10th International Mine Ventilation Congress, IMVC2014

`© 2014, The Mine Ventilation Society of South Africa

1 MOTIVATION

1.1 Fan Type SelectionDuring the ordering phase for a main ventilation fana fan pre-selection process has to be followed,similar to the process in Figure 1.

Figure 1: Fan selction process, for the selection of the correctfan technology for the required ventilation task.

This paper will focus on Axial fans and analyseone of their major components, namely the impellerblades, regarding safety aspects.

For heavy duty rotary equipment endurance is amust, however commercial factors are becomingmore and more dominating, and it makes the

operating life of the equipment even more difficult.The direct comparison between overpressure type

fans and impulse type fans would reveal that there isnot a robust, reliable design on either side

Beside aerodynamic performance differencesboth fan solutions have their own strengths andweaknesses in respect of static and dynamicbehavior.

2 LOADS ON AXIAL FAN BLADES

2.1 Static and quasi-static loadsThe impeller blades of an axial fan is whereelectrical energy from the electric motor is convertedto mechanical engineering, by the movement of thegas through the fan impeller.

Impeller blades are subjected to variousmechanical loads of which a typical simplification isgiven in Figure 2.

These loads are independent of the fan type.

The most dominating force is the centrifugalforce ZF

2rmFZ (1)

m = generalized mass or a mass elementr = radial distance of m from the rotation axis

= the angular frequency

Moreover there are other loads also acting on theimpeller blade, such as the aerodynamic force,inertial force and the torque, all at the same time.

Structural statics and dynamics on axial fan blades

T. Neff & A. LahmTLT-Turbo GmbH, Germany

ABSTRACT: The failure of a fan blade in heavy duty rotating equipment can have fatal consequences andlead to a major loss of production. The static, dynamic and aero-elastic loads over the entire operatingrange of a fan was analyzed to determine the design rules that must to be implemented for the differenttypes of axial fan blades in order to achieve the highest possible life and ensure the safety of the system.

10th International Mine Ventilation Congress, IMVC2014© 2014, The Mine Ventilation Society of South Africa

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Figure 2. Typical load vectors representing the static anddynamic forces on an axial fan blade during operation.

As the impeller blades are typically twisted foraerodynamic efficiency improvements, not all of theblade mass is concentrated on a single plane or acentral axis. It is due to these offset masses thatadditional inertial forcessubsequently increasing the internal blade materialstresses.

The aerodynamic forces are initiated by thepressure distribution at the airfoil. They can befractionalized into a circumferential part, actingagainst the rotation and a tangential part, actingagainst the direction of flow. For more details, seeEckert&Schnell (1980).

2.2 Dynamic LoadsIn order to keep the hub in it s place in many cases,it cannot be avoided that structural support ribs areplaced upstream of the fan impeller. Depending onthe form and thickness of the support ribs, theintensity of vortices and wake turbulencesdownstream will vary.

Each impeller blade with its own naturalfrequencies fbl,i will be subjected to various otherdisturbances, within a single rotation of the impellerblade.

This blade passing frequency fpass is the mostimportant one

fpass = Zbl n / 60 (2)

Zbl = number of rotor bladesn = rotations per minute.Another important and powerful excitation is the

rotational speed and its multiples fn,i

fn,j = j n / 60

j = rotational multiples (3)

In case the structural elements in front of theimpeller that are not equally distributed, each

i leads to a frequency fsup,i

fsup,i i n / 60 (4)

i = angle between two structural elementsi = consecutive number

The downstream static guide vanes are also asource for generating vibrational excitations withinthe rotating blades, typically these guide vanes areequally spaced and the excitation frequency due tothese guide vanes fgv can be described as:

fgv = Zgv n / 60 (5)

where: Zgv is the number of guide vanes

Furthermore the electrical power supplyfrequency fel and 2 fel shall be checked forresonances with fbl,i, too. It is the challenging task ofthe blade design engineer to ensure that:

fbl,i pass n,i fsup,i gv el fel (6)

To ensure that the above equation (6) holds truethe blade stiffness can be altered. It is thereforeadvisable not only to tailor the blade from root to topaccording to stress calculations, but moreover to alsoincrease the thickness locally that will have aninfluence on the blade stiffness. A blade with a highstiffness will have a high natural frequency and viceversa.

What is the magnitude of stresses at the rotatingblades, caused by aerodynamic excitation fromstructural support elements upstream anddownstream of the impeller? As shown by Staiger(1991) the stress amplitudes can be considerablyhigh and careful positioning can improve the loadsituation of the impeller significantly.

The blades of an axial test rotor, with 12 bladeswere equipped with strain gauges and the influenceof different strut configurations on the blade stresseswere tested.

In the worst case, the comparison between nostruts and three equally distributed, radial struts ofcylindrical shape gave a stress amplification factorof eight.

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When a fan is operated inconditions the gas might contain elements whichmay cause abrasion and adhesion effects on theimpeller blades. Abrasion is the removal of originalblade material (weight reduction) by the working gaswhile adhesion is increasing the blade mass due toparticles in the working gas sticking to the originalblade material.

Both abrasion and adhesion will alter the naturalfrequency of the blades and can cause unwantedresonances additional to the original designedresonances.

To complete the picture regarding importantloads on axial fan blades, the driving torquevariances have to be mentioned as well.

These torque forces act as foot excitations to theblades and should not coincide with the naturalblade frequencies. When a variable speed drives isused, the pulsations in rotational speed can correlatewith e.g. the torsional resonance of the motor / fanimpeller system. This could increase the amplitudesof the above mentioned blade foot excitations.

It is however very unlikely that the blade will failearlier than the shaft system.

3 DESIGN RULES

3.1 Conventional engineering approachThree decades ago, when computers and FiniteElement Analysis (FEA) tools were much lesspowerful than today, the engineers had to simplifycomplex blade geometry in order to calculate thevarious blade stresses.

The sectional approach the most commoncalculation method during that time. This method isstill in use today. The blade is split from foot to tipinto portions (Fig 3) and each of these is loadedaccording to its radial position. While at the base (A-A) the centrifugal forces of all elements above areresponsible for the sectional stresses, at the top (H-H) there is only the element itself that contributes tothe load.

In case the blade has a circular foot to beconnected to the impeller, it can be assumed thatwithin an angle of 45° the flux of force isdeveloping.

Accordingly, the loaded section is reduced (Fig 3,A-A, B-B and C-C).

By approximation, the section modulus can bederived from the geometry together with the bendingmoment, the bending stresses per section can besuperimposed to the tensional stresses.

Usually von Mises combined stresses arecompared with the selected material properties. Thiskind of conventional stress calculation develops only

one value per section - the maximum of stress orstrain.

Other methods like FEA can be used whenstress distribution details are essential. This wouldalso enable stress values to be for any point on therelevant component.

Figure 3. Sketch about sections (A-A I-I) for conventionalblade calculation

3.2 Determination of natural frequenciesMost ambitious was - and still is - the estimation ofthe blade natural frequencies. The influence of thesupporting structure and consequently the fixing ofthe impeller blade to the hub is essential.

Figure 4: Typical impeller of a blade pitch axial fan with thepivot-mounted blade shaft and the bolted connection to theblade

It is obvious that a bolted blade (Fig 4) willbehave different from a welded blade (Fig 6).Nevertheless the bolted connection is also very rigidduring operation when the very high centrifugal

Blade toshaftconnection

10th International Mine Ventilation Congress, IMVC2014© 2014, The Mine Ventilation Society of South Africa

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force FZ stiffens the angular ball bearing of the bladeshaft to a remarkable extent. For a fan impeller ofØ3.7m and operated at 750rpm, FZ of one singlecasted steel blade can reach values of 350kN or evenmore.

In the case of individually mounted blades,measurements of natural frequencies can beconducted on a workbench. The blade is bolted on abig mass when a hammer is used to excite anoscillation and a vibration pick-up sensor records theresponse.

If the measured natural frequency would coincidewith one of the possible excitation frequencies, theblades would need a tuning e.g. by mass reduction(grinding, drilling holes from the blade tip), leadingto a higher natural frequency. This method allowsonly small steps in the range of up to 5Hz.

When measured frequencies are compared to thecalculated excitations during operation, thestiffening influence of the centrifugal force must beconsidered. More details about the analyticalcalculation are given in Traupel (1982). It might be achallenging task to find suitable blade dimensions,especially for the first three natural frequencies ofthe rotating blade fbl,1, fbl,2 and fbl,3 so that noconcord with the known excitations (see section 2.2)exists. Figure 5 illustrates it graphically.

Figure 5. Graphical overview of possible blade excitationfrequencies and the first 3 blade natural frequencies (example).

Conservative, safe intervals between fbl,1, fbl,2 andfbl,3 to some of the important excitation frequenciesare given in Table 1.

Table 1. Recommended safety margin between first three bladenatural bending frequencies and selected excitations.

fn1 fn2 fn3 fsup,i fpass

% % % % %+30 ±20 ±10 ±15 ±10

The analytical calculation of blade frequenciesrequires a lot of experience and tests for methodverification. But there is no compromise possiblewhen safety is of the essence.

The impeller of an axial impulse fan (Fig 6) is acompletely welded structure and its dynamicalbehaviour can be compared to that ofmeans, that there is one very dominant base modewhich even with a strong punch can be excited,leading to an audible tone.

Figure 6. Typical impeller of a mixed flow fan with weldedblade to hub connection.

And there are the harmonics where packages ofblades together are moving contrary to otherpackages. Although this fan type is commonlyknown as robust and easy to handle, care has to betaken to avoid any kind of excitation, which maylead to fatigue cracks, Section 5 describes impulseimpellers in greater detail. Blade flutter is a seriousconcern which has to be addressed by the bladeengineer as well. This aero-elastic phenomenon isespecially critical for long, slender, thin blades. Byaltering the blade geometry, e.g. the blade thicknessin the root portion of the blade, the blade stiffnesswill be increased and blade flutter would beminimized or avoided all together.

3.3 State of the art calculationsThe state of the art method for the determination ofblade natural frequencies is using F.E.A. simulationsoftware, which is an integrated part of modern,three dimensional mechanical design software. TheF.E.A method is an appropriate way to get numericaland animated results regarding various vibrationalmodes of the blade under investigation.

Figure 7 shows typical results of a FEAsimulation for a fan blade, where different colorsindicate the displacement at the given naturalfrequency, at various point on the blade surface.When applying the operation loads on the blade, thesame model can be used for stress determination aswell.

BladesHub

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Figure 7. Examples of first 3 natural frequency modes with andwithout the influence of centrifugal force. Red color = largestdisplacements

3.4 Safety factorsIn order to make fans cost effective, high tip speedsof 190 m/s and even higher are desired. Blademechanical design limits are defined by the materialproperties in combination with the necessary safetyfactors.

Each fan supplier may have its own philosophyregarding the required safety factors which would beapplied to the blade design.. Other important factorsthat would typically be considered and which would

influence the required safety factor include, but arenot limited to: material of construction, the qualityof surface, the ductility of the blade material, theoperation temperature of the fan and the gas quality.Typical values of safety factors against mechanicalyield point can vary from 2.0 to 3.0, but may beincreased subject to the above mentioned factors.

3.5 Manufacturing boundaries and restrictionsWhile the blade of a mixed flow fan usually isfabricated from a single thickness, formed steelplate, the manufacturing of an airfoil profile of anoverpressure fan is usually more complex.

There are numerous fabrication methods availablefor the manufacturing of impeller blades, however adetailed discussion regarding these processes wouldfall outside the scope of this paper.

The preferred method of fabrication for airfoilblades of heavy duty fans would be casting, forgingor hot forming. These casted or forged blades mustbe accurately machined after casting and variousquality checks must be performed to verify that theblades are acceptable.

4 QUALITY REQUIREMENTS

4.1 Welded partsImpellers of mixed flow fans as well as some smallaxial fan impellers have their blades directly weldedto the hub. This crucial connection should be donewith the highest care and checked with the same careafterwards.

The quality of the welding seams must bespecified and evaluated according to ISO 5817Group B or another comparable welding standard.Irregularities such as excessive peaking, linear offsetor spatters must be avoided at all cost.

Regardless of the welding standard utilized avisual inspection followed by a non-destructive(ND) test procedure should be employed todetermine any hidden failures (e.g cracks).

4.2 Casted piecesTypically larger blades are more difficult to castwithout flaws. Large casted blades, regardlesswhether they are made from aluminum, steel, or castiron should therefore always be subjected tointensive testing. The test methods are defined bythe relevant original equipment manufacturer

inspection plan may prescribe that from each castingbatch at least one blade should be tested with

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destructive methods to confirm the correctness of thematerial properties, both chemical and mechanical.

Visual inspection, dye penetrant or magneticparticle testing is the first choice here due to therelative low cost. Cavities, cracks and otherirregularities must however be ruled out byadditional ND-testing on each blade.

On thick section blades it is also worthwhile toemploy radiographic or ultrasonic testing proceduresas these methods are able to detect failures deepwithin the material structure. Figure 8 shows anexample.

Figure 8. Colored blade foot regions for zones, indicatingdifferent quality requirements (red = higher; green = lower).

For each of the sections the blade design engineerhas to find the best compromise between acceptablefailure sizes and the risk of a blade failure duringoperation.

4.3 Material selectionMine ventilation is one of the most complexapplications in terms of fan blade material selection.

In order to achieve long life-cycles, operationalconsiderations regarding gas quality (dusty orabrasive) as well as mechanical considerations haveto be made. Dust load and humidity, together withthe underground and surface temperature conditionsmay cause serious problems for the durability of theblades. At worst, the blade material will be subjectedto not only particle impacts, but also chemicalerosion may occur, by the formation of acids due tothe presence of various corrosive elements in theprocess gas.

Due to the variety of mines and the specificoperating conditions of each mine, there is nogeneral applicable wear protection method, theselection of a suitable wear protection system isdone on a case by case study, which takes intoaccount the specific details for every mine site.

This first step in selecting the appropriate surfaceprotection is in the correct selection of a blade basematerial. Aluminum, SG iron, stainless steel andcomposite materials are all proven on mining fans.

Depending on the primary threat, the basematerial should be selected and an additionalprotective coat has to be applied. For example if

there is a high load of acid-forming contentstogether with moderate amount of medium-sizedparticles in the air, stainless steel with leading edgeprotectors would usually be a good choice tomaintain the expected lifetime of the impellerblades.

However if this crucial step is misinterpreted bythe fan engineer, or in-sufficient details from the enduser were received, the wear protection might notselected correctly the predicted blade lifetime will bedramatically reduced. Even small defects, which aretolerable during the quality assurance process, willvery quickly develop into larger cracks or cratersand the blade may fail prematurely.

4.4 Operational experience

4.4.1 Mixed flow fansIn Section 3.2 it was mentioned that the operatingconditions and the built-in situation has to beconsidered in order to avoid damages. At lowvolume flows when the inlet guide vanes are nearlyclosed, mixed flow fans can deliver only lowpressure. The stall line is low and during start-up orpart load conditions the operating points may crossthat line (Fig 9a).

The impeller will subject to severe resonancesduring this time, and continued operation within thestall zone may lead to premature impeller failure,concentrated at the welds between the impeller bladeand hub, due to fatigue cracks caused by theresonance.

The installation of a performance curvestabilization ring can raise the stall line at low flowvolumes and avoid such dangerous difficulties (Fig9b).

Figure 9. Comparison of mixed flow fan without (9a) and with(9b) performance curve stabilization.

As shown by Maddox (1991) there are importantfeatures of welds in relation to fatigue. Thedifference between welded and unwelded materialsis significant in respect of endurance and fatiguelimits. Maddox (1991) showed that for a BS 4360

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Grade 50B steel specimen the stress range for 106

cycles is approx. 300 MPa. When at an equivalentspecimen additional material is welded from bothsides, the stress range drops down to 100 MPa. At107 cycles the ratio becomes even worse, namely270/45.

Consequently the quality acceptance have to beof the highest grade for all welds, however thepresence of micro-cracks cannot be avoided.Therefore procedures for the fatigue design ofwelded structures, e.g. according to the Europeanstandard EN 1993-1-9 (2005) or equivalent are notnegotiable.

When mixed flow fans are operated with variablespeed drives, the possibility that structural, weldedelements of the impeller fall into a resonance zone isrealistic. Even when the loads at resonance speedsmight be low, the number of cycles will reachmillions within some days. The addition of loadcycles is like adding something to a reservoir whichhave limited capacity, eventually it will be full andfailure cannot be avoided.

Alternatively these fans can be fitted with inletguide vanes to modulate the airflow at constantdriving speed.

4.4.2 Overpressure fansThe general arrangement for this type of fan is tohave individually mounted airfoil blades which arefixed to a central impeller hub. Duty modulation andenergy efficiency would be optimized by blade pitchadjustments. Blade pitch adjustments alter the angleof attack of the blades and thus the airflow can bemodulated. Blade pitch adjustments are available inmanual or automatic versions. A manual methodwould mean an artisan will have to stop the fan,adjust all the blades individually and then restart thesystem, while an automatic system would adjust theblades all at once while the fan is online.

Further optimization by using variablespeed\frequency drives is possible as well, howeverthese variable speed\frequency drives are stillrelatively expensive and it is a matter of weighing upinvestment expenditure versus operationalexpenditure.

Reversal of airflow is also possible with bladeadjustments\reversal so that at least 70% of the

flow in the reverse direction can beachieved.

The advantage of a constant speed drive isevident in respect of rotor dynamics. The risk to runthrough an aerodynamic unstable condition (stall) isrelatively low, because the characteristic curve issimilar to Figure 9b.

In-flight variable pitch fans can be started withclosed blades. This will ensure that the starting

torque is kept on the lowest possible level and theflow is started smoothly.

For less demanding requirements, simpler fanswith blades adjustable at rest are chosen. Auxiliarymining fans are usually constructed using thisprinciple. Designed for onerous and robustconditions they can be used where flow control orenergy efficiency are not the dominating factors.

5 CONCLUSION

There is no other part of a fan which is of similarimportance as its blade. It is the aerodynamicperformance which the buyer of a fan is specifyingand where his primary focus is on. He will expectthat the fan supplier did his utmost to diminish thefailure probability of his rotating machinery to anabsolute minimum.

specification complies with both the aerodynamicaland mechanical characteristics of the selected fantype.In this context, the article aimed to provide aninsight into how loads, design rules and qualityaspects are interwoven in a typical fan design.Following a well-proven engineering approach, willensure a safe and reliable product, free of failuresand with the longest possible service life.

6 REFERENCES

Eckert, B. & Schnell, E. 1980. Axial und Radial-kompressoren.Berlin: Springer-Verlag

European Committee for Standardisation 2005. EN 1993-1-9:Eurocode 3: Design of steel structures - Part 1-9 Fatigue

Maddox, S.J. 1991. Fatigue Strength of Welded Structures. 2ndEdition. Cambridge: Woodhead Publishing Ltd.

Staiger, M. et al. 1991. Beanspruchung der Laufschaufeln vonAxialventilatoren bei gestörter Zuströmung: 125-145. VDI-Berichte Nr. 872. Düsseldorf: VDI-Verlag

Traupel, W. 1982. Thermische Turbomaschinen: 405-409.Band II, 3.Auflage. Berlin: Springer-Verlag