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The high cost of power has made manufacturers look into equipment to reduce power consumption. Cooling Towers, heat exchanger, humidification and ventilator are areas where a major saving is possible at reasonable cost for manufacturers.

Wind Water Industries dynamically designed energy efficient FRP Hollow axial flow fans have helped in 20% reduction in power consumption and recovering the cost of fan in short time.

Traditionally Cast aluminum fans consume more power due to low efficiency and heavy weight. The fans operate throughout 24 hours using excess power, which can be saved by using Wind Water Industries energy efficient FRP Hollow bladed fans.

Wind Water Industries has designed and developed structurally strong FRP hollow fans which are light-in-weight with help of USA based NACA* computerized software, having more lift and low drag airfoil shape, Engineered twist, optimized chamber to increase the fan efficiency resulting a nominal low operating cost.

Wind Water Industries method of Axial Flow Fan design is based on data obtained from the combination of the blade element theory, model fan tests in a wind tunnel, and tests of full scale fans in the field. The data are plotted as curves for fans of standard theory form, making the actual operation in designing the fan very short and simple. For the analysis or design of special blade which is not conforming to the standards, the modified blade element theory is used, with airfoil section characteristics, which give resultant powers and efficiencies checking the standard model data.

Application of Axial Flow Fans

The axial flow fans are widely used for providing the required airflow for the heat & mass transfer operations in various industrial equipment and processes. These includes cooling tower for air conditioning & ventilation, humidifiers in textile mills, air–heat exchangers for various chemical processes, ventilation & exhaust as in mining industry etc. All major industries use large number of axial flow fans operation, such as:

  • Cooling Towers
  • Heat Exchangers
  • Humidifiers
  • Ventilation
  • Industrial Air Circulator
  • Man Cooler
  • All type of Industrial Axial Flow Fan


  • Energy efficient FRP Axial Flow Hollow fan Assemblies
  • Aluminum Axial Flow Hollow fan Assemblies
  • FRP fan Stack
  • FRP Inlet Bell Mouth For ACC
  • Drive Shaft
  • Gear Box

Range of FRP Axial Flow Fans

  • Fan diameter : 250mm to 15000mm
  • Air Flow : 0.5 M3/s to 2500m3/s
  • Pressure : 1mm to 100mm
  • Tip speed Limit : up to 70 M/s

Range of Aluminum Axial Flow Fans

  • Fan diameter : 250mm to 5500mm
  • Air Flow : 0.5 M3/s to 400m3/s
  • Pressure : 1mm to 100mm
  • Tip speed Limit : up to 100 M/s

  • Product Efficiency of Fans designed by Aerotech : 90%
  • Raw Materials for Blades : Epoxy / As per Client requirement
  • Raw Materials for Hub : Steel + Galvanized /SS/ As per Client requirement
  • Raw Materials for Hardware : SS-304 / SS-316/ As per Client requirement


Corrosion Free Material : The raw material i.e. Fiber Reinforced Plastics (FRP) provides the desired non-corrosive quality to the fan blades, resulting in the operation of the fans even in the chemical environment.

Lightweight : FRP is Lightweight material which ensures a low moment of inertia, minimum wear & tears on motor, bearing and drive system. Hollow FRP blades is lighter than aluminum fan assembly which makes it easy to carry and due to light weight there is no possibility of damage to fan & drive during sudden stops.

Tailor Made Designs : Composite structural design can be tailor made by using various glass fibers (Glass cloth & woven roving mat and roving) in right direction while moulding the fan blades with Epoxy resin thus imparting the desired mechanical strength by improving its industrial stability and enhanced mechanical properties.

Wind Water Industries design of the fan blades ensure more airflow, lower noise level & less power consumption means high efficiency. The Wind Water Industries DYNAMICALLY designed fan impeller of the fans, fabricated by composite material can be excellent alternative to ensure enhanced efficiency & appropriate energy saving apart from wide gamut of critical advantages.

USP of Wind Water Industries FRP Fans

HOLLOW FAN BLADES : The hollow design of fan blades The basic purpose of a “fan” is to move a mass of gas or vapor at the desired velocity. For achieving this objective there is a slight increase in the gas pressure across the fan rotor or impeller. However main aim remains to move air or gas without any appreciable increase in its pressure. The total pressure developed by the fan is of the order of a few millimeter of water gauge. A “blower” which is also referred to as “fan” in some literature deliver the gas or air with the appreciable race in pressure to overcome some kind of resistance in the flow. Some applications develop pressure of the order or 1000 mm W.G. or more. An axial flow fan stage in its simplest form consists of a rotor made up of number of blades fitted to the hub. When it is rotated by an electric motor or any other drive, a flow is established through the rotor. The actions of the rotor causes an increase in the stagnation pressure of air or gas across it. A cylindrical casing encloses the rotor.

Wind Water Industries DYNAMIC DESIGN OF FRP HOLLOW BLADE :As mentioned previously the choice of correct twist and of special airfoil sections to reduce compressibility losses are of major importance in modern fan blades design. However other considerations still remain to be studied with care if the efficiency of the blade is to be kept at its maximum under severe operating conditions.

CENTRIFUGAL AND Wind Water Industries DYNAMIC TWISTING : In variable-pitch fan as described later the blades are turned in the hub about their longitudinal or pitch-change axis. Clearly the mechanism provided to produce this pitch change must be capable of exerting sufficient force to overcome any mechanical or Wind Water Industries dynamic opposing force set up by the blades themselves. It will be useful then to see exactly what these force are.

The “mechanical force” involved is known as Centrifugal Twisting Moment (C.T.M)which as its name implies is closely allied to the normal centrifugal force acting on the blades when the fan is rotating about the shaft . It is turning couple brought about by the fact that the blade section are inclined at an angle to the fan of rotation and results in the natural tendency for any fan blade, when rotating turned about its longitudinal axis towards zero pitch so that the blades section are turned in to the fan blade rotation .

THRUST AND TORQUE FORCES :The main factor that determines the force developed by an aerofoil section is the angle at which it is inclined to the relative airflow, i.e. the angle of attack in order, therefore, for the fan blade section to develop the requisite aerodynamic force for propulsion, each section along the blade must be inclined at the appropriate angle of attack to the relative airflow direction pertaining to the section hence knowing angle of attack is a simple mater to indicate the force developed as shown in figure 1) from which it will be noted that the total blade angle of attack and the helix angle.

Why Wind Wand Industries

Why to choose Wind Water Industries FRP Fan’s

  • 2 years Guarantee for all fan assemblies from installation period.
  • Power Saving guarantee of upto 30% without reduction of air flow.
  • Higher Durability of fan blades. The blades will easily last for 10 years minimum as we use top grade material, which is superior quality of resin (Epoxy) designed to with stand harshest condition for cooling tower.
  • Wind Water Industries maintains all the fan tip clearance with the ratio of 0.03% but it is variation depend to improper round shape of the fan cylinder.
  • Our fans maintain the track variation within range of less than 25mm for large size of the fans as per recommended international standard.
  • Longer life for fan as we maintain special hard build on leading edge providing Erosion resistance.
  • All fans are built with UV protected and corrosion free material.
  • Tailor made design possible due to special moulds which helps in producing any diameter as per clients requirement.
  • Our fan assembly not only saves power it has low maintenance cost, inter changeable parts and gives less load on gear box which helps in extending life of gear box.
  • Our typical balancing method like momentum balancing of individual blades thereby maintaining equal centre-of-gravity of all fan blades and giving momentum matching mark for easy installation in hub. Hub is dynamically balanced to ensure minimum vibration (Grade of Balancing G-6.3)
  • Manually adjustable pitch, Steel galvanized hub with FRP hollow blade from 2’Ø to 50’Ø. Adjustable power setting corresponding to climatic condition.
  • Our ideal elliptic inlet type velocity recovery-cone FRP fan cylinder, resonance stable, reduces HP, and increase air moment-an ideal mate.
  • Our fans resist moisture. Dripping water and repeated condensation from a saturated atmosphere are to be regarded as abnormal hazards. Wind Water Industries provides higher grade materials and finishes which resist corrosion. The impellers are constructed from materials selected to resist erosion, and pitting are avoid from the points of impacts with the liquid drops.
  • Resist's Corrosion. Heavily polluted industrial atmospheres will be sufficiently corrosive to shorten the life of most general- purpose fans. Wind Water Industries provides the concentration which remains within what might be described as the “breathable” range, additional coats of protective paints will usually suffice. Chlorinated rubber paints are commonly used for this purpose. Wind Water Industries Fans handle highly corrosive gases, particularly when wet or hot, may be require to be constructed from special corrosion-resistant of FRP raw materials.
  • Heat resistance : Wind Water Industries fans are suitable for operating continuously in normal indoor and outdoor atmospheres at temperatures up to 40oC~120 oC.
  • Abrasion resistance : The impellers of general-purpose fans may be worn away by the passage of too much dust of an abrasive character. If the particles cannot be sufficiently removed, as by a cyclone, an abrasion resistance fan must be specified. FRP fan blades moulded to aerofoil shape used for axial fans, and will deal with particles of any hardness.


Wind Water Industries Quality Management Principles

To lead and operate our organization successfully, it is necessary that we direct and control it in a systematic and transparent manner. Success can result from implementing and maintaining a management system that is designed to continually improve performance while addressing the needs of all interested parties. Managing an organization encompasses quality management amongst other management disciplines

Few quality management principles have been identified by Aerotech that can be used by top management in order to lead our organization towards improved our performance.

Customer focus : Our Organization depends on our customers and therefore we understand current and future customer needs, we meet customer requirements and strive to exceed customer expectations.

Leadership : Wind Water Industries unity of purpose and direction of the organization. We create and maintain the internal environment in which people can become fully involved in achieving the organization’s objectives.

Involvement of people : The people at all levels are the essence of our organization and their fully involvement enables their abilities to be used for the organization’s benefits.

Process approach : desired result is achieved more efficiently when activities and related resources are managed as a process.

System approach to management : We identify, understand and manage interrelated processes as a system contributes to the organization’s effectiveness and efficiency in achieving its objectives.

Continual improvement : Continual improvement of the organization’s overall performance is a permanent objective of the organization.

Factual approach to decision making : Effective decisions are based on the analysis of data and information.

Mutually beneficial supplier relationships : Our organization and our suppliers are independent and a mutually beneficial relationship enhances the ability of both to create value.

These eight quality management principles form the basis for the quality management system standards within the ISO 9001:2008 and ISO 14001:2004 family.

Pressure Drop

Estimation of Total Pressure Drop

While the resistance of a system can be estimated in terms of static pressure changes, the total pressure method is adopted here. This avoids the use of the somewhat confusing “static regain” concept, and is logical in the total pressure drop is true measure of energy loss. When using total pressure it is important to remember that the loss of total pressure at system outlet must always be included. This appliers whether the air leaves through an outlet grille into a room, leaves through whether louvers to the atmospheres or is directly discharged by the fan itself. In each case there is an item of total pressure drop equal to the velocity pressure corresponding to the average velocity to an outlet.

The loss in each elements of the systems are depend to the average velocity through it, which is taken as:

Average Velocity V (m/s) =  Volume flow, Q
                                               Gross cross section

From this velocity, and the air density, p (kg/m3) the conventional velocity pressure is determined:

Velocity Pressure = ½ p V2 Pa.

This is then multiplied by a factor, K, and the result is the total pressure drop from the inlet (1) to the outlet (2) of the element. Since we are calling it a pressure drop, we can ignore the negative sign which it would have as a pressure change (from high low in the direction of flow).

Total Pressure drop Pf12 = K ½ p V12

Or Pf12 = K ½ p V22

Influence of a Density on Total Pressure Drop

To calculate the total pressure drops in an airways system, Pf first on the assumption that it is handling air at 1.20 kg/m3 standard density. This will give the correct volume flow rate irrespective of the actual density (or gas) when used in conjunction with the standard fan characteristic. The reason is that both the fans total pressure and total pressure drop are equally in balance as density changes.

At any density, pkg/m3, the total pressure drop is: 

Pf = P * (Pf at 1.2 kg/m3)


Design & Testing


  • Fan diameter: in meter
  • P= pa of water gauge
  • P = p g h = N/ m2
  • Area =Π /4 x (d) 2 = m2
  • Total load acting on fan = P x A = N
  • Load acting on each blade = kg
  • Taking Factor of safety =1.5
  • So, Actual applied Load = Kg
  • Time of load application for each blade =30 min
  • Evaluate the creep to avoid the failure

Blade Mechanical Design

Tensile strength of the blade material of frp = 50 kg/cm/2 F 
                                                                                             w x t
Cross breaking strength = 1.5 W L
                                                  B D2 
W= Fc
Fc= m x ω2 x r = N
ω = (2 x Π x N)/(60) = rad / sec

Shearing strength of blade materials = W 
Bond strength = W 
Percentage of liner shrinkage = {( L0 - L )/ L0}x 100
Flexural strength = 3 x P x L/2 x 2x b x d2
Flexural Rigidity = D =Es bt3 +2b+d2 +Ecbc3
                                              6                       12
Modules of elasticity EB= (L3 M) / (4 B D3 )


The fan deflection testing is carried out to ensure the quality as follows

  • To load the FRP fan blade till failure, in the proportion 20% of load at a cross section close to neck of the blade, 30% of load at cross-section approximately at the middle and 50% of load at a cross-section close to the free end.
  • To record the load and deflection at the three loading locations.
  • To observe any abnormal behavior during the load test and observe the failure pattern.

The fan deflection test is carried out for the predetermined load as follows

  • The FRP fan blade is fixed on a rigid platform with the help of holding clamps.
  • Three loading positions are marked for the load runners. The outer most position is at a distance of d1 from the tip (free end) and takes 50% of superposed load. The innermost position is at a distance of d2 mm from tip and takes 20 % of superposed load schematic view of loading positions. The deflection measuring points are same made using iron runners and matching the aerofoil profile at the respective loading points. Pre-calibrated dead weights are used for the test. Three suitable weighing pans are fabricated in-house and weighed. It was found that the total weight of the three pans came out to be about W1 Kg. Hence, the weighing pans were not used initially; instead hangers were used till the total load of about W Kg.
  • Initial deflection readings with only self-weight of the blade and the holding clamp with load runners were noted down. The deflections were measured with the help of pointers pointing on to a wooden meter scale clamped to a fix support. The hangers were hooked to the three loading clamps respectively. The loading was done in steps of w Kg. Therefore, the load at loading position W1 was increased in steps w1Kg. The load at loading position W2 was increased in steps of w2. and the load at the loading positions W1, W2 and W3 were after each loading steps were noted. This loading system was continued up-to W Kg. Then the load hangers were removed and weighing pans were hung at three loading locations. Subsequently, dead weights were added to the weighing pans in step and readings for loads and deflections were recorded till failure.


Following observations are made on the basis of readings achieved

  • Initial zero readings was taken with self-weight of the blade and a superposed load of holding clamps with wooden load runners.
  • The creep effect was studied at the superposed load in Kg. kept for about 30 minute. After 30 minutes, it was observed that the loading position W1 became a creep in mm, the loading position W2 also became a creep effected.
  • Cross checking up-to failure de-lamination of the chord joint (edge of the blade near the innermost loading positions) was observed along the longitudinal axis of the blade over a length of about X mm. The total failure load was observed in Kg.
  • Maximum deflection observed by the outermost loading point in mm at the failure load.
  • Load-deflection curve expectedly shows a typical curve of a FRP specimen under a flexural load.
  • The blade was cut at the three loading locations to observe the thickness of cross-section of chord of the blade across the cut sections with identification numbers.


The ultimate superposed load on the given sample of FRP blade was determined in Kg. No abnormal behavior was stopped for a period of 30 minutes. This however, is natural for FRP material. The maximum deflection at the outermost loading location was found in mm. After failure when the superposed loads were removed, the blade almost regained its original position, showing that even at failure the FRP material remained elastic. This is because, the failure was entirely due to de-lamination of FRP layers in the vicity of the loading point W1.

Shear stress and Bending moment in Fan

(E.g. Diagram of 6’F Fan at 20 mm wg of total Pressure)

Hub design

The material used for manufacturing the hub is mild steel/ Stainless steel. The following are the properties of mild steel

  • Tensile stress = N/mm2
  • Shear stress = N/mm2

  • The area of hub under the shear stress=? x d x t = mm2
  • Therefore the cross breaking shear strength of the hub N = Area under shear stress x shear stress
  • Total shear load acting on the hub (N) = No. of blades x centrifugal force on each blade+ Thrust load acting on all the blades
  • The shear strength of the hub must be greater than the total shear load acting on the hub.

Design of U bolts

  • The area of u bolts under the shear load = 4 xΠ/4 x d2 = mm2
  • Therefore the cross breaking shear strength of the u bolt = A x Fs The total shear load acting on each pair of u bolt (N) = centrifugal force acting on each blade + thrust on each blade.
  • The shear force acting on the I - bolts (N) = centrifugal force due to the hub and blades + thrust force acting on the bolts.
  • The shear strength of the u bolts must be higher than the shear load acting on the U-bolts.

Impeller Design

Practical Impeller Design

This idealized treatment is only a starting point. Further factors we consider while designing include the following

  • The air will leave the impeller at the angle set by the trailing edge of the blade. There will be a deviation angle tending to reduce input and output power. This is a function of aerofoil shape and blade loading.
  • The air will meet the leading edge of the blade at the optimum angle only at the design volume flow. At other flows the incidence angle will change. The effect of this change on performance over the whole fan characteristics is again a function of aerofoil shape and loading.
  • To keep the portion of flow considered between the imaginary surfaces defining its intended path requires a balancing transverse pressure gradient. In the axial case, for example, only the so-called free vortex pressure and velocity distribution will secure flow at constant radius along concentric cylinders. This distribution requires a constant values of rv cos a, i.e. a spin components which is smaller the bigger the radius. This limits the workdone towards the blade tips, giving a fan of poor power-size ratio; To overcome this weakness, forced vortex designs are used, increasing tip work. To maintain balance the streamlines tilt outwards through the impeller, so that r2and u2 are greater than r1 and u1.
  • Viscous drag forces at the blade surface and wake effects behind the blade convert some of the work input into heat instead of useful pressure rise
  • Tip clearance effects and boundary layer retorted flow along casing hub, back plate spoils the flow pattern at the ends of the blade and limits the work done (input and output).
  • The velocity leaving the blades is usually far from uniform in magnitude or direction. Since the energy is proportional to the square of the velocity, more is required by the peak velocities than is saved in the troughs. Thus excess kinetic energy is supplied which not all will be available when the air has reached the downstream test plane.

On account of the considerable variation in the flow conditions and the blade section along the span, it is divided in to a number of infinitesimal section of small radial thickness dr. the flow through such a section is assumed to be independent of the flow through other elements.

Velocities is a blade force for the flow through the elemental section are shown below. The flow has a mean velocity w and direction b (from the axial direction). The lift force L is normal of mean flow and the drag D parallel to this. The axial (#Fx) and tangential (#Fy ) force acting on the element. Also (#Fr) is the resultant force inclined at an angle 0 to the direction of lift.

An expression for the pressure rise (#P) across the element is now developed.

Resolving the force in the axial and tangential direction

  • #Fx = #L sinb - #D cosb
  • #Fy = #L cosb + #D sin b

By definition lift and drag force are

  • #L = ½ CL p w2 (ldr)
  • #D = ½ CD p w2 (ldr)
  • tanf = #D/#L = CD/CL


Since the fan laws are valid for any particular point on the fan pressure/volume characteristic, similar laws will be valid for every other point of operation, the only difference being numerical values of the coefficients. Thus a plot of Kp against Kq will have exactly same form as the pressure /volume characteristic of each fan in homologous series, and also be used to compare the performance of the series design with that of another series design. A system of performance coefficients based on this

Volume coefficient.    Kq                        Q
 = ----------------------------
               d3 (n / 1000) 
Pressure coefficient.    Kr          Fan total pressure Pt 
 = -------------------------------
             d2 (n / 1000)2
Fan power efficient.    Kp                Fan h.p.
 = -------------------------------
             d5 (n / 1000)3
Volume coefficient   f              Volume flow
 = --------------------------------
               (Pd2/ 4)u
Total pressure coefficient   y               Fan total pressure
 = -----------------------------
Static pressure coefficient   yst              Fan static pressure 
 = ------------------------------
Power coefficient.    l                       fy
 = -------------------------------

Application of Fan Engineering Formula

The application of the fan theory can best shown by example.

D    =      ÖQ/k
Hub Dia.    =   Tpf/n x 10
RPM    =   (U)/(p x d x 60)
Stp    =   0.5 x p x ts2 (1- f αtan2θ)
Tip Speed    =   p x D x n
a    =   U/cx
Tp    =   Stp + Pv
Fan Velocity pressure (Vp)    =   0.5 x r x u2
Required Impeller pressure    =   Tp/h
Impeller Power    =   Fp/Gh/Mh
Wind Water Industries dynamic efficiency    =   Ap/Ip
Tip Clearance    =   D x 0.025 x 25.4
Mass flow rate    =   r x Q
Natural frequency    =   1/2p Ög´ 1000 
Blade pass Frequency    =   (n x Z )/60
Frequency of natural vibration    =   1Ög
     2*p d
Solid ratio    =   ( Z x C)/2 pr
Lift force    =   0.5 x p x V 2 x S x CL
Drag force    =   0.5 x p x V 2 x S x CD
Axial thrust    =   5.202 x D r x NFA2 x W
Centrifugal force    =   2 x p x N_
        60   x 2 x x M `
Reynolds number    =   DVr 
Fan efficiency    =     Q x tp       x (100)
      kW x 1000

Fan Laws

The performance of a fan in terms of pressure, volume flow and power absorbed depends on a number of factors, the most obvious of which are

  • The design and type of fan
  • The point of operation on the volume flow/pressure characteristic
  • The size of fan
  • The speed of rotation of impeller
  • The condition of the air or gas passing through the fan.

It is customary for a manufacturer to make a range of fans of varying size to single design, thus producing a series of geometrically similar fans (homologous series). It is convenient to be able to compute the performance of each fan from the minimum test data. The pressure/volume flow relationship it is not generally capable of being expressed as a simple mathematical function. However, by considering any single point of operation on the characteristic curve (for example, the point at which the fan efficiency is a maximum), it is possible to drive some simple relationships, generally known as the fan laws.

For all fan laws: hTa = hTb and (point of rating)a = (point of rating)b

Fan Noise

Noise is undesirable or unwanted sound. With better understanding of the effects of environment on the inmates of the dwellings and factory workers, noise has become an important subject in the design, installation, and operation of the fans.

In a well-balanced properly installed fan, the mechanical noise originating from bearing and vibration of various parts is not as prominent as the wind water industries dynamically generated noise. The latter is due to the various flow phenomena occurring within the fan.

The main causes of wind water industries dynamically generated noise are

  • The flow at entry and exit of the fan. i.e., suction and exhaust noise
  • Rotation of blade through air or gas
  • Turbulence of air
  • Shedding and vortices, from blades
  • Separation, stalling and surging

Some parameters on which the noise level radiated from a fan depends are: fan Wind Water Industries dynamic performance, duct configurations at the entry and exit, housing geometry, relative number of blades, magnitudes of clearances, blade thickness and fan speed.

Some methods of reducing fan noise are

  • Operation of fans at their maximum efficiencies
  • Use of low speed and low pressure fans
  • Employment of uniform flow in ducts
  • Use flexible fan mounting
  • Use of sound absorbing walls; duct should be lined by sound absorbing material
  • Use of silencers at the suction and exhaust
  • Reinforcing fan casings.

Fan Laws for Sound

Dimension analysis indicate that the sound power ratio SWR = SW/W of geometrically similar series of fans should be dependent only on Mach number, Ma = V/Va, (ratio of a typical fan velocity to the velocity of sound) and Reynolds number, Re = DVp m which is the ratio of the dynamic to viscous forces involved, SWR = constant x (Ma)xx (Re)y

Taking the rotational speed N, diameter D, and output power W of the impeller as the typical quantities

  • W µ N3D5 Maµ ND Re µ ND2
  • Therefore SW = constant x N3 + x + y D5 + x +2y
  • Acoustical theory, considering only the Mach number effect, yields the following relations for the source types mentioned
  • Monopole Source x = 1 so that SW varies as V4
  • Dipole Source x = 3 so that SW varies as V6
  • Quadrupole Source x = 3 so that SW varies as V8
  • Fan noise is principally dipole in origin and, for the simplest cases of boundary layer separation, there is reason to believe that the Reynolds number exponents y should be about - 0.4, leading to
  • SW = constant x N5.6 D7.2
  • Axial fans seems to follow this relationship quite closely (e.g. N5.5 D7.5), but other type may well introduce more complex generation mechanism. For example, centrifugal fans are usually considered to follow the law:
  • SW = constant x N5 D7
  • Geometrical similarity as its affects Reynolds number is not well maintained in practice from size to size (thickness, clearances, even number of blades). It is wise to keep the index of D2 greater than the experimental index of N and to treat small changes with size as a scale effect on the constant. The influence of atmospheric conditions is small -generally within 1 dB.

Erosion of Fan

Minor erosion of fan parts due to the presence of dust is quite common. However, in some applications erosion of fan blades and casing due to dust-laden air is very serious. This is one of the causes of failure of I.D fans.

When dust particles directly hit the moving blades, they cause cracking of the blades, whereas flow of abrasive dust through the passages causing scraping action leading to surface erosion. Some aspects of dust erosions are given below.

  • The worn-out blade surfaces alter the geometry of the flow far from the design. This is reflected in poor fan performance.
  • If considerable erosion has occurred in highly stressed regions, the affected part can be suddenly fail after some time.
  • The wear of the rotor due to dust erosion is not ax symmetric. This leads to imbalance of the rotor and increases the load on bearings.
  • The imbalance and the resulting vibration are further increased due to the collection of the dust in the pockets created by erosion.

Dust particles collected in the stalled regions of the fan which erodes the surface by a milling action.

In view of erosion problems, the selection of the right type of fan is important. However a fan which suffers least due to erosion may not always be the best choice for given application. Dust erosion has been found to be inversely proportional to the pressure coefficient. It has found that erosion is more serious in axial type fan.

In view of erosion problems, the selection of the right type of fan is important. However a fan which suffers least due to erosion may not always be the best choice for given application. Dust erosion has been found to be inversely proportional to the pressure coefficient. It has found that erosion is more serious in axial type fan.

Wind Water Industries can minimize the erosion problem which is

  • Employing more efficient dust removing apparatus.
  • Regulating fan speeds at part load.
  • Employing large and low speed fan.
  • Providing erosion shields on the blades.