Design & Analysis of Fan

September 10, 2017 | Author: Mrudula Chintana | Category: Mechanical Fan, Boiler, Gas Compressor, Computer Aided Design, Turbine
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DESIGN AND ANALYSIS OF A LOW SPECIFIC SPEED CENTRIFUGAL FAN

A Dissertation Work Submitted to Jawaharlal Nehru Technological University In Partial Fulfilment of the requirements of the award of BACHELOR OF TECHNOLOGY IN MECHANICAL ENGINEERING By K.KARTHIK A.SAI CHARAN D.ASHESH GOPAL NATH A.SREENU

06311A0368 06311A0380 06311A03B4 06311A03B5

Department of Mechanical Engineering, SREE NIDHI INSTITUTE OF SCIENCE & TECHNOLOGY Yamnampet, Ghatkesar, Hyderabad-501301. (Accredited by AICTE, New Delhi & Affiliated to JNT University, Hyderabad)

DESIGN AND ANALYSIS OF A LOW SPECIFIC SPEED CENTRIFUGAL FAN

A Dissertation Work Submitted to Jawaharlal Nehru Technological University In Partial Fulfilment of the requirements of the award of BACHELOR OF TECHNOLOGY IN MECHANICAL ENGINEERING By K.KARTHIK A.SAI CHARAN D.ASHESH GOPAL NATH A.SREENU

06311A0368 06311A0380 06311A03B4 06311A03B5

Under The Guidance of Dr.M.V.S.S.S.M.PRASAD B.Tech(IITM), M.Tech(IITM), Ph.D(IITM) Professor, Department of Mechanical Engineering

Department of Mechanical Engineering, SREE NIDHI INSTITUTE OF SCIENCE & TECHNOLOGY Yamnampet, Ghatkesar, Hyderabad-501301. (Accredited by AICTE, New Delhi & Affiliated to JNT University, Hyderabad)

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ACKNOWLEDGEMENT We would take immense pleasure to acknowledge with gratitude, the help & support extended during the course of our project entitled DESIGN AND ANALYSIS OF A LOW SPEED CENTRIFUGAL FAN from all people who have helped in the successful completion of this project.

We are highly indebted to Dr. M.V.S.S.S.M.PRASAD, Professor, Department of Mechanical Engineering, for his guidance and help at all stages of the project.

We are highly grateful to Dr. Ch.SIVA REDDY, Professor, Head of Department of Mechanical Engineering for the facilities provided to carry out the project.

We are highly thankful to Mr. RAVINDER REDDY, Assistant professor, Department of Mechanical Engineering for helping us in learning the software required for this project.

We express our sincere thanks to Mr. VENKAT NARAYANA, incharge of CAD/CAM laboratory for providing us the computer systems and the required software tools.

We also thank our parents, class mates and friends for the kind support given by them at all stages of the project.

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ABSTRACT :

The current project is aimed to design a low specific speed centrifugal fan. Fans belong to the family of turbo machines and they move air or gas continuously at desired velocity by action of a rotor. Flow investigation of the fan is planned to be carried out by using ANSYS-CFX software for different designed off design points of operation. The performance of the fan generated from the CFD analysis at the design point will be compared with that of the designed data assumed for calculation. This will also be compared with the best efficiency point of operation. For the analysis, an Auto CAD drawing and a 3-D model the fan impeller and casing are developed for the designed fan. This is followed by the generation of Grid and aerodynamic analysis using the available CFD solver. The work is concluded by identifying possible zones of improvements in the design of impeller and casing and suggest suitable modifications.

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Nomenclature, Greek letters and Subscripts: A

Area

b

Impeller Width

c

Absolute velocity

dP

Incremental change in pressure

d

Diameter

D

Impeller diameter

E

Energy

H

Head, blade span or height

m

Mass flow rate

n

Speed in rpm

nsh

Shape number

nq

Specific speed

P

Pressure

p

Slip power Factor

R

Gas constant

r

Radius

Rc

Radius of curvature of vane

u

Blade speed

W

Specific work

z

Number of blades 5

GREEK LETTERS:

a Nozzle blade angle w.r.t. Blade speed u ν

Taper angle at shroud

β Impeller blade angle,relative,flow direction w.r.t. Negative of blade speed Φ

Flow coefficient

η

Efficiency

ρ

Density

w

Angular Velocity

Pressure coefficient , Energy coefficient SUBSCRIPT





Far upstream or Downstream Flow conditions with infinite number of blades or vane congruent flow

bl

Blade or Impeller

b h m t u

Blade or Vane Hydraulic Meridional Tip Tangential or Peripheral component

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CONTENTS

1 INTRODUCTION……………………………………………

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1.1 Introduction to Turbo machines 1.2 Fans – Principle of operation. 1.3 Classification of fans.

2 LITERATURE SURVEY……………………………………. 2.1 Specific work and static pressure rise 2.2 Impeller 2.2.1 Slip 2.2.2 Inlet Vane angle 2.2.3 Pre whirl 2.2.4 Impeller outlet angle 2.2.5 Impeller outlet diameter 2.2.6 Effect of Viscosity 2.2.7

Inlet passage

2.2.8

Effect of surface roughness

2.2.9

Volute casing

2.3 Effects of geometric and flow parameters of fan 2.3.1 Impeller size 2.3.2 Blade shape 2.3.3 Number of blades 2.3.4 Volute and Diffuser 2.3.5 Effect of Friction 2.4 Losses 2.4.1 Losses in the impeller 2.4.2 Leakage losses 2.4.3 Volute and diffuser losses

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2.5 Applications

3 DESIGN OF THE LOW SPECIFIC SPEED CENTRIFUGAL FAN ……

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3.1 Fan Specifications. 3.2 Calculations 3.3 Auto CAD design of the Fan Impeller.

4 EXTRACTION OF COORDINATES…………………………….

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4.1 Method of extraction 4.2 Coordinates of the blade profile (hub side) 4.3 Coordinates of the blade profile (shroud side) 4.4 Coordinates of the hub 4.5

Coordinates of the shroud

5 CFD THEORY……………………………………………………

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5.1 CFD Theory 5.1.1 Continuity Equation 5.1.2 Momentum Equation 5.1.3 Energy Equation 5.2 Turbulence Modules 5.2.1 K- Epsilon module 5.3 Discretization of governing equations 5.3.1 Finite difference method 5.3.2 Finite Control volume method 5.3.3 Finite element method

6 ANSYS – CFX………………………………………………….. 6.1 Introduction to ansys cfx 6.2 Ansys Cfx and the Ansys workbench Environment 6.3 CFD Pre-Processing in CFX-Pre 6.4 The ANSYS CFX Solver 6.5 Post-Processing with ANSYS CFD-Post

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6.6 Industry solutions using ANSYS

7 METHODOLOGY………………………………………………….

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7.1 Modelling and CFD analysis of centrifugal fan. 7.2 Meridional data for Hub and Shroud contour 7.3 Mesh data for 3-D impeller blades 7.4 Selection of solver parameters and convergence criteria 7.5 Blade geometry plot

8 RESULTS AND DISCUSSIONS……………………………………

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8.1 General 8.2 Variation of flow parameters in the chosen impeller 8.3 Results 8.4 Pictorial analysis 8.5 Graphs

9 CONCLUSIONS……………………………………………………….

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10 SUGGESTION………………………………………………………..

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11 REFERENCES…………………………………………………………

93

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1. INTRODUCTION 1.1 Introduction

to Turbo machines

Turbomachines used for the compression of gases are classified under radial, axial or

mixed flow types depending on the flow through the impeller. In a radial or centrifugal machine, the pressure increase due to the centrifugal action forms an important factor in its operation. The energy is transferred by dynamic means from the impeller to the fluid. The fluid because of centrifugal action is continuously thrown outwards making way for fresh fluid to be inducted in because of the reduced local pressure. Another characteristic feature of the centrifugal impeller is the angular momentum of the fluid flowing through the impeller is increased by virtue of the impeller outer diameter being significantly larger than the inlet diameter. In axial flow machines, a large mass of gas is set in motion by the rotating impeller and is made to move forward because of the aerodynamic action of the blades. A mixed flow machine encompasses the properties of both the above types. Depending on the pressure rise attained, these machines are named as fans and blower or compressors. There is however no distinct demarcation among the different types. Fans handle gases in large volumes without appreciable density variation. Pressure ratio attainable is of the order of 1.05. They are invariably single stage machines. Blowers cover pressure ratios from 1.05 to about 4. They are made

either as single

stage or two or three stages. No inter cooling is required. Compressors include pressure ratios from 3 to 12 or higher. They are invariably multistage with or without intercooling. For higher pressure ratios appreciable compression takes place followed by a reduction in volume. The calculations are done on the basis of mass flow in such cases.

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The selection of a type of impeller namely axial, radial or mixed flow for a specified pressure rise, speed and flow rate follows from shape number considerations defined by Nshape = n √(v)/ w^0.75 The shape number is important to achieve an optimum efficiency. Radial machines have low shape numbers ranging from 0.033 to 0.12 and are known as slow running impellers. Axial flow types have shape numbers from 0.33 to 1.5. Mixed flow types have values in between those of radial and axial impellers.. An idea of the shape of impeller can be obtained from the shape number. For example, slow running impellers have long and narrow vane channel passages and large shroud diameters. This increases the friction losses and lowers the efficiency, high shape numbers are desirable. The energy which is converted into pressure in the impeller is indicated by the degree of reaction which is the ratio of specific pressure energy to the specific work of the machine. Blowers and compressors operate with degree of reaction greater than zero, and mostly than 0.5. The reason is that the static pressure can be generated more efficiently in the impeller than in the guide vanes as the centrifugal forces in the rotating channels of the impeller help in the suction of the boundary layer and dead zones. If the specified pressure rise cannot be obtained in one stage, two or more stages as required are built in series, the individual stages being joined by what are known as return guide passages or return channels. In such a multistage centrifugal compressor or blower, the chief problems encountered are regarding the design of efficient guide and return channel passages as well as carefully designed shroud and vane contours. Though compressors with more than eight or ten stages are in existence, the number of stages is generally restricted to two or three. The desired pressure rise is obtained by employing high rotational speeds made possible by the steam and gas turbine drives and using high strength forged impellers with straight radial blades and devoid of front shroud in order to minimize the stresses in the hub and back shroud. In blowers and fans dealing with large volumes of gas but relatively low pressure rise, sheet metal construction is employed, with suitable hub design to take care of stresses and guide the flow. The sheets are suitably pressed to shape and the joining is through riveting or welding.

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Blade loading, shroud or disc stresses and critical speed considerations impose serious restrictions on the dimensions of the machine to lower values. However, s the pressure rise increases with increasing peripheral speeds, minimum number of stages is preferred for a compact blower, thus necessitating the use of high peripheral speeds limited by the strength of the material.

1.2 FAN : A fan can be defined as a volumetric machine, which, like a pump, moves a quantity of air or gas from one place to another. In doing this, it overcomes resistance to flow by supplying the fluid with the energy necessary for continued motion. Physically essential elements of a fan are a bladed impeller (rotor) and a housing to collect the incoming air or gas and direct its flow. Fans, Blowers or Compressors all move air, but at different pressures. At any point in the flow of air through the impeller, a pressure head obtains the centripetal acceleration, so that the static pressure of the air increases from the eye to the tip of the impeller. I

1.3 CLASSIFICATION OF FANS Depending upon the nature of the flow through the impeller blades, fans can be categorized as axial, centrifugal, mixed or cross flow type.

The major categories can be further categorized as given below: Centrifugal flow fans: a. Forward Curved b. Radial Curved c. Backward Curved Axial flow fans: a. Propeller type. b. Tube-axial type c. Contra rotating d. Guide-vane type e. Axial type 12

Mixed flow fans: a. Axial Casing Cross flow fans: a. J-Casing b. S-Casing c. U-Casing The above said fans have different characteristics suitable for specific applications. If the requirement is to blow air in large volume rate capacity, but relatively low-pressure gain, axial flow fans may be suited by contrast a fan required to blow air through filtrate system offering a high flow resistance will have a relatively small volume flow rate capacity with high pressure rise.

CENTRIFUGAL FLOW FANS

Air or gas enters the impeller of the fan axially through the suction chamber. This gas flows through the flow passage between the impeller blades while impeller rotates. The action of the impeller swings the gas from a smaller radius to a larger radius and delivers the gas at a high pressure and velocity to the casing. Due to impeller rotation centrifugal force also contributes to the stage pressure rise. At the exit of the impeller a spiral shaped casing known as scroll or volute collects the flow from impeller which can further increase the static pressure of air. Forward Curved Centrifugal Fans In forward curved centrifugal fans the blades are inclined in the direction of motion. This type of fan is best suited for application requiring high volume flow at low to medium pressure rise. This type is sometimes referred to as a ‘Volume Blower’. It can compete with tube axial and guide vane axial fans for some duties. Its efficiency is less than axial fans. Radial Discharge Centrifugal Fans This type of fan is mainly suited for handling of air borne particles. In this type of fan blades tend to be self-cleaning in moderately dirty conditions and in efficient units with curved heel blades is thus often used for draught induction in the boilers. Because of tolerance these fans are suitable for handling particulate matter in filtration duties. Back-bladed Centrifugal Fans 13

In backward curved centrifugal fans, the blades at the impeller are inclined away from the direction of motion. The static pressure rise in the rotor results from the centrifugal energy and the diffusion of the relative flow. The stagnation pressure rise and stage work depends on the whirl components (Cu , Cu ) of the absolute velocity vectors C and C respectively. 1 2 1 2 These impellers are employed for lower pressure and lower flow rates.

AXIAL FLOW FANS The major categorizes of the axial flow fans are sub-categorized into four types: Propeller Fans, Tube-Axial Fans, Contra Rotating Fans and Guide-Vane Axial Fans. Most axial fans are available with many blade angle settings that in some cases may be adjusted when stationary, by slackening a clamping mechanism in the impeller hub. The variable pitch facility is an advantage in sophisticated fans that can alter the impeller blade angle while the fan is in operation. The flow coefficient of the fan is predominantly affected by the changing of blade angles. Fans optimized to produce high flow coefficients are set with large blade angles.

MIXED FLOW FANS The characteristics of the mixed flow fans are different from those of axial flow fans and those of centrifugal fans. These fans are frequently used when characteristics approximating those of backward curved centrifugal fans are required but the installation dictates an axial inlet and outlet configuration. One most common type is axial casing mixed-flow fan.

CROSS-FLOW FANS In this type of fans the air enters the impeller through peripheral segment other than through hub. These fans are used where convenience is more important than efficiency. These fans are suitable for low-pressure rise applications. The applications of cross flow fans are domestic fan assisted heaters, handhold hair dryers and air curtain.

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2. LITERATURE SURVEY

2.1 Specific work and Static pressure rise In any centrifugal machine, the most important requirement is that it should develop the required specific work with the desired static pressure rise. In other words the specific pressure rise is directly dependent on the specific work developed by the machine .

The specific work is developed in the impeller only through the energy transfer to the fluid through the vanes and is given by Euler's equation W = U2C2 – U1C1 W= specific work developed by the stage (N.m/Kg) U1 = impeller speed at start of vane U2 = impeller tip peripheral speed C1 and C2 are the components the absolute velocity in the tangential direction at points just before the inlet to the impeller vane and the exit from the impeller vane respectively. The above Equation can be rewritten as: W = (U22 – U12 + C12-C22+W02-W32)/2 As the flow energy of the fluid comprises the pressure energy, the kinetic energy and that due to the geodetic head, the energy at any section of the passage (except where energy is being added) can be written as: E = P/ρ + C2/Z + g.h

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2.2 BLADE ANGLES: Inlet vane angle As the temperature of the air at the inlet is less. The sonic velocity is also less. There is the danger of the velocity in this region reaching a sonic value .For incompressible flow, the relative inlet velocity is a minimum when β 1 =35°. In compressible flow,

the relative inlet Mach number is a minimum when β 1 is in

between 25° to 30°.

Exit vane angle There are three considerations for β2b namely forward curved blades if β2b90°. In all the three cases β1b, the fan speed, the inlet velocity cm and size are kept the same. Therefore the velocity triangles at 1 are the same for three cases. The velocity triangles at 2 are shown in the figures for each case. It can be seen c2u increases with β2b and likewise the specific work. As β2b increases, the blades are more cambered finally resulting in the highly cambered impulse profile this means increase in the B 2b results in increase in C 2u , likewise the specific work. The kinetic energy of the fluid at the impeller outlet becomes a smaller percentage of the total energy as blades become more backwardly curved. Therefore, a larger portion of the static pressure can be recovered in the impeller with backward curved vanes.

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FIG : 2.1 Effect of Exit Vane Angle on Outlet Velocity of Impeller

IMPELLER BLADE ANGLE AT THE SUCTION END (β1b) β1b used in impeller is with in a limited range for all machines. It is the angle at inlet for pump/comp and at exit for turbines. For radial fans and blowers, values outside this range reducing upto 20° are found to be in use. In the case of turbines, a low β 1b would mean more flow deflection in the impeller blade row with corresponding increase in specific work. With decreasing β 1b, the blade tangential thickness t1u at exit increases. From strength considerations, trailing edge thickness cannot be reduced to small values. Also this causes formation of eddied behind the blade trailing edge and results in wider wakes and more losses values between 15° to 35° are used.

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2.3VELOCITY TRIANGLES: The three velocities that make a velocity triangle are namely i Blade speed U ii Absolute velocity C iii Relative velocity W Generally the blade speed is taken as the base of the triangle, the direction of U1 and U2 follow the direction of rotation of impeller and W and C's direction vary depending on that and such that W=CU (In vectorial notation) is satisfied

FIG 2.2 : Velocity Triangle at Inlet of Impeller

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FIG 2.3: Velocity Triangle at outlet of Impeller

In a radial machine U2 greater than U1. Angle between C 'absolute velocity' and 'relative velocity' U is α and β is the angle between W and –U. The flow velocities are resolved into two components with respect to U, the component along U is Cu {may be C1u or C2u} and perpendicular to U i.e. along meridional plane is Cm and similarly Wu and Wm are obtained. To get the volume flow rate at the particular section Cm can be multiplied by flow area at that section hence its is called the 'flow velocity'. If the pre whirl is 0 then C1u = 0, hence it is desirable to design with consideration C1m = C2m whenever possible which also helps to maintain the blade angle within considerable range.

2.4 Impeller 19

The impeller forms the major component in the whole machine where the actual energy transfer to the fluid takes place.

In an actual impeller, complete guidance to

the fluid cannot be expected due to the limited number of vanes. The vane thickness, the viscous effects, the relative circulation, return flows and the effect due to bends make the velocity and pressure distribution far from uniform. The actual flow deflection is less than that obtained when the flow truly follows the vanes. The difference between the vane angle and the actual flow angle is accounted by the introduction of a factor called slip factor.

2.4.1 Slip In the case of vane congruent flow, the specific work of the machine is given by

W ∞ = U2 C2U - U1 CIU

The peripheral components of velocity just outside the impeller are different from those just within. This difference in specific work is due to the slip in the impeller that is the flow does not wholly follow the impeller vanes. The energy transfer obtained in practice is less than that calculated assuming the flow is one - dimensional and that the fluid outlet angle equals the impeller vane angle due to the relative eddy and nonuniform velocity profile at the impeller.

Pfleiderer defined the slip power factor p given : W bl∞ = (p+1)W ∞

Stodola assumed that the slip is due to the relative eddy and that the slip velocity is given by: σ = 1 –( (Π/Z)(Sin β2 /(1-Ф2 Cot β2 )) 20

2.4.2 Inlet Vane angle As the temperature of the air at the inlet is less. The sonic velocity is also less. There is the danger of the velocity in this region reaching a sonic value .For incompressible flow, the relative inlet velocity is a minimum when β 1 =35°. In compressible flow,

the relative inlet Mach Number is a minimum when β 1 is in

between 25° to 30°.

2.4.3 Pre Whirl The relative inlet mach number at impeller inlet can be reduced further by giving whirl velocity in the direction of rotation of the impeller. However this has the other effect of reducing the specific work of the stage. In designing usually the fluid is assumed to enter radially so that α 1 = 90°. As the fluid approaches the vans inlet it comes into contact with the rotating shaft and impeller. This tends to cause it to rotate with the wheel. This makes larger as shown by solid line

Effect of pre-rotation on the inlet diagram

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2.4.4 Impeller outlet angle The vane outlet angle has a major effect in the design and performance of the impeller. The optimum inlet angle having been fixed- by sonic velocity criterion in the case of a blower, the outlet angle directly controls the size, performance as well as the specific world developed The component C2u increases with increasing β 2

.

For a given specific work, the peripheral speed will come down or if the rotating speed is also fixed, the diameter comes down. But an increase in β 2 could cause adverse effects at the vane boundary.

2.4.5 Impeller outlet diameter The impeller outlet diameter as a ratio of the inner diameter should not be too large as otherwise the vane channels become long and narrow increasing the friction losses. On the other hand, a smaller ratio makes the length of the flow traverse inside the impeller quite small hampering the energy transfer between the impeller vanes and the fluid for radial machines the optimum value of this ratio is about 2.

2.4.6 Effect of viscosity The viscosity of the flowing medium causes the boundary layer to develop along the shroud and the vane faces in the channel resulting in a decrease of the area available for the flow of the fluid. Also pressure losses result because of this. Even simple friction losses are appreciable because of the high relative velocities and the large amount of wetted flow surface. Boundary layer effects may be appreciable because of the adverse velocity gradients of considerable magnitude present along the channel walls.

When the boundary

layer is not in equilibrium with the pressure gradient across the channel, a flow normal to the through flow may arise which will alter the desired potential flow pattern and cause direct losses as a result of the partial dissipation of the energy absorbed from the through flow to create the secondary motion.

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2.4.7 Inlet passage The inlet passage is meant to slowly accelerate the fluid from the entrance to the eye with minimum losses. An inlet nozzle is usually fitted at the entrance of the inlet nozzle design is important as otherwise it may affect the flow conditions at the entrance to the impeller.

2.4.8 Effect of Surface roughness The effect of the surface roughness becomes appreciable in small impellers where vane channels are very narrow. Varley found that in the case of a centrifugal pump, the effect of surface roughness is to increase the specific work developed and slightly reduce the efficiency without altering the shape of the specific work versus discharge curve.

2.4.9 Volute Casing This is normally employed in the single stage machines and in the last stage of the multi-stage machines.

Its main purpose is to collect the fluid emanating from

all around the periphery and discharge it into the exit flange. A spiral casing can be used with or without a diffuser ring.

The flow condition in the spiral casing is

given by the free vortex condition that is Cu. r

= constant

Another type of casing normally employed is the constant velocity volute having a constant average velocity at all sections and the volute area increases in proportion to the angular displacement from the torque where the velocity is zero.

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2.5 EFFECT OF GEOMETRIC AND FLOW PARAMETERS ON FAN PERFORMANCE

2.5.1 Size of impeller: The flow rate depends on impeller diameter and the width. For particular stage pressure rise the peripheral speed and geometry of the impeller can be decided. The diameter ratio (d1/d2) of the impeller determines the length of the blade passage. Smaller the ratio, larger is the blade passage. With slight acceleration of the flow from the impeller eye to the blade entry the following relation for the blade width to diameter ratio is recommended. b1/d2 = 0.2

Impellers with backward swept blades are narrower i.e. b1/d2
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