (Figure 10.1.)LW Total sound power generated

by fanLWD Gas-borne sound power radiat-

ing in flow direction through thedischarge area

LWDL Sound power radiating from the

discharge duct as air sound due to LWD and body sound transmission

LWS Gas-borne sound power radiat-

ing against flow direction through the inlet area.

LWSL Sound power radiating from the

inlet duct as air sound due to LWS and body sound transmis-sion (Figure 10.2. and 10.3.)

LWG Sound power radiating from the

fan housing as air sound due tobody sound excitation through the sound energy in the gas stream

LWU Sound power radiating as air

sound from the support struc-ture due to body sound con-duction from the fan housing

LWM Sound radiation by attached or

neighboring machinery (for in-stance fan motor drive)

Sound Protection of a ForcedDraft Fan by Means of Inlet Silencerand Acoustic Insulation andLagging (Figure 10.2.)- Attenuation of the sound power

LWS through an inlet silencer

- Insulation of the sound power LWG, LWDL, LWSL through

acoustic insulation and lagging.Since no reduction of the gas-borne sound energy occurs inside the system the sound will radiate at full level from all surfaces where there are gaps in the acoustic insulation and lagging.

- The sound power LWD radiates

into the duct system.

Sound attenuation:(through absorption)Sound insulation:

Sound energy penetrates through porous walls and is con-verted to thermal energy through viscosity friction. Soundenergy strikes non-porous walls and is reflected.

39

Sound Protection of a ForcedDraft Fan by Means of Inlet andDischarge Silencers and AcousticInsulation and Lagging (Figure 10.3.)

In addition to the protection shown inFigure 10.2. the air-borne sound powerLWD radiating through the discharge area

is reduced by a discharge silencer.

Sound Protection of a ForcedDraft Fan by Means of SoundEnclosurewith 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 thesound 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 LWD radiates into the duct

system) or through the additionof a discharge silencer.

For the design of sound enclosuresappropriate ventilation must be assuredto remove fan and main drive generatedheat so that the maximum permissibletemperature within the sound enclosurecan be maintained. If necessary, forcedventilation has to be used (for examplefor hot gas fans).

40

Stronger Together

Die TLT wurde Anfang 2003 von der Franken-thaler Kühnle, Kopp und Kausch AG übernommenund firmiert als eigenständige Konzerngesell-schaft unter dem Namen TLT-Turbo GmbH.

Die Aktiengesellschaft KK&K, wurde im Jahr1899 durch den Zusammenschluß dreier Fami-lienunternehmen gegründet. Im Verlauf desvergangenen Jahrhunderts hat es dieKK&K stets verstanden, den Verän-derungen – vor allem techni-schen Erfordernissen – raschanzupassen.

Mit der Zielsetzung, sichwieder auf das traditionelleKerngeschäft zu konzentrie-ren, hat die KK&K die Turbo-ladersparte im Jahr 1998verkauft. Seither widmet sichder Konzern mit stetig wachsen-dem Erfolg wieder ausschließlichder Entwicklung, der Fertigung unddem Verkauf von Turbomaschinen – Tur-binen, Verdichtern und Ventilatoren für sämtlicheAnwendungsbereiche.

Bei allen den Tätigkeiten steht immer der Kundemit seinen individuellen Bedürfnissen im Vorder-grund. Auf der Basis dieser Kundenorientierungist das gesamte Handeln von einem Miteinander

geprägt – einem Miteinander mit dem Kunden,dem Mitarbeiter und dem Aktionär.

Eine ausgewogene Kombination von strategischerWeitsicht, kaufmännischer Umsicht und weltweiterKooperation mit kompetenten Partnern macht dieKK&K so erfolgreich und bestärkt uns in der

Auffassung, bestens auf die Zukunft vorbe-reitet zu sein. Den Motor des Erfolgs

bilden qualifizierte, zielorientierteund zufriedene Mitarbeiter so-

wie sehr moderne Geschäfts-prozesse.

Neben dem Verfolgen wirt-schaftlicher Ziele ist es derKK&K besonders wichtig,mit den Produkten einenaktiven Beitrag zum Umwelt-

schutz zu leisten. Dies doku-mentiert der Konzern mit der

Wahl der Worte „Clean Air”,„Clean Water” und „Clean Energy”

als strategische Leitbegriffe für das unter-nehmerische Handeln. Zum Konzern gehören ne-ben der TLT-Turbo GmbH auch die HV-Turbo A/Sin Dänemark sowie die PGW-Turbo in Leipzig.

Weitere Informationen zum Konzern erhalten Sieauf der KK&K-Website unter: www.agkkk.de

41

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Phone: + 49.6621.950-0Fax: + 49.6621.950-115

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© 2003 - TLT-Turbo GmbH · 3th edition 07/03/2.0/e · errors and changes reserved.

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TLT Industrial Fans