DESIGN AND ANALYSIS OF A LOW SPECIFIC SPEED CENTRIFUGAL FAN

A Dissertation Work Submitted to Jawaharlal NehruTechnological University

In Partial Fulfilment of the requirements of the award of

BACHELOR OF TECHNOLOGYIN

MECHANICAL ENGINEERING

By

K.KARTHIK 06311A0368A.SAI CHARAN 06311A0380D.ASHESH GOPAL NATH 06311A03B4

A.SREENU 06311A03B5

Department of Mechanical Engineering,

SREE NIDHI INSTITUTE OF SCIENCE & TECHNOLOGYYamnampet, Ghatkesar, Hyderabad-501301.

(Accredited by AICTE, New Delhi & Affiliated toJNT University, Hyderabad)

DESIGN AND ANALYSIS OF A LOW SPECIFIC SPEED CENTRIFUGAL FAN

A Dissertation Work Submitted to Jawaharlal NehruTechnological University

In Partial Fulfilment of the requirements of the award of

BACHELOR OF TECHNOLOGYIN

MECHANICAL ENGINEERING

By

K.KARTHIK 06311A0368A.SAI CHARAN 06311A0380D.ASHESH GOPAL NATH 06311A03B4

A.SREENU 06311A03B5

Under The Guidance ofDr.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 & TECHNOLOGYYamnampet, Ghatkesar, Hyderabad-501301.

(Accredited by AICTE, New Delhi & Affiliated toJNT University, Hyderabad)

2

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.

4

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

5

z Number of blades

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 Blade or Vaneh Hydraulicm Meridionalt Tipu Tangential or Peripheral component

6

CONTENTS

1 INTRODUCTION…………………………………………… 10

1.1 Introduction to Turbo machines

1.2 Fans – Principle of operation.

1.3 Classification of fans.

2 LITERATURE SURVEY……………………………………. 15

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

7

2.4.3 Volute and diffuser losses

2.5 Applications

3 DESIGN OF THE LOW SPECIFIC SPEED CENTRIFUGAL FAN …… 30

3.1 Fan Specifications.

3.2 Calculations

3.3 Auto CAD design of the Fan Impeller.

4 EXTRACTION OF COORDINATES……………………………. 38

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…………………………………………………… 42

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………………………………………………….. 51

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

6.6 Industry solutions using ANSYS

8

7 METHODOLOGY…………………………………………………. 55

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…………………………………… 71

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………………………………………………………. 92

10 SUGGESTION……………………………………………………….. 92

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.

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

10

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.

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

11

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

Mixed flow fans:

a. Axial Casing

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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

In backward curved centrifugal fans, the blades at the impeller are inclined away from

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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 (Cu1, Cu

2) of the absolute velocity vectors C

1 and C

2

respectively. 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.

14

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 – U1

2 + C12-C2

2+W02-W3

2)/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

15

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 β1 =35°. In

compressible flow, the relative inlet Mach number is a minimum when β1 is in

between 25° to 30°.

Exit vane angle

There are three considerations for β2b namely forward curved blades if β2b<90°, radial

blades when β2b=90° and backward curved blades if the angle β2b>90°. 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 C2u, 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 β1b would mean more flow deflection in the impeller

blade row with corresponding increase in specific work. With decreasing β1b, 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.

17

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=C-

U (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 ))

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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 β1 =35°. In

compressible flow, the relative inlet Mach Number is a minimum when β1 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 α1= 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

21

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

β2 . 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 β2 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.

22

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.

23

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<0.2

d1 / d2 = 1.2(Φ)1/3

d1 - Impeller inlet diameter

d2 - Impeller outlet diameterΦ - Flow coefficient

24

FIG : 1.1 Effect of Shape Number

2.5.2 Blade Shape

The blade loading, local acceleration and deceleration characteristics and impeller

performance are influenced by impeller blade shape. It plays an important role. Straight or

curved sheet metal blades or aerofoil shaped blades have been used in the centrifugal fans

and blowers. Sheet metal blades are arc-shaped or of different curve and can either be

welded or riveted to the impeller disc. These are classified as backward-swept, radial and

forward swept depending on the exit angles. The optimum blade angle at inlet was found to

be 35°.

2.5.3 Number of Blades

Too few blades are unable to fully impose their geometry on the flow, where as too

many of them restrict the flow passage and leads to higher losses. The number of blades in

centrifugal fan can wary from 3 to 64 depending on the application type and size. Some

empirical relations to determine the no of blades are given below.

Pfleiderer has recommended the following relation:

Z = k * (d2+d1)/ (d2-d

1) sin (0.5 (β

1+β

2)) (where k varies from 6 – 8.5)

25

2.5.4 Volutes and Diffuser

At the impeller exit of the fan the flow has considerable kinetic energy. This kinetic

energy can be converted into static energy by providing vaneless and vaned diffuser at the

exit of the impeller. The spiral casing as a collector of flow from the impeller or diffuser is

an essential pan of the centrifugal fan. The provision of vaned diffuser in a blower can give

a slightly higher efficiency than the blower with only a volute casing. However for majority

of the centrifugal fans and the blowers the higher cost and the size that result by employing

a diffuser outweigh its advantage. Therefore most of the single stage centrifugal fan

impellers discharge directly into volute casings. Some static pressure rise can also occur in a

volute casing. Volutes can be designed for constant pressure or constant average velocity.

The cross section of the volute passage may be square or rectangular, circular or

trapezoidal .The fabrication of the rectangular volute is from sheet metal.

2.5.5 Effects of Friction on the Characteristics

Frictional losses within the impeller and shock losses have a considerable effect on the

characteristics of a fan. These losses are proportional to average relative entry velocity,

which is proportional to the volume flow. The effects of frictional losses are influenced by

flow-separation and back flow.

26

2.6 LOSSES IN CENTRIFUGAL FANS

Losses occur in both the stationary as well as moving parts of the centrifugal fan stage. By

accounting for the stage losses, the actual performance of a fan or blower can be predicted.

The various losses are given below:

2.6.1 Losses in Impeller

The losses are categorized here as

a) Impeller internal losses and

b) Impeller external losses

Impeller Internal Losses

The impeller internal losses are those due to skin friction, blade loading, and blade-wake

mixing and impeller shroud clearance. Impeller skin-friction loss is defined as the loss

experienced by the fluid while flowing through the channels formed by the bounding

surface of the impeller. These losses specifically exclude the effects of the non uniform

velocity distribution caused by the work-addition process in the impeller on the blade-

surface boundary-layer behavior.

Impeller External Losses

The impeller external losses are those due to disk friction, recirculation at the impeller edges,

and leakage around shrouded impellers. The disk-friction loss is that due to the shear force

acting on the impeller caused by the fluid between the rotating and stationary surfaces. The

recirculation and scrubbing loss is that due to internal recirculation at either impeller-shroud

clearance or at the impeller exit, where in the fluid loses momentum in the process of

flowing back to the impeller and therefore necessitates an increase in the amount of work

required to be supplied by the impeller.

27

2.6.2 Leakage Losses

A clearance is provided between the rotating periphery of the impeller and the casing at

the entry. This leads to the leakage of some air and disturbance in the main flow field.

Besides this, leakage also occurs through the clearance between the fan shaft and the casing.

2.6.3 Diffuser and Volute Losses

Losses in the diffuser or volute occur due to friction and separation .At off-design

condition there are additional losses due to incidence .The flow form the impeller or diffuser

expands to a large cross sectional area in the volute. This leads to losses due to eddy

formation .Further losses occur due to the volute passage friction and flow separation.

2.7 FAN APPLICATIONS

Some of the important applications are Steam Power stations, Ventilation systems,

cooling of electric motors; Gas based power plants, Generators and many industrial process

plants.

Power Plant Auxiliaries

In Steam Power plants forced draft and induced draft fans are used to raise the pressure of

air and flue gases to overcome the draught losses in the flow passage of steam boiler. The

forced draft fan raises the pressure of the ambient air and delivers it to the boiler furnace

through air pre-heater. The induced draft fan is located between the furnace and the flue gas

chimney. Therefore these fans work in the hostile atmosphere of high temperature (150

degrees to 350 degrees centigrade) abrasive and corrosive gases. These fans are either axial

or centrifugal type and generally driven by electric motors. For pulverizing coal or fuel oil

small and large fans are used.

Cooling of Motors, Generators and Engines

In internal combustion engines and electric motors and generators considerable extent of

heat is needed to be removed. The cooling of the hot water in the radiators of an automobile

28

vehicle is a well-known example. The air sucked through the radiators cools the circulating

water as well as the engine. For this propeller fans are used and driven by the engine

through belt transmission drive. For cooling the electric motors, fans are generally mounted

on the extension of their shafts.

Air Circulation and Mine Ventilation

Fans of various ratings are used to circulate air in air conditioning systems. Besides this

fans are used to circulate air in a number of other applications as centrifugal separators,

furnaces, drying equipment and cooling of electric and optical equipment. Fans employed

for ventilation of mines and tunnels are heavy duty fans. The rating of fan is be obtained

from the number of workers in the mine and the total resistance to be overcome .Normally

centrifugal flow fans are frequently used compared to axial flow fans.

Steel Plants

In steel plant applications large and small fans are used. One or more high-pressure

blowers are also employed to supply blast furnace gases to the steam boilers. In such cases

impellers must be able to operate at high temperatures and speed. Main blast furnace

blowers are required to develop high pressures and therefore they apply many centrifugal

stages.

Other applications include pneumatic transport of granular materials, centrifugal

separators, furnace and drying equipment. The miniature fans are used in much equipment

for component cooling.

29

3.DESIGN OF THE LOW SPECIFIC SPEED

CENTRIFUGAL FAN

3.1 Fan Specifications

The following are the necessary specifications required for the design of a centrifugal fan.

i. Design flow rate : 11000 cum/hr

ii. Static pressure raise : 300 mmwc

iii. Approximate total pr : 315 mmwc

iv. Static Head rise : 250mair

v. Total head rise : 262.5mair

vi. Specific work : 2575.125 m^2/s^2

vii. Volume flow rate : 3.06 m^3/s

viii. Reference density : 1.2 kg/m^3

ix. Fan input power : 11.6 KW

x. Operating speed : 980 rpm

xi. Reference pressure : 1.0132 bar (76mm.Hg)

xii. Reference temperature : 20 deg.C

3.2 Design Calculations:

From the above specifications, the dimensions and other parameters of the fan are calculated.

Specific work, W = (g Ht) / ρ = 2575 m2/s2

Shape number, nsh = (N/60).√ ( Q / W0.75) = 0.0822

Specific speed, (σ) = 2.108 nsh = 0.171

For design, proper selection specific shape and specific diameter ensures good efficiency. So

cordier diagram is used, the x cordinate is σ and y is specific diameter δ

30

The value is found out to be δ = 5.562

Impeller tip diameter (D2) = (δ/1.8652)*(√ (Q)/H.25) {selected value is 1.33 m}

Impeller tip speed (U2) = 68.24 m/s

Entrance coefficient ε = C1m/√ (2W)=0.25 (assuming ε =0.2–0.3 for blowers and fans)

Therefore meridional velocity at inlet C1m = 17.94 m/s

Also C2m = 17.94 {assumed that pre whirl is zero}

C0 = 16.31 m/s (C1m/ C0 =1.1)

Eye diameter = 4 *√ ((Q / c0)/ Π) = 0.500 m

Impeller Inlet diameter D1 = 0.550 m ( De/D1=1.1)

31

Corresponding A1 (Q / C1m) = 0.1788m2

Width b1 = (A1 / D1) =0.1035m

Tip in let speed U1 = Π D1 N /60 = 28.22 m/s

From velocity triangles Tan β1 = C1m/U1 =0.6357

β1 = 32.44

Vane contraction factor = 1.1 (assuming)

C1mb = 19.75

Tan β1 = C1mb/U1 = 19.75

Final β1 =25.30 = 25

D2/D1 = 2.42

C2u = 37.73

Hydraulic efficiency from graph = 81.7%

Hydraulic efficiency assumed ηhyd = 81 %

Wbl = (W / ηhyd) = 3179.2

C2ubl = (Wbl / U2) = 46.58

Slip pow.Factor p (assumed to be) = 0.35

Wbl∞=Wbl / (1+ p)= 4291.9

C2ubl∞=Wbl∞/U2= 62.89

A1*(A2/A1) =0.1788 (since A1/A2=1)

b2=A2/(π.D2)= 0.0428 {b2 selected =0.043}

Vane C F 2= 1.0

C2mb =17.94

32

tan β2b = 3.349 (tan β2b =c2m/(U2-C2ubl∞))

Β2b = 73.37 {β2b selected = 73.0}

First Trial :

Z = 12.7 {Z = k (r2+r1)/(r2-r1)*sin(β1b+β2b)/2}

Z selected = 13

A2/A1 mean = 1.71

A2= 0.3056

b2 = 0.073 {b2 (selected) = 0.073}

Vane C.F2 = 1.0 (Assumed)

C2m = 10.5

C2mb =10.5

tan β2b =1.959

β2b = 62.96

β2b selected = 63.0

Second Trial :

Z = 11.7

Z (selected) =12

Vane thickness=4.0

Bf1 = 0.066

Cf2 =1.013

C2mb = 10.635

Estimation of slip factor (Pfleiderer)

ψ' = 1.333 { ψ'=k(1+β2b/60) where k =0.65}

33

P= 0.268 {P= 2*ψ'/z*(1/(1-(r1/r2)2)}

1 / (1+P) = 0.788

Wbl∞ = 4031 (wbl/sig-pfl)

C2ubl∞ = 59.06 (wbl∞/U2)

tan β2b = 1.1582{c2mb/(u2-c2ubl∞)}

Β2b = 49.2

Β2b (selected) = 50

Third Trial:-

Z = 10.3 {z selected = 10}

Vane thickness=4.0

Bf1 = 0.104

Vcf1 = 1.116

Bf2 = 0.012

Cf2 = 1.013

C2mb = 10.63

Estimation of slip factor (Pfleiderer)

ψ' = 1.192

P = 0.287

Radius of curvature Rc = 1.0285 m

Xc =0.7878m {xc=√ (rc2+r1

2-2rcr1cosβ1b)}

Shroud Taper = tan(ν)=0.0794 {tanν=(b1-b2)/(r2-r1)}

Taper angle (ν) = 4.54

34

3.3 AUTO CAD DESIGN OF THE FAN IMPELLER:-

Computer-aided design (CAD) is the use of computer technology for the design of

objects, real or virtual. CAD often involves more than just shapes. As in the manual drafting

of technical and engineering drawings, the output of CAD often must convey also symbolic

information such as materials, processes, dimensions, and tolerances, according to

application-specific conventions.

CAD may be used to design curves and figures in two-dimensional space; or curves,

surfaces, and solids in three-dimensional objects. It is an important industrial art extensively

used in many applications, including automotive, shipbuilding, and aerospace industries,

industrial and architectural design, prosthetics, and many more. CAD is also widely used to

produce computer animation for special effects in movies, advertising and technical

manuals.

CAD has become an especially important technology within the scope of computer-

aided technologies, with benefits such as lower product development costs and a greatly

shortened design cycle. CAD enables designers to lay out and develop work on screen, print

it out and save it for future editing, saving time on their drawings.

AutoCAD software is used to design a two-dimensional model of the impeller fan and it

is also used in extraction of co-ordinates. The process is explained in detailed steps with the

assist of figures below

35

Fig 4.1 Fan Auto CAD design 1.

1) Taking intersection of axes as the centre and radius draw 2 circles of radius 275 mm and 665 mm.

These form the inner diameter and outer diameter of the impeller. (Figure 1)

2) Draw another circle taking radius as 788 mm, and then draw of radius 1028mm and centre as the

intersection point of the x axis and the inner diameter. (Figure 1)

3) Two intersection points are obtained on either side of the horizontal axes. Depending on the direction

of the blades one of the points is chosen. Since we went for clockwise direction we choose the left

hand side point. (Figure 1)

4) From this point another circle of radius rc = 1028mm is drawn. (Figure 2)

5) This circle passes through the inner and outer diameter circles and the arc contained by these two

circles forms the blade. (Figure 2)

Fig 4.2 Fan Auto CAD design 2 Fig 4.3 Fan Auto CAD design 3

6) The enclosed arc is the median of the blade and it is shown in figure 3.7) Taking 2mm off set on either side of the blade median curve, two identical curves are drawn. The top

curve is the pressure side and the bottom curve is the suction side. (Figure 4)

36

Fig 4.4 Fan Auto CAD design 4

8) After obtaining one blade mirroring is used, where the numbers of blades are specified as 10 and angle as 360.

9) To generate the side view the taper is considered and the following figure is generated

37

4. EXTRACTION OF COORDINATES

4.1 Method of Extraction

The coordinates of blade, hub and shroud are extracted from the 2-d diagram of fan impeller.

Coordinates are used in generating a 3D figure in turbo grid.

A series of coordinates are absorbed from a 2D cad diagram,

i. The CAD diagram is first simplified to represent one blade passing through one of the axis.

ii. Further more the area between the inner radius and outer radius are divided at a series of

equal intervals.

iii. For example a series of concentric circles are drawn considering the center of the impeller as

shown in the fig.

iv. These lines intersect the blade profile at both pressure and suction side and also intersecting

the axis as shown.

v. Considering the geometrical x axis as y axis an geometrical y axis as x axis, using the crock

screw thumb rule the meridional geometrical x axis represents z axis.

vi. Now, considering the intersection point on the blade profile the perpendicular distance from

x and y as shown in fig., the x and y coordinates are absorbed.

vii. For the similar point the circle passing through the intersection also passes through the

geometrical X axis as seen in fig., a perpendicular is drawn to the meridional diagram.

viii. From the meridional diagram, as defined earlier the geometrical x axis is the z axis, from

this the perpendicular intersection the meridional diagram at both hub and shroud the "z-

hub" and "z-shroud" coordinates are extracted, as the representation uses the crock screw

thumb rule the values of z is considered negative.

ix. And for the leading edge a series of concentric circles with a difference of "2mm" are drawn

and coordinates are generated for the x, y, z-hub, z-shroud.

x. As the value of z is generated for both hub and shroud, by varying the values of z,

profile.curve coordinates are generated as a set for hub using the z-hub coordinates, and a

set for the z-shroud.

38

4.2 COORDINATES OF THE BLADE PROFILE (HUB SIDE)

R X Y Z

275 275 0 0

275.9 275.899 -0.1864 0

276.689 276.689 0 0

277.6588 277.6571 0.9718 0

305 300.2754 53.4761 0

345 323.6533 119.0809 0

385 341.0639 178.6069 0

425 353.9198 235.2993 0

465 363.2347 290.3198 0

505 369.43 344.3058 0

545 372.7376 397.6074 0

585 373.1085 450.4228 0

625 371.1548 502.8609 0

665 366.3716 554.9747 0

665 362.0027 557.8342 0

625 366.9361 505.9475 0

585 369.2318 453.7542 0

545 368.8591 401.2081 0

505 365.7523 348.21 0

465 359.6106 294.5772 0

425 350.7593 239.9852 0

385 338.2689 183.8454 0

345 321.3768 125.1295 0

305 298.8742 60.8211 0

277.6588 277.4731 10.1544 0

276.689 276.5684 8.1688 0

275.9 275.8225 6.5400 0

275 275 0 0

39

4.3COORDINATES OF THE BLADE PROFILE( SHROUD SIDE)

R X Y Z

275 275 0 104.000

275.9 275.899 -0.1864 103.859

276.689 276.689 0 103.736

277.6588 277.6571 0.9718 103.584

305 300.2754 53.4761 99.308

345 323.6533 119.0809 93.051

385 341.0639 178.6069 86.795

425 353.9198 235.2993 80.538

465 363.2347 290.3198 74.282

505 369.43 344.3058 68.026

545 372.7376 397.6074 61.769

585 373.1085 450.4228 55.513

625 371.1548 502.8609 49.256

665 366.3716 554.9747 43.000

665 362.0027 557.8342 43.000

625 366.9361 505.9475 49.256

585 369.2318 453.7542 55.513

545 368.8591 401.2081 61.769

505 365.7523 348.21 68.026

465 359.6106 294.5772 74.282

425 350.7593 239.9852 80.538

385 338.2689 183.8454 86.795

345 321.3768 125.1295 93.051

305 298.8742 60.8211 99.308

277.6588 277.4731 10.1544 103.584

276.689 276.5684 8.1688 103.736

275.9 275.8225 6.5400 103.859

275 275 0 104.000

40

4.4 COORDINATES OF HUB CURVE AND SHROUD CURVE

I In generating the hub.curve and shroud.curve file from the meridional view, a series of

horizontal lines intersection both hub and shroud lines are drawn, and the coordinates for

these intersection points are considered as shown in fig.

Hub curve

Shroud curve

5. CFD THEORY

X Y Z

0 0 -195

40 0 -155

80 0 -115

120 0 -75

160 0 -35

195 0 0

275 0 0

345 0 0

665 0 0

X Y Z

249.9640 0 -129.6601

250.4691 0 -124.6601

252.0511 0 -119.6601

254.964 0 -114.6601

259.964 0 -109.6601

271.0792 0 -104.9638

665 0 -43

41

5. CFD THEORY:

CFD is playing a strong role as a design tool as well research tool. In CFD, the fundamental

equations of fluid mechanics are based on the following universal laws of conservation:

1. Conservation of mass

2. Conservation of momentum

3. Conservation of energy.

5.1.1 Continuity Equation:

Physical principle: Mass is conserved.

Net mass flow out Time rate ofof control volume = decrease of massthrough surface S inside control volumePartial differential equation form of the continuity equation in differentiable conservative form

can be expressed as

42

Fundamental physical principles Governing equations of fluid flow

Mass is conserved Continuity equation

Newton’s second law Momentum equation

Energy equation

Energy conserved

Where,

→ Density

x, y, z → Cartesian Coordinates

u, v, w → velocity vectors in x, y, z directions.

L.H.S → Net mass flow out of the control Volume

R.H.S → Time Rate of Decrease of mass inside the control volume

The basic continuity equation of fluid flow is as follows:

Where, = Fluid density

= the rate of increase of density in the control volume.

The first term in this equation represents the rate of increase of density in the control volume

and the second term represents the rate of mass flux passing out of the control surface,

which surrounds the control volume. This equation is based on Eulerian approach. In this

approach, a fixed control volume is defined and the changes in the fluid are recorded as the

fluid passes through the control volume. In the alternative Lagrangian approach, an observer

moving with the fluid element records the changes in the properties of the fluid element.

Eulerian approach is more commonly used in fluid mechanics. For a Cartesian coordinate

system, where u, v, w represent the x, y, z components of the velocity vector, the continuity

equation becomes

/t +/ x (u) + / y (v) + / z (w) =0

5.1.2 Momentum Equation:

Here, Physical principle: F = ma (Newton's second law)

Newton's Second Law applied to a fluid passing through an infinitesimal, small, moving fluid

element. Only the forces in the x direction are considered and the momentum is conserved

in this direction and thus the X component of the momentum equation is derived.

43

Forces on a fluid element can be classified in a tree diagram as:

Based on the above classification of forces the momentum equation in differentiable

conservative form can be expressed as

in X direction

in Y direction

in Z direction

44

Where, V stands for the velocity vector of the fluid.

L.H.S represents the Substantial derivative of the product of mass and acceleration

R.H.S represents the summation of Pressure force, Normal and shear force, body force

t →

represents rate of increase of momentum per unit volume.

V →

represents the rate of momentum lost by convection

through the control volume surface.

f →

represents the body force per unit volume.

5.1.3 Energy Equation:

Physical principle: Energy is conserved.

The physical principle stated above is nothing more than the first law of

thermodynamics. When applied to a fluid passing through an infinitesimal fixed control

volume yields the energy equation i.e. increase in energy in the system is equal to the heat

added to the system plus the work done on the system.`

For a fluid element it can be represented as:

Energy in different conservation form is expressed as:

45

Where,

e → internal energy

V^2/2 → Kinetic Energy

K → Coefficient of thermal conductivity

L.H.S → the rate of Change of energy inside a fluid element

First four terms in the R.H.S corresponds to the Net Flux of heat into the element

Rest of the Terms in the R.H.S corresponds to the Rate of Work Done on the Fluid Element

Due to Surface and Body Forces.

In terms of enthalpy, the final form of Energy equation is

q

t

Q

Dt

Dp

Dt

Dh.

Where Φ is known as dissipation function.

5.2. Turbulence Models:

Special attention needs to be paid to accurate modeling of turbulence. The purpose of a

turbulence model is to provide numerical values for the Reynolds stresses at each point in

the flow. The objective is to represent the Reynolds stresses as realistically as possible,

while maintaining a low level of complexity. The turbulence model chosen should be best

suited to the particular flow problem. A wide range of models is available and type of model

that is chosen must be done so with care. It is understood that these models are not used

when modeling laminar flows.

The final result of the flow, turbulence, reaction, heat transfer, and multiphase

calculations will be a detailed map of the local liquid velocities, temperatures, chemical

reactant concentrations, reaction rates, and volume fractions of the various phases. These

outcomes can be analyzed in detail using graphical visualization, calculation of overall

parameters and integral volume or surface averages, and comparison with experimental or

plant data. This analysis phase is referred to as post processing. Because of improvements in

46

computer power and enhanced graphics software, it is now much easier for CFD analysts to

create animations of their data. These often help in understanding complex flow phenomena

that are sometimes difficult to see from static plots.

5.2.1. K-Epsilon Model:

Boussinesq suggested that the apparent turbulent shearing stresses might be related to the

rate of mean strain through an apparent scalar turbulent or "eddy" viscosity. For the general

Reynolds stress tensor the Boussinesq assumption gives

Where T is the turbulent viscosity, k is the kinetic energy of turbulence given by,

By analogy with kinetic theory, by which molecular (laminar) viscosity for gases be evaluated

with reasonable accuracy, we might expect that the turbulent viscosity can be modeled as:

Where vT and l are characteristic velocity and length scale of turbulence respectively. The

problem is to find suitable means of evaluating them.

Algebraic turbulence models invariably utilize boussinesq assumption. One of the most

successful of this type of model was suggested by Prandtl and is known as "mixing length

hypothesis".

Where l a mixing length can be thought of as a transverse distance over which particles

maintain their original momentum, some what on the order of a mean free path for the

collision or mixing of globules of fluid. The product l * u/y can be interpreted as the

characteristic velocity of turbulence, VT. In the above equation, u is the component of

velocity in the primary flow direction, and y is the coordinate transverse to the primary flow

direction.

There are other models, which use one partial differential equation for the transport of

turbulent kinetic energy (TKE) from which velocity scales are obtained. The length scale is

prescribed by an algebraic formulation. The most common turbulence model generally used

is the two-equation turbulence model or k-Є model. There are so many variants of this

model. In these models the length scale is also obtained from solving a partial differential

equation.

47

The most commonly used variable for obtaining the length scale is dissipation rate of

turbulent kinetic energy denoted by E. Generally the turbulent kinetic energy is expressed as

turbulent intensity as defined below.

, k= (Actual K.E in flow – mean K.E in flow)

The transport PDE used in standard k-f model are as follows

5. 3. Discretization of Governing Equations:

The above governing partial differential equations are continuous functions of x, y, z. In

the finite difference approach, the continuous problem domain "discretized", so that the

dependent variables are considered to exist only at discrete points.

Equilibrium problems usually result in a system of algebraic equations that must be solved

simultaneously throughout the domain in conjunction with specified boundary values. These

are mathematically known as elliptic problems. Marching problems result in algebraic

equations that usually solved one at a time. These are known as parabolic or hyperbolic

problems.

Three methods are generally used for discretization,

1. Finite difference method.

2. Finite control volume method.

3. Finite element method.

5.3.1 Finite Difference Method:

In terms of the flow-field variables, partial differential equations are totally replaced by a

system of algebraic equations, which can be solved for the values of the flow-field variables

48

at the discrete points only. In this sense partial differential equations have been discretized.

This method of discretization is called Finite difference method. Most common finite-

difference representations of derivatives are based on Taylor’s series expansion.

Forward difference

= Backward difference

O ( x) = truncation error due to neglected terms in series.

These are called first-order difference equations. So the partial difference equations have re-

placed by finite difference representation & finally converted into algebraic equations. It is

perhaps the simplest method to apply on uniform meshes, but it requires a high degree of

regularity of the mesh. This scheme was once popular.

5.3.2 Finite Volume Method:

The governing equations of fluid dynamics have been mathematically expressed in differential

form when numerical scheme applied to these differential equations, the computational

domain is subdivided into grid points, and the finite difference equations are solved at each

point. An alternate approach is integral form of the governing equations. In this approach,

the physical domain is sub divided into small volumes for 3D case and small areas for 2-D

and the dependent variables are evaluated either at the centers of volumes or corners of the

volumes. The conservation principles are applied to a fixed region in space known as

“control volume”. This integral form of the conservation statement is usually well known

from first principles, or it can in most cases, be developed from the PDE form of the

conservation law.

Consider unsteady 2-D heat conduction. The appropriate form of the conservation statement

for the control volume can be represented mathematically,

The first term in the above equation is an integral over the control volume, represents the time

rate of increase in the energy stored in the volume. The second term, an integral over the

surface of the volume, represents the net rate at which energy is conducted out through the

surface of the volume. This is the integral or control-volume form of conservation law. The

integral approach includes the Finite volume method and Finite element method. The FVM

method has an obvious advantage over a FDM. If the physical domain is highly irregular

and complicated since arbitrary volumes can be utilized to subdivide the physical domain.

49

Also since the integral equations are solved directly in the physical domain, no co-ordinate

transformations required. Another advantage of FVM is that mass, momentum and energy

are automatically conserved

\

6. ANSYS CFX

6.1 Introduction to ANSYS CFX

ANSYS CFX is a high-performance, general purpose CFD program that has been

50

applied to solve wide-ranging fluid flow problems for over 20 years. At the heart of ANSYS

CFX is its advanced solver technology, the key to achieving reliable and accurate solutions

quickly and robustly. The modern, highly parallelized solver is the foundation for an

abundant choice of physical models to capture virtually any type of phenomena related to

fluid flow: laminar to turbulent (including transition), incompressible to fully compressible,

subsonic to trans- and supersonic, isothermal or with heat transfer by convection and/or

radiation, non-reacting to combusting, stationary and/or rotating devices, single fluids and

mixtures of fluids in one or more phases (incl. free surfaces), and much, much more. The

solver and its many physical models are wrapped in a modern, intuitive, and flexible GUI

and user environment, with extensive capabilities for customization and automation using

session files, scripting, and a powerful expression language.

6.2 ANSYS CFX and the ANSYS Workbench Environment

ANSYS CFX software is fully integrated into the ANSYS Workbench environment, the

framework for the full suite of engineering simulation solutions from ANSYS. Its adaptive

architecture enables users to easily set up anything from standard fluid flow analyses to

complex interacting systems with simple drag-and-drop operations. Users can easily assess

performance at multiple design points or compare several alternative designs. Within the

ANSYS Workbench environment, applications from multiple simulation disciplines can

access tools common to all, such as geometry and meshing tools.

Geometry: ANSYS DesignModeler software is specifically designed for the creation and

preparation of geometry for simulation. Its easy-to-use, fully parametric environment with

direct, bidirectional links to all leading CAD packages acts as the geometry portal for all

ANSYS products to provide a consistent geometry source for all engineering simulations.

Meshing: Providing accurate CFD results requires superior meshing technology. ANSYS

Meshing provides a multitude of meshing technologies in a single application to allow users

to select the best option on a part-by-part basis. ANSYS ICEM CFD meshing tools also are

available and include unlimited mesh editing capabilities as well as structured hexahedral

meshing.

6.3 CFD Pre-Processing in CFX-Pre

The ANSYS CFX physics pre-processor is a modern and intuitive interface for the setup

51

of CFD analyses. In addition to a general mode of operation, predefined wizards are

available to guide users through the setup of common fluid flow simulations. A powerful

expression language gives users the ability to customize their problem definition in

numerous ways, such as with complex boundary conditions, proprietary material models or

additional transport equations. The adaptive architecture of CFX-Pre even allows users to

create their own custom GUI panels to standardize input for selected applications, and

thereby ensure adherence to established best practices.

6.4 The ANSYS CFX Solver

At the heart of ANSYS CFX software is its advanced solver technology using coupled

algebraic multigrid, the key to achieving reliable and accurate solutions quickly and

robustly. Its engineered scalability ensures a linear increase in CPU time with problem size

and parallel performance that is second to none. Users can follow convergence progress and

dynamically monitor numerical and physical solution quantities. Solver parameters,

boundary conditions and other parameters can be adjusted ‘on the fly’, without stopping the

solver. The ANSYS CFX solver uses second order numerics by default, ensuring users

always get the most accurate predictions possible. All simulations, whether for rotating

machinery, multiphase flows, combustion or any other physical model benefit enormously

from the coupled solver technology in ANSYS CFX software to achieve robust and scalable

flow solutions.

6.5 Post-Processing with ANSYS CFD-Post

Complete and powerful post-processing capabilities for ANSYS CFX results are provided

with ANSYS CFD-Post for both graphical and quantitative analysis. Together with full

scripting and automation, including report generation, CFD-Post ensures users get the most

out of their CFD simulations.

6.6 Industry solutions using CFX

52

1. Vortex structures in a four-stroke engine just after injection of fuel and intake valve

opening.

2. Nucleate boiling downstream of spacers in a fuel rod bundle assembly.

3. Prediction of heat transfer distribution in a shell and tube heat

Exchanger.

53

4. Prediction of wetness dispersion under non-equilibrium conditions for quan-

tification of thermo-dynamic performance in a low- pressure steam turbine.

7.METHODOLOGY

54

7.1Modelling and CFD Analysis of Centrifugal Fan Stage

Problem Solving Approach in CFD

The basic steps involved in solving any CFD problem are as follows:

Identification of flow domain.

Geometry construction or Component Modelling.

Grid generation.

Specification of boundary conditions and initial conditions.

Selection of solver parameters and convergence criteria.

Results and post processing.

The Centrifugal Fan Stage is modelled and analysis is carried out by following above steps.

Identification of Flow Domain:-

Before constructing grid, it is required to understand the exact flow domain properly. The

flow domain in the case of Centrifugal fan consists of Impeller, where Impeller is a rotating

component and others are stationary. It is therefore required that before going ahead with 3D

modelling and grid generation, the common interfaces should be clearly defined. The

software that is used is decided later based on nature and complexity of the geometry. For

axis-symmetry bladed geometry, the data for hub, shroud and blade profiles are obtained

from 2D drawing and subsequently grids are generated using Turbo-Grid software..

The boundary wall is the region where no slip condition exists and the velocity gradually

increases and reaches to mainstream velocities. That means, velocity gradient exists there

and that region close to the boundary wall should have fine grids.

3D CAD MODELLING:-

3D Geometrical Model of Impeller:-

The blade of the present Impeller is of 3D type and the modelling of Impeller blade is

rather complex compared to 2D curved blades. 3D blade involves thickness and twist

distribution as the blade extends between hub and shroud surfaces.

The geometrical design of blade profile is extracted from blade co-ordinates of line

elements, camber surface and distribution of thickness on the camber surface. The basic

design data is given in the form x, y, z co-ordinates of line elements. Line elements are

located along the radial positions of the blade, and some of the line elements are located

upstream of the blade leading edge, and like-wise also extends downstream of the trailing

55

edge. The sample data for line elements are given in the, this data is arranged in order to

obtain hub and shroud blade profiles. This process requires programming file in

TURBOGRID, which can transfer large amount line data instantly.

CUTTING THE TRAILING AND DRIVING SURFACES

Hub.curve

X Y Z0 0 -195

40 0 -15580 0 -115

120 0 -75160 0 -35195 0 0275 0 0345 0 0665 0 0

Shroud.curve

X Y Z249.9640 0 -129.6601250.4691 0 -124.6601252.0511 0 -119.6601254.964 0 -114.6601259.964 0 -109.6601

271.0792 0 -104.9638665 0 -43

7.2MERIDIONAL DATA FOR HUB & SHROUD CONTOURS

By using the above data we get the meridional view of the hub and shroud contours of

the impeller as shown

The hub curve runs upstream to downstream and must extend of the blade leading

edge. The hub data file contains the hub curve data points in Cartesian form and

downstream of the blade trailing edge. The profile points are listed, line-by-line, in order

from upstream to downstream. These data points are used to place the nodes on the hub

surface, which is defined as the surface of revolution of a curve joined by these points.

Shroud Data File

The shroud data file contains the shroud curve data points in Cartesian or cylindrical form the

shroud curve runs upstream to downstream and must extend upstream of the blade leading

56

edge and downstream of the blade trailing edge the points are listed, line by line in free

format style in order from upstream to downstream. These data points are used to place the

nodes on the shroud surface, which is defined as the surface of revolution of a curve joined

by these points.

Example: Considering XZ Plane with ‘Z’ as Axis of Rotation

Fig: Hub Curve and Shroud Curve Profile curve Data File:

The “profile” data file contains the blade “profile” curves in Cartesian or cylindrical form.

The profile points are listed, line-by-line, in a closed loop surrounding the blade. The blade

profiles should lie on a surface of revolution to facilitate transformation to m-prime, theta

conformal space.

A minimum of two blade profiles are required, one which lies exactly on the hub surface

and one which lies exactly on the shroud surface. The profiles must be listed in the file in

order from hub to shroud. Multi bladed geometries are handled by placing multiple blade

profile definitions in the same profile.

Profile. Curve:

57

# Profile 1X Y Z275 0 0275.899 -0.1864 0276.689 0 0277.6571 0.9718 0300.2754 53.4761 0323.6533 119.0809 0341.0639 178.6069 0353.9198 235.2993 0363.2347 290.3198 0369.43 344.3058 0372.7376 397.6074 0373.1085 450.4228 0371.1548 502.8609 0366.3716 554.9747 0362.0027 557.8342 0366.9361 505.9475 0369.2318 453.7542 0368.8591 401.2081 0365.7523 348.21 0359.6106 294.5772 0350.7593 239.9852 0338.2689 183.8454 0321.3768 125.1295 0298.8742 60.8211 0277.4731 10.1544 0276.5684 8.1688 0275.8225 6.5400 0275 0 0

#Profile 2

X Y Z275 0 104.000

58

275.899 -0.1864 103.859276.689 0 103.736277.6571 0.9718 103.584300.2754 53.4761 99.308323.6533 119.0809 93.051341.0639 178.6069 86.795353.9198 235.2993 80.538363.2347 290.3198 74.282369.43 344.3058 68.026372.7376 397.6074 61.769373.1085 450.4228 55.513371.1548 502.8609 49.256366.3716 554.9747 43.000362.0027 557.8342 43.000366.9361 505.9475 49.256369.2318 453.7542 55.513368.8591 401.2081 61.769365.7523 348.21 68.026359.6106 294.5772 74.282350.7593 239.9852 80.538338.2689 183.8454 86.795321.3768 125.1295 93.051298.8742 60.8211 99.308277.4731 10.1544 103.584276.5684 8.1688 103.736275.8225 6.5400 103.859275 0 104.000

The first step is to check whether the blade profile data obtained from solid model is

intersecting hub and shroud curves or not. We use CFX-Turbogrid intersect option for this

purpose. Using this option, we have to see that blade profile must lie on the surface of

revolution of hub and shroud as shown in fig Turbo grid intersecting capability can convert

an existing set of blade profiles that does not necessarily lie on the surface of revolution into

one that can be used in a CFX-Turbogrid template.

Next step is generating grid. Among the various templates available in turbogrid, Multi

Block Grid template as shown in fig is used. By the way of adjusting control points in fig a

good quality hexahedral grid can be generated. Flip topology is used to correct negative grid

volume due to left-handed system. The mesh command creates mesh grid but also calculates

and displays the minimum and maximum skew angle in the grid and the node at which it

occurs. The ‘View’ command in the GUI window can be used to see the different views of

the grid like Cartesian view, Meridional view and blade-to-blade view as shown in the

figure.

59

Setting the topology for the mesh grid

Adjusting the control points at the Leading Edge & Trailing Edge

60

3-D view of impeller without shroud surface

61

3D View of Impeller with hub Surfaces

The mesh generated by adjusting the control points as shown in Fig and

correspondingly Circumferential view of 3d Impeller surfaces & Periodical

arrangement of blades through out the circumference are shown in Figs

VIEWS FOR 3D IMPELLER BLADE MESH

The following parameters were considered to check the quality of the grids:

Skew angle: It is defined as the internal angle of the octahedron. Ideally, all the

angles should be equal to 90 degrees to get a perfect orthogonal grid. However, for

practical purposes, the grid is considered to be of high quality if the minimum

skew angle is not lower than 15 degrees and the maximum skew angle is not

greater than 165 degrees.

Grid volume: Negative volume meant overlapping of adjacent grids, which would

lead to errors in solver. Care was taken to ensure that there was no negative

volume in the grids.

Aspect ratio: It is defined as the ratio of the longest side to the shortest side. Its

minimum value is 1. For good quality grid creation, the maximum aspect ratio

63

should be less than 200.

The mesh is generated for the 3D Impeller with the total number of nodes, maximum

and minimum skew angle and aspect ratio obtained from TURBOGRID are given

in Table 4.3.

7.3MESH DATA FOR 3D IMPELLER BLADES

S.No Component Number of

Nodes

Number of

Elements

Minimum

Skew Angle

Maximum

Skew Angle

1 3DIMPELLER 19380 16352 18 163

Specification of boundary conditions and initial conditions: - This part of

simulation is done in CFX-Pre processing .The files with the extensions: “. grd”,

“.gci”, “.bcf” of 3DImpeller are copied into a new folder separately and these grid

files are read into pre-processing model of CFX-11 software.

PHYSICS DEFINITION:

Physics definition involves defining the physical parameters such as pressure,

temperature, mass flow, etc. and other boundary conditions relevant for the

problem. Pre-processing involves the following steps. The software used was CFX

PRE 11.

I. Importing the mesh assembly and region definition: The mesh file (.grd) file

was imported separately for the 3DImpeller. The grid file (with extension .grd) is

the file necessary to generate the grid.

The .gcf file contains the topographical details of the file, while the boundary

conditions file (.bci) specifies the inlet, exit, periodic and the blade regions for

each assembly.

II. Defining the domain and boundaries: The 3DImpeller regions like inlet,

outlet, and blade are defined. The hub, shroud and the blades of both assemblies

were treated as walls. The interface between the periodic1 and periodic2 is defined

as rotational periodicity. The boundary conditions were applied at inlet and outlet.

III. Initial conditions: - The initial condition for the pressure field should be the

average of the highest value of pressure specified on any of the Outlet boundaries

and the lowest value of pressure specified on any of the Inlet boundaries. This

reduces the likelihood of spurious inflow at Outlets, or outflow at Inlets, during the

course of the solution. A sensible initial guess for the temperature field is an

64

average of the boundary condition temperatures.

In the pre processing the following fluid domains and boundary conditions are

specified.

1. Simulation : Steady State

2. Domains : Fluid

R1 : Impeller (Rotating)

3. Boundary Conditions:

Inlet : Impeller inlet

Outlet : Impeller exit

Inlet Relative Pressure : 1.0132 bar

Wall : smooth

Mass flow : 3.672 kg/s

4. Fluid Properties:

Working Fluid : air at 25C

Density : 1.2 kg/m3

5. Rotation Axis : Z

65

6. Turbulence Model:

Turbulence Model : k-Epsilon

Heat transfer Model : None

7. Interface:

Type : Fluid -Fluid

Interface models : Rotational periodicity

8. Write Solver File:

After specifying all conditions write definition file using write solver file

command.

7.4 Selection of solver parameters and convergence criteria:

The flow governing equations are solved in CFX-Solver. The CFX-Solver Manager is

a graphical user interface used to set attributes for a CFD calculation, Control the CFX-

Solver interactively and to View information about the emerging solution.

The solver solves the mass, momentum and energy equations and calculates

pressure, velocity, enthalpy etc in the flow domain in each control volumes. The

inlet relative pressure and reference pressure plays a vital role to avoid round-off

errors. Reference pressure is the absolute pressure datum from which all other

pressure values are taken. It is a property of the entire simulation. So all domains

must use the same reference pressure value. The reference pressure will affect the

value of every other pressure set in simulation. It is used to avoid problems with

round-off errors which can occur when the dynamic pressure change in a fluid, that

drives the flow are small compared to the absolute pressure level. The relative

pressure specification set is measured relative to the reference pressure value. The

solver parameters are

1. Basic Settings: Steady State Simulations Advection Scheme is carried out using a

Numerical Advection Correction Scheme (Specify Blend). This selection allows

setting a Blend Factor between 0.0 and 1.0 for the advection scheme. A value of

0.0 is equivalent to using the First Order Advection Scheme and is the most robust

option. A value of 1.0 uses Second Order differencing for the advection terms; this

is not the same as the High Resolution advection scheme. This setting is more ac-

curate but less robust. Values between 0.0 and 1.0 blend First and Second Order

differencing, with increased accuracy and reduced robustness as you approach 1.0.

66

At the higher values overshoots and undershoots can appear, at lower values ex-

cessive diffusivity can occur. It is therefore recommended to use a value of 0.75 for

good accuracy of CFD results.

2. Timescale Control for a steady state simulation: The selection of an appropriate

time step size is essential in order to obtain good convergence rates for simulation.

In general there are two situations in which we use a physical time step:

to provide sufficient relaxation of the equation non-linearity’s so that a con-

verged steady state solution is obtained, or,

To evolve the solution through time in order to provide transient informa-

tion about a time dependent simulation.

Physical Time step

This option allows a fixed time step size to be used for the selected equations over the

entire flow domain. For advection-dominated flows, the physical time step size

should be some fraction of a length scale divided by a velocity scale. A good

approximation is the Dynamical Time for the flow. This is the time taken for a

point in the flow to make its way through the fluid domain. For many simulations a

reasonable estimate is easy to make based on the length of the fluid domain and the

mean velocity,

3. Max. No. Iterations are the maximum number of iterations the CFX-Solver will

run.

4. Residual Type is set to either RMS or MAX and a residual target is specified for

the convergence. The residual is a measure of the local imbalance of each conser-

vative control volume equation. It is the most important measure of convergence as

it relates directly to whether the equations have been solved. We can either select

MAX (maximum) or RMS (root mean square) normalized values of the equation

residuals as your check for convergence. The CFX-Solver will terminate the run

when the equation residuals calculated using the method specified is below the Re-

sidual Target value.

For the present simulation Solver Parameters are specified as follows:

Advection scheme :Specified Blend Factor (0.75)

Time Scale Control :Physical Time Scale (0.0003 sec)

Maximum Iterations : 200

Residual Convergence criteria : RMS

67

Residual Convergence Target: 1E-3

5. Run the solver monitor.

The solver is allowed to run till the required convergence is obtained.

7.5 Blade Geometry Plot

Isometric 3-D view of blade, hub & shroud

68

Meridional view

Blade mesh plot

Mesh element at 50% span

POST PROCESSING:-

CFX-Post is a flexible state-of-the-art post-processor. It is designed to allow easy

visualization qualitative and quantitative post-processing of the results of CFD

simulations.

69

Once the solution is converged, the solver writes all the data related to grid,

boundary conditions and flow parameters are stored in the result file. It is a binary

file, which can be opened by loading result file in CFX-Post, and the results are

analyzed. The performance of compressor stage is studied by using suitable

macros. The various plots are drawn and listed in results. Using the function

calculator option parameters like Mass flow rate, Velocity, Pressure, Enthalpy,

Entropy etc can be calculated. Plots are also available for various parameters like

Velocity, Pressure and Mach number etc, which show the variation of parameters

through out the domain. The efficiency, torque and power are obtained using

software’s macro.

The similar type of stage analysis is carried out for different mass flows i.e.

70%,80%,90%,110%,120%,130%etc.

70

8.RESULTS AND DISSCUSSION

8.1GENERAL

The simulated investigation on the impeller of a centrifugal fan are presented and

interpreted in this chapter. Data extraction and interpretation form a very important

part of CFD analysis to show conformity of simulated data with the experimental

results

The chosen centrifugal fan has an impeller diameter of 900 mm and an exit width of

83 mm.

The simulation is conducted on the impeller of a fan at various speeds. The various

speeds that were considered are Design Speed of 1450 RPM, 980 RPM and 2900

RPM rpm. Flow is analysed for different flow rates. The flow rates considered are

75,85, 90, 100, 110, 120,130.

The different parameters chosen for comparison are:

1) Velocity magnitude

2) Pressure ratio

3) Total and static pressure

4) Head coefficient

5) Shaft power

6) Overall efficiency

The above mentioned parameters are plotted with respect to radius, mass flow rate

and speed. The following plots are generated from the present CFD analysis for

better understanding of the following phenomenon to centrifugal fan:

71

1) Vector Plots

2) Path Lines

3) Contour Plots

8.2VARIATION OF FLOW PARAMETERS IN THE CHOSEN

IMPELLER

With respect to flow rate and speed

8.2.1Variation of velocity magnitude with flow rate and effect of speed

For a speed of 1450 RPM rpm it is clear that the velocity increases upto the

design flow and after that it slightly falls. Similarly in the case of higher speeds,

for various flow rates the velocity magnitudes are given in the table. The other

values of higher flow rates can be observed from the figure. It is also observed that

the absolute velocity is higher for higher speeds i.e. as speed increases, absolute

velocity increases.

8.2.2Variation of relative velocity with mass flow rate

For a given speed, absolute flow and is found to increase with flow rate. This is

evident for the increase of relative velocity from 32 to 75 for 1450 RPM

Relative velocity is the tangent inverse of the ratio of radial velocity and tangential

velocity. It can be seen form the figure that the tangential and radial velocity is

increasing with flow rate. The reason to justify this increase of relative velocity is

the greater increase of tangential velocity than radial velocity.

Since for higher speeds it results in higher velocity, absolute relative velocity

72

increases for various flow rates.

8.2.3Variation of Static Pressure and Total Pressure with flow rate and

effect of Speed

Static Pressure: It is found that static pressure decreases with flow rate and static

pressure ratio is found to increase with speed for respective mass flow rates. Static

pressure values are tabulated in the table. Variation of static pressure ratio is also

found to be similar to static pressure variation.

Total Pressure: It is found that the total pressure decrease with flow rate. Total

pressure ratio is found to increase with speed for respective mass flow. The total

pressure values are tabulated in the table. Variation of total pressure ratio is also found

to be similar to total pressure variation

8.2.4Variation of static pressure along the pressure and suction side of the

impeller vane

It is clear that static pressure increases up to a certain radius, but reduces there after

on the pressure side of the blade. This behaviour is because the blade extends only

upto a particular radius.

73

8.3RESULTS

8.3.1Values obtained from CFX for 980 rpm

8.3.2Values obtained from CFX for 1450 rpm

% flow %

flow per pas-sage, cu.m Flow coeff

exit total pressure

(pa)Pr

rise(pa)pre .stati

chead coeff Cu2total efficiencyflow angle al-

pha

70 0.25 0.009 104924 3606.67 103609 0.1638 -45.4287 97.8786 75.1928

80 0.29 0.0103 104795 3476.97 103540 0.1579 -43.7344 98.1630 72.5490

90 0.33 0.0116 104673 3355.63 103464 0.1524 -42.2680 98.2193 69.8385

100 0.36 0.0129 104558 3240.39 103380 0.1471 -40.9320 98.0767 67.1102

110 0.40 0.0142 104428 3114.03 103295 0.1414 -39.5281 97.6966 64.3117

120 0.44 0.0155 104295 2978.13 103160 0.1352 -38.1747 96.8224 61.4162

130 0.47 0.0167 104138 2821.91 103011 0.1281 -36.7975 95.2642 58.4596

% flow %

flow per pas-sage, cu.m Flow coeff

exit total pressure

(pa)Pr

rise(pa)pre .stati

chead coeff Cu2total effi-

ciencyflow angle al-

pha

70 0.25 0.0061 109831 8513.37 106513 0.1766 -71.51 96.42 63.94

80 0.29 0.007 109512 8195.59 106385 0.1700 -68.13 97.17 57.76

90 0.33 0.0078 109475 8159.74 106435 0.1692 -69.86 97.41 77.44

100 0.36 0.0087 109286 7971.2 106364 0.1653 -67.93 97.82 75.82

110 0.40 0.0096 109092 7777.36 106273 0.1613 -66.12 98.11 74.05

120 0.44 0.0104 108905 7589.8 106167 0.1574 -64.51 98.25 72.26

130 0.47 0.0113 108727 7412.34 106056 0.1537 -63.05 98.30 70.44

74

8.3.3Values obtained from CFX for 2000 rpm

8.4GRAPHS AND PICTORIAL ANALYSIS OF CFD

CONTOURS:

The form of representation of area under the action of a particular force, which are

shown in the form of colors representing a significant value. The following are a

few contour plots representing pressure, velocity and relative velocity at various

speeds of 980, 1450 and 2000 RPM and for various flow rates.

% flow %

flow per pas-sage, cu.m Flow coeff

exit total pressure

(pa)Pr

rise(pa)pre .stati

chead coeff Cu2total effi-

ciencyflow angle al-

pha

70 0.25 .0044 118556 17240.2 111481 0.188 -106.23 93.78 37.64

80 0.29 .005 118163 16846.7 111397 .0.183 -103.5 94.9 37.04

90 0.33 .0057 117753 16437.4 111285 0.179 -100.4 95.88 64.70

100 0.36 0.0063 117372 16057.0 111149 0.175 -97.5 96.7 63.53

110 0.40 0.0069 116913 15600 110958 0.17 -94 97.21 56.53

120 0.44 0.0076 116860 15547 111042 0.169 -96.6 97.33 74.57

130 0.47 0.0082 116889 15378 111002 0.167 -95.2 97.62 76.82

75

PRESSURE

Total pressure:-

Total pressure for 980 RPM at 70% flow Total pressure at 980 RPM at 100% flow

Total pressure for 980 RPM at 130% flow

76

Relative pressure:-

Rel. pressure for 980 RPM at 70% flow Rel. pressure at 980 RPM at 100% flow

Rel. pressure for 980 RPM at 130% flow

77

Static pressure:-

Ps for 980 RPM at 70% flow Ps for 980 RPM at 100% flow

Ps for 980 RPM at 130% flow

78

Pt at 980 RPM at 100% flow Pt at 1450 RPM at 100% flow

Pt at 2000 RPM at 100% flow

79

VELOCITY

Velocity vectors at 50% span:-

Velocity at 980 RPM at 70% flow Velocity at 980 RPM at 100% flow

Velocity at 980 RPM at 130% flow

Velocity at 1450 RPM at 100% flow Velocity at 2000 RPM at 100% flow

80

MERIDONIAL PLOTS

Contour plot of Pt at 980 RPM at 70% Contour plot of Pt at 980 RPM at 100%

Contour plot of Pt at 980 RPM at 130

81

STREAM LINE PLOT

Vel. stream at 980 RPM at 70% flow Vel. stream at 980 RPM at 100% flow

Vel. stream at 980 RPM at 130% flow

Vel. stream at 1450 RPM at 100% flow Vel. stream at 2000 RPM at 100% flow

82

BLADE LOADING

At 1450 RPM and 70% flow :-

At 1450 RPM and 100% flow :-

At 1450 RPM and 130% flow :-

83

At 980 RPM and 70% flow :-

At 980 RPM and 100% flow :-

84

At 980 RPM and 130% flow :-

At 2000 RPM and 70% flow :-

85

At 2000 RPM and 100% flow :-

At 2000 RPM and 130% flow :-

86

8.5 GRAPHS

PRESSURE RISE VS MASS FLOW

87

TOTAL PRESSURE VS MASS FLOW RATE

88

HEAD COEFFICIENT VS MASS FLOW RATE

89

TOTAL EFFICIENCY VS MASS FLOW

90

SHAFT POWER VS MASS FLOW (AT 980 RPM)

FLOW COEFFICIENT VS HEAD COEFFICIENT

91

9.CONCLUSIONS

A low specific speed centrifugal fan was designed for the given flow and

head conditions. The fan impeller was modelled using ANSYS Turbo Grid and was

analysed using CFX package.

The fan performance was evaluated and studied for different flow

conditions covering design and off-design points of operation and also for different

speeds.

The performance is seen to be following the normal trend for a low specific

speed fan and the flow and head curve shifts upwards with increasing speed.

The impeller efficiency seen to be maximum at the design point and decreasing

at off-design conditions. The efficiency is found to be above 90%, this is because

the windage losses, frictional losses have not been accounted.

The different contour and vector plots as well as the blade loading curve are

included for typical cases of design and off-design conditions.

The pressure rise is seen to increase uniformly along the impeller passage.

10.SUGGESTIONS FOR FUTURE WORK

This work may be extended by varying the number of impeller blades and

also by including the volute casing to get the total fan performance.

92

11.REFERENCES

a) Prithvi Raj & Gopala Krishnan Treatise on Turbo Machine b) Wolfgang Scheer Introduction to Turbo Machinery c) Balje, O.D. A Contribution to the problem of Designing Radial Turbo Machines d) Pfleiderer Die kreisel pumpen

e) Wikipedia.org

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