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 & d 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 & d to JNT University, Hyderabad)

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ACKNOWLEDGEMENT We would take immense pleasure to acknowledge with gratitude, the help & 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 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 age

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 Diff 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 diff losses

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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…………………………….

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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……………………………………

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 encomes 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 ages 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 ed by what are known as return guide ages 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 ages 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 ing 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 age 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 age (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 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 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 ed 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 deg usually the fluid is assumed to enter radially so that α 1 = 90°. As the fluid approaches the vans inlet it comes into 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 age The inlet age 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 diff 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 age. Smaller the ratio, larger is the blade age. 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 1/3 d1 / d2 = 1.2(Φ) d1 - Impeller inlet diameter d2 - Impeller outlet diameter Φ - Flow coefficient

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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 age 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+d )/ (d2-d ) sin (0.5 (β +β )) 1 1 1 2

(where k varies from 6 – 8.5)

2.5.4 Volutes and Diff 25

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 diff at the exit of the impeller. The spiral casing as a collector of flow from the impeller or diff is an essential pan of the centrifugal fan. The provision of vaned diff 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 diff 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 age 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 flowseparation and back flow.

2.6 LOSSES IN CENTRIFUGAL FANS 26

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

2.6.2 Leakage Losses A clearance is provided between the rotating periphery of the impeller and the casing at 27

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 Diff and Volute Losses Losses in the diff 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 diff 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 age 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 age 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 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. 28

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.

3.DESIGN OF THE LOW SPECIFIC SPEED CENTRIFUGAL FAN

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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) Corresponding A1 (Q / C1m) = 0.1788m2 Width b1 = (A1 / D1) =0.1035m 31

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 tan β2b = 3.349 (tan β2b =c2m/(U2-C2ubl∞)) Β2b = 73.37 {β2b selected = 73.0} 32

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}

P= 0.268 {P= 2*ψ'/z*(1/(1-(r1/r2)2)} 1 / (1+P) = 0.788 33

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+r12-2rcr1cosβ1b)} Shroud Taper = tan(ν)=0.0794

{tanν=(b1-b2)/(r2-r1)}

Taper angle (ν) = 4.54

3.3 AUTO CAD DESIGN OF THE FAN IMPELLER:-

34

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

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)

35

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 es 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

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,

37

i.

The CAD diagram is first simplified to represent one blade ing 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 ing through the intersection also es 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 "zshroud" 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.

4.2 COORDINATES OF THE BLADE PROFILE (HUB SIDE)

38

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)

40

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

41

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

42

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

Shroud curve 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

5. CFD THEORY 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. 43

Fundamental physical principles

Governing equations of fluid flow

Mass is conserved

Continuity equation

Newton’s second law

Momentum equation

Energy conserved Energy equation

5.1.1 Continuity Equation: Physical principle: Mass is conserved.

Net mass flow out Time rate of of control volume = decrease of mass through surface S inside control volume Partial differential equation form of the continuity equation in differentiable conservative form can be expressed as

Where,  

→ Density

x, y, z

→ Cartesian Coordinates

u, v, w

→ velocity vectors in x, y, z directions. 44

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 ing 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 es 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 ing 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.

45

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 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.

46

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 ing 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:

Where, e

→ internal energy

V^2/2

→ Kinetic Energy

K → Coefficient of thermal conductivity 47

L.H.S → the rate of Change of energy inside a fluid element First four in the R.H.S corresponds to the Net Flux of heat into the element Rest of the in the R.H.S corresponds to the Rate of Work Done on the Fluid Element Due to Surface and Body Forces. In of enthalpy, the final form of Energy equation is

ρ

Dh Dp δQ = + − ∇.q + φ Dt Dt δt

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 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  ∂u ∂u  2  ∂u  ' − ρui u 'j = µT  i + j  − δ ij  µT k + ρ k   ∂x    j ∂xi  3  ∂xk Where µT is the turbulent viscosity, k is the kinetic energy of turbulence given by,

48

k=

ui'u 'j 2

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:

µT = ρν T l 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".

µT = ρl 2

∂u ∂y

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 * | δ u/δy| 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. 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 σ as defined below.

(

'2

'2

k = 1/ 2 u + ν + w

1  u ' + ν ' + w' σ = 3 U  2

2

)

'2 , 2

   

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

1/ 2

µ T = Cµ ρ k 2 / ε The transport PDE used in standard k-f model are as follows

49

ρ

 ∂u Dk ∂  ∂ k    ∂ ui ∂ u j  2 = − ρ kδ ij  i − ρ ε  ( µ + µ T // P rk )  +  µ T  + D t ∂ x j  ∂ x j    ∂ x j ∂ xi  3  ∂ x j

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 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 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.

ui , j+ 1 − ui , j ∆x

+ O(∆ x)

Forward difference

50

 ∂u     ∂x  i , j

=

ui, j − ui, j − 1 ∆x

+ O (∆ x)

Backward difference

O ( ∆ x) = truncation error due to neglected in series. These are called first-order difference equations. So the partial difference equations have replaced 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,

∫∫

∫ρc

R

∂T dR + ∫ ∂t

∫ q.nds = 0 S

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. 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

51

\

6. ANSYS CFX 6.1 Introduction to ANSYS CFX ANSYS CFX is a high-performance, general purpose CFD program that has been 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 52

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 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 s to easily set up anything from standard fluid flow analyses to complex interacting systems with simple drag-and-drop operations. s 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 s 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 of CFD analyses. In addition to a general mode of operation, predefined wizards are available to guide s through the setup of common fluid flow simulations. A powerful expression language gives s 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 s to create their own custom 53

GUI s 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 U time with problem size and parallel performance that is second to none. s 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 s 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 s get the most out of their CFD simulations.

6.6 Industry solutions using CFX 1. Vortex structures in a four-stroke engine just after injection of fuel and intake valve opening.

54

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.

55

4.

Prediction of wetness dispersion under non-equilibrium

conditions for quanti-

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

7.METHODOLOGY 56

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 axissymmetry 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 edge. The 57

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 0 40 80 120 160 195 275 345 665

0 0 0 0 0 0 0 0 0

Z -195 -155 -115 -75 -35 0 0 0 0

Shroud.curve X 249.9640 250.4691 252.0511 254.964 259.964 271.0792 665

Y 0 0 0 0 0 0 0

Z -129.6601 -124.6601 -119.6601 -114.6601 -109.6601 -104.9638 -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 ed 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 58

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 ed 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: 59

# Profile 1 X 275 275.899 276.689 277.6571 300.2754 323.6533 341.0639 353.9198 363.2347 369.43 372.7376 373.1085 371.1548 366.3716 362.0027 366.9361 369.2318 368.8591 365.7523 359.6106 350.7593 338.2689 321.3768 298.8742 277.4731 276.5684 275.8225 275

Y 0 -0.1864 0 0.9718 53.4761 119.0809 178.6069 235.2993 290.3198 344.3058 397.6074 450.4228 502.8609 554.9747 557.8342 505.9475 453.7542 401.2081 348.21 294.5772 239.9852 183.8454 125.1295 60.8211 10.1544 8.1688 6.5400 0

Z 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0

#Profile 2 X 275

Y 0

Z 104.000 60

275.899 276.689 277.6571 300.2754 323.6533 341.0639 353.9198 363.2347 369.43 372.7376 373.1085 371.1548 366.3716 362.0027 366.9361 369.2318 368.8591 365.7523 359.6106 350.7593 338.2689 321.3768 298.8742 277.4731 276.5684 275.8225 275

-0.1864 0 0.9718 53.4761 119.0809 178.6069 235.2993 290.3198 344.3058 397.6074 450.4228 502.8609 554.9747 557.8342 505.9475 453.7542 401.2081 348.21 294.5772 239.9852 183.8454 125.1295 60.8211 10.1544 8.1688 6.5400 0

103.859 103.736 103.584 99.308 93.051 86.795 80.538 74.282 68.026 61.769 55.513 49.256 43.000 43.000 49.256 55.513 61.769 68.026 74.282 80.538 86.795 93.051 99.308 103.584 103.736 103.859 104.000

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.

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Setting the topology for the mesh grid

Adjusting the control points at the Leading Edge & Trailing Edge

62

3-D view of impeller without shroud surface

63

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. 65

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 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

1

Number of Number of

Minimum

Nodes

Skew Angle Skew Angle

3DIMPELLER 19380

Elements 16352

Maximum

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 66

course of the solution. A sensible initial guess for the temperature field is an 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 67

5. Rotation Axis

: Z

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 interface used to set attributes for a CFD calculation, Control the CFXSolver 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 ; this is not the same as the High Resolution advection scheme. This setting is more accurate but less robust. Values between 0.0 and 68

1.0 blend First and Second Order differencing, with increased accuracy and reduced robustness as you approach 1.0. At the higher values overshoots and undershoots can appear, at lower values excessive 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 converged steady state solution is obtained, or,

•

To evolve the solution through time in order to provide transient information 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 conservative 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 Residual Target value. For the present simulation Solver Parameters are specified as follows:

•

Advection scheme

•

Time Scale Control

:Physical Time Scale (0.0003 sec)

•

Maximum Iterations

: 200

•

Residual Convergence criteria

:Specified Blend Factor (0.75)

: RMS 69

•

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

70

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. 71

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.

72

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: 1)

Vector Plots

73

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 increases for various flow rates.

74

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.

75

8.3RESULTS 8.3.1Values obtained from CFX for 980 rpm

% flow %

flow per age, cu.m

Flow coeff

exit total pressure (pa)

Pr pre .statrise(pa) ic head coeff

flow angle alCu2 total efficiency 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

8.3.2Values obtained from CFX for 1450 rpm

76

% flow %

flow per age, cu.m

Flow coeff

exit total pressure (pa)

Pr pre .statrise(pa) head ic coeff

total efficien-flow angle alCu2 cy 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

8.3.3Values obtained from CFX for 2000 rpm

77

% flow %

flow per age, cu.m

Flow coeff

exit total pressure (pa)

Pr pre .statrise(pa) head ic coeff

total efficien-flow angle alCu2 cy 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

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

78

contour plots representing pressure, velocity and relative velocity at various speeds of 980, 1450 and 2000 RPM and for various flow rates.

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

79

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 80

Static pressure:-

Ps for 980 RPM at 70% flow

Ps for 980 RPM at 100% flow

Ps for 980 RPM at 130% flow

81

Pt at 980 RPM at 100% flow

Pt at 1450 RPM at 100% flow

Pt at 2000 RPM at 100% flow

82

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

83

Velocity at 1450 RPM at 100% flow

Velocity at 2000 RPM at 100% flow

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

84

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

85

BLADE LOADING At 1450 RPM and 70% flow :-

At 1450 RPM and 100% flow :-

At 1450 RPM and 130% flow :-

86

At 980 RPM and 70% flow :-

At 980 RPM and 100% flow :-

87

At 980 RPM and 130% flow :-

At 2000 RPM and 70% flow :-

88

At 2000 RPM and 100% flow :-

At 2000 RPM and 130% flow :-

89

8.5 GRAPHS

PRESSURE RISE VS MASS FLOW

90

TOTAL PRESSURE VS MASS FLOW RATE

91

HEAD COEFFICIENT VS MASS FLOW RATE

92

TOTAL EFFICIENCY VS MASS FLOW

93

SHAFT POWER VS MASS FLOW (AT 980 RPM)

FLOW COEFFICIENT VS HEAD COEFFICIENT

94

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 ed. 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 age.

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.

95

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 Deg Radial Turbo Machines

d) Pfleiderer

Die kreisel pumpen

e) Wikipedia.org

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