Almost Final Draft Project

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Design of Coolant Re-circulating Pump used in Nuclear Power Plants

A

PROJECT

ON

DESIGN OF COOLANT RECIRCULATING PUMP

USED IN NUCLEAR POWER PLANTS

Submitted

in partial fulfillment of

the requirements of the degree of

B.TECH. IN MECHANICAL ENGINEERING

By

AJINKYA A. PARAB (081020070)

BHARAT SUBRAMONY (081020067)

NAKUL R. GAWATRE (081020215)

2011-2012

Under the guidance of

PROF. S.M. GUNADAL

Department of Mechanical Engineering

Veermata Jijabai Technological Institute

Mumbai 400 019


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CERTIFICATE

This is to certify that Ajinkya A. Parab, Bharat Subramony and Nakul R. Gawatre students of

B.Tech (Mechanical engineering), have completed the thesis entitled, DESIGN OF

COOLANT RECIRCULATING PUMP USED IN NUCLEAR POWER PLANTS to our

satisfaction.

Prof. S.M. GUNADAL

Project Guide

Department of Mechanical engineering

Dr. M.A. Dharap

The Head,

Department of Mechanical engineering


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CERTIFICATE

The thesis, DESIGN OF COOLANT RECIRCULATING PUMP USED IN NUCLEARPOWER PLANTS submitted by Ajinkya A. Parab, Bharat Subramony and Nakul R.

Gawatre is found to be satisfactory and is approved for the Degree of Bachelors in

Technology in Mechanical engineering.

Prof. S.M. GUNADAL

Project Guide

Department of mechanical engineering

Examiner

Date: Place:


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Declaration of the Students

We declare that this written submission represents our ideas in our own words and where

others' ideas or words have been included, we have adequately cited and referenced theoriginal sources.

We also declare that we have adhered to all principles of academic honesty and integrity and

have not misrepresented or fabricated or falsified any idea / data / fact / source in our

submission.

We understand that any violation of the above will be cause for disciplinary action by the

Institute and can also evoke penal action from the sources which have thus not been properly

cited or from whom proper permission has not been taken when needed.

Signature of the students

Ajinkya A. Parab-(081020070)

Bharat Subramony-(081020067)

Nakul R. Gawatre-(081020215)

Date: Place:


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List of symbols

A area of flow, cross section

bi width of impeller at a given section i

c absolute velocity

cs flow velocity in suction nozzle

D,d diameter

db average diameter of impeller at inlet or outlet

dn

hub diameter

ds shaft diameter

dD shaft diameter at seal

e vane/blade thickness

F force

Fax axial force

Hth theoretical manometric head

Hth theoretical manometric head for infinite number of blades

i angle of incidence

kn blockage factor due to hub

n, N rotational speed in rpm.

nq,ss specific speed, suction specific speed

P (without subscript) power

p (with subscript) pressure

Q volumetric flow rate

Qla volumetric flow through impeller

q*

volumetric flow rate at best efficiency point (q*=Q/Qopt)

R, r radius


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s clearance

sax axial clearance between shroud of impeller and volute casing

tax sidewall clearance behind back-shroud of impeller

U, u circumferential velocity

V volume

Zla number of impeller blades

angle between relative velocity and negative circumferential velocity

specific weight of liquid or slip factor

angle in polar co-ordinate system

sp wrap angle for the volute or the blade

efficiency

angle between impeller vane and back shroud

density in kg/m3

shear stress, blade blockage factor

flow co-efficient

pressure co-efficient

blade loading


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List of subscripts, superscripts and abbreviations

1 impeller inlet leading edge (low pressure)

2 impeller outlet trailing edge (high pressure)

ax axial

a, m, i outer, meridional, inner streamline

B blade angle after all corrections

BEP best efficiency point

hyd hydraulic

La impeller

m meridional component

max maximum

mano manometric

mech mechanical

min minimum

opt optimum (at best efficiency point)

o/a overall

ref reference value

r radial

s suction, shaft

sp volute

th theoretical

vol volumetric

u circumferential component


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List of figures

Figure 1.1 Working of a general nuclear power plant Page 14

Figure 1.2 Working of a Pressurised Water Reactor Page 3

Figure 2.1 Different pumps used in a reactor circuit Page 8

Figure 4.1 Dimensioning of mixed-flow impeller Page 17

Figure 4.2 Inlet velocity triangle Page 21

Figure 4.3 Outlet velocity triangle Page 21

Figure 4.4 Blade cross-section Page 23

Figure 4.5 Point-by-point method of plotting blade Page 24

Figure 5.1 Blade loading Page 28

Figure 6.1 Types of volute sections Page 50

Figure 6.2 Volute axis profile Page 51

Figure 6.3 Casing thickness Page 31

Figure 7.1 Free body diagram of impeller Page 36

Figure 8.1 Gaspac T type mechanical seal Page 40

Figure 8.2 Gaspac L type mechanical seal Page 41


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List of tables and graphs

Table 4.1 Blade geometry on inner and outer streamlines Page 25

Table 5.1 Blade thickness calculation Page 28Table 6.1 Volute cross-section geometry Page 33

Table 6.2 Stress calculation for volute Page 34

Table 6.3 Materials for volute casting Page 34

Graph 4.1 Pressure co-efficient v/s specific speed Page 16

Graph 4.2 Normalized suction specific speed v/s inlet blade angle Page 19


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Contents

1. Introduction ................................................................................................................................13

1.1 Scope of the project .......................................................................................................................13

1.2 General Working of a Nuclear Reactor ..........................................................................................13

1.3 Nuclear Reactor Types ...................................................................................................................15

1.3.1 Pressurized Water Reactor ..........................................................................................................15

1.3.1.1 Coolant .................................................................................................................................16

1.3.1.2 Moderator .............................................................................................................................17

1.3.1.3Advantages ........................................................................................................................ 18

1.3.1.4Disadvantages.................................................................................................................... 19

2. Arrangement of different pumps in PWR reactor ..................................................................21

2.1 Functions of different type of pumps in the reactor .......................................................................22

2.1.1 Primary System ...........................................................................................................................22

2.1.1.1 Primary Reactor Coolant Pump ................................................................................................22

2.1.1.2 High-pressure injection pump ..................................................................................................22

2.1.1.3 Filling pump ..............................................................................................................................22

2.1.1.4 Residual heat removal pump ...................................................................................................22

2.1.1.5 Containment spray pump.........................................................................................................22

2.1.1.6 Nuclear reactor component cooling water pump ....................................................................22

2.1.1.7 Seawater pump ........................................................................................................................23

2.1.1.8 Motor driven auxiliary feed water pump .................................................................................23

2.1.1.9 Turbine driven auxiliary feed water pump ...............................................................................23

2.1.2 Secondary system .............................................................................................................. 23

2.1.2.1 Feed water pump .....................................................................................................................23

2.1.2.2 Feed water booster pump ........................................................................................................24

2.1.2.3 Condensate booster .................................................................................................................24

2.1.2.4 Condensate booster pump .......................................................................................................24

2.1.2.5 Moisture separator drain pump ...............................................................................................24

2.1.2.6 Low pressure feed water pump ...............................................................................................24

2.1.2.7 Circulating water pump ............................................................................................................25

3. I.A.E.A. specifications for an ACR-1000 PWR .......................................................................27

4. Calculation of pump parameters ..............................................................................................29

4.1 Specific speed.................................................................................................................................29


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4.2 Efficiencies.....................................................................................................................................29

4.3 Shaft diameter ds ............................................................................................................................30

4.4 Semi-axial impeller design .............................................................................................................31

4.4.1 Pressure co-efficient ....................................................................................................................31

4.4.2 Impeller outlet diameter d2a .........................................................................................................32

4.4.3 Mean Impeller outlet diameter d2m & diameter of inner streamline d2i .......................................32

4.4.4 Impeller inlet velocity .................................................................................................................32

4.4.5 Flow co-efficient ...................................................................................................................334.4.6 Normalized suction specific speed & suction specific speed ...................................334.4.7 Outlet width ...........................................................................................................................344.4.8 Axial extension ()....................................................................................................................354.4.9 Blade angles at inlet and outlet of the impeller at outer, meridional and inlet streams...........36

4.4.10 Blade thickness e and angle of incidence i...........................................................................37

4.4.11 Blade leading edge profile.........................................................................................................37

4.5 Impeller blade shape (profile) ........................................................................................................38

5. Blade loading and stress analysis ..............................................................................................43

6. Volute casing design ...................................................................................................................46

6.1 Wrap angle of partial volutes .........................................................................................................47

6.2 Casing design flow rate QLe ...........................................................................................................47

6.3 Inlet velocity...................................................................................................................................47

6.4 Cutwater diameter dz ......................................................................................................................47

Shape of the volute cross sections ........................................................................................................48

6.5 Casing thickness .............................................................................................................................50

7. Determination of axial thrust ....................................................................................................54

8. Bearing and Seal selection .........................................................................................................57

8.1 Bearing specifications: ...................................................................................................................57

Selection of seals .................................................................................................................................58

8.2 Bi-directional face pattern: .............................................................................................................58

8.3 Uni-directional face pattern:...........................................................................................................59

8.4.1 Gaspac T...............................................................................................................................59

8.4.2 Gaspac L...............................................................................................................................60


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Chapter 1

Introduction


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1. Introduction

1.1 Scope of the project

The primary objective of the project is to design the primary coolant re-circulating

pump of a Pressurized Heavy Water Reactor. The project aims at designing the basic

components of a pump i.e., Impeller, shaft, blades, volute, bearing and seals. The design

procedures followed are mostly based on empirical relations (which are observed to be true in

a number of experiments performed with varying parameters) and geometrical constraints.

The project does not address the problems arising due to flow separation, eddy losses,

vibrations, etc. The pump is designed for specific operating conditions as referred from the

I.A.E.A manual.

1.2 General Working of a Nuclear Reactor

Large electrical generating plants which provide most of our electricity all work on the

same principle - they are giant steam engines. Power plants use heat supplied by a fuel to boil

water and make steam, which drives a generator to make electricity. A generating plant's fuel,

whether it is coal, gas, oil or uranium, heats water and turns it into steam. The pressure of the

steam spins the blades of a giant rotating metal fan called a turbine. That turbine turns the

shaft of a huge generator. Inside the generator, coils of wire and magnetic fields interact - and

electricity is produced.


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Figure 1.1

The reactor in a nuclear power plant does the same thing that a boiler does in a fossil

fuel plant - it produces heat. The basic parts of a reactor are the core, a moderator, control

rods, a coolant, and shielding. The core of a reactor contains the uranium fuel. For a light

water reactor with an output of 1,000 megawatts, the core would contain about 75 tonnes of

uranium enclosed in approximately 200 fuel assemblies.

The neutrons produced by fission are travelling at great speeds, and in most reactors,

are deliberately slowed down by a material known as a moderator. Slow neutrons are much

more likely, when they collide with the nuclei of U-235, to cause fission and keep the reaction

going. A moderator is composed of light atoms and the materials most commonly used are

carbon in the form of graphite, and water. For more precise control of the chain

reaction, control rods are inserted into the core of the reactor. Pushed in, they absorb neutrons

and slow down the reaction, when pulled out, they allow the reaction to speed up again. In

this way the chain reaction is controlled. Fissions occurring in the reactor generate an

enormous amount of heat. A liquid or gas coolant carries this heat away from the reactor to a

boiler where steam is made. Shielding, typically made of steel and concrete about two meters

thick, is an outer casing that prevents radiation from escaping into the environment.


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1.3 Nuclear Reactor Types

Many different reactor systems have been proposed and some of these have been

developed to prototype and commercial scale. Several types of reactor (Magnox, AGR, PWR,

BWR, CANDU and RBMK) have emerged as the designs used to produce commercial

electricity around the world. A further reactor type, the so called fast reactor, has been

developed to the full-scale demonstration stage.

1.3.1 Pressurized Water Reactor

Figure 1.2

Design of PWR

Nuclear fuel in the reactor vessel is engaged in a fission chain reaction, which

produces heat, heating the water in the primary coolant loop by thermal conduction through

the fuel cladding. The hot primary coolant is pumped into a heat exchanger called the steam

generator, where it flows through hundreds or thousands of tubes (usually 3/4 inch in
http://en.wikipedia.org/wiki/Nuclear_fuelhttp://en.wikipedia.org/wiki/Nuclear_chain_reactionhttp://en.wikipedia.org/wiki/Heat_exchangerhttp://en.wikipedia.org/wiki/Steam_generator_(nuclear_power)http://en.wikipedia.org/wiki/Steam_generator_(nuclear_power)http://en.wikipedia.org/wiki/Steam_generator_(nuclear_power)http://en.wikipedia.org/wiki/Steam_generator_(nuclear_power)http://en.wikipedia.org/wiki/Heat_exchangerhttp://en.wikipedia.org/wiki/Nuclear_chain_reactionhttp://en.wikipedia.org/wiki/Nuclear_fuel


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diameter). Heat is transferred through the walls of these tubes to the lower pressure secondary

coolant located on the sheet side of the exchanger where it evaporates to pressurized steam.

The transfer of heat is accomplished without mixing the two fluids, which is desirable since

the primary coolant might become radioactive. Some common steam generator arrangements

are u-tubes or single pass heat exchangers.

In a nuclear power station, the pressurized steam is fed through a steam turbine, which

drives an electrical generator connected to the electric grid for distribution. After passing

through the turbine the secondary coolant (water-steam mixture) is cooled down and

condensed in a condenser. The condenser converts the steam to a liquid so that it can be

pumped back into the steam generator, and maintains a vacuum at the turbine outlet so that

the pressure drop across the turbine, and hence the energy extracted from the steam, is

maximized. Before being fed into the steam generator, the condensed steam (referred to as

feed water) is sometimes preheated in order to minimize thermal shock.

Two things are characteristic for the pressurized water reactor (PWR) when compared

with other reactor types: coolant loop separation from the steam system and pressure inside

the primary coolant loop.

In a PWR, there are two separate coolant loops (primary and secondary), which are

both filled with de-mineralized or de-ionized water. A boiling water reactor, by contrast, has

only one coolant loop, while more exotic designs such as breeder reactors use substances

other than water for coolant and moderator (e.g. sodium in its liquid state as coolant or

graphite as a moderator).

The pressure in the primary coolant loop is typically 1516 Mega Pascal (150

160 bar), which is notably higher than in other nuclear reactors, and nearly twice that of a

boiling water reactor (BWR). As an effect of this, only localized boiling occurs and steam will

recondense promptly in the bulk fluid. By contrast, in a boiling water reactor the primary

coolant is designed to boil.

1.3.1.1 Coolant

Light water is used as the primary coolant in a PWR. It enters the bottom of the

reactor core at about 275 C (530 F) and is heated as it flows upwards through the reactor
http://en.wikipedia.org/wiki/Electrical_generatorhttp://en.wikipedia.org/wiki/Condenser_(heat_transfer)http://en.wikipedia.org/wiki/Breeder_reactorhttp://en.wikipedia.org/wiki/Pascal_(unit)http://en.wikipedia.org/wiki/Bar_(unit)http://en.wikipedia.org/wiki/Nuclear_reactorhttp://en.wikipedia.org/wiki/Waterhttp://en.wikipedia.org/wiki/Waterhttp://en.wikipedia.org/wiki/Nuclear_reactorhttp://en.wikipedia.org/wiki/Bar_(unit)http://en.wikipedia.org/wiki/Pascal_(unit)http://en.wikipedia.org/wiki/Breeder_reactorhttp://en.wikipedia.org/wiki/Condenser_(heat_transfer)http://en.wikipedia.org/wiki/Electrical_generator


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core to a temperature of about 315 C (600 F). The water remains liquid despite the high

temperature due to the high pressure in the primary coolant loop, usually around 155 bar

(15.5 MPa 153 atm, 2,250 psig).

Pressure in the primary circuit is maintained by a pressurizer, a separate vessel that is

connected to the primary circuit and partially filled with water which is heated to the

saturation temperature (boiling point) for the desired pressure by submerged electrical heaters.

To achieve a pressure of 155 bar, the pressurizer temperature is maintained at 345 C, which

gives sub-cooling margin (the difference between the pressurizer temperature and the highest

temperature in the reactor core) of 30 C. Thermal transients in the reactor coolant system

result in large swings in pressurizer liquid volume; total pressurizer volume is designed

around absorbing these transients without uncovering the heaters or emptying the pressurizer.

Pressure transients in the primary coolant system manifest as temperature transients in the

pressurizer and are controlled through the use of automatic heaters and water spray, which

raise and lower pressurizer temperature, respectively.

To achieve maximum heat transfer, the primary circuit temperature, pressure and flow

rate are arranged such that sub-cooled nucleate boiling takes place as the coolant passes over

the nuclear fuel rods.

The coolant is pumped around the primary circuit by powerful pumps, which can

consume up to 6 MW each. After picking up heat as it passes through the reactor core, the

primary coolant transfers heat in a steam generator to water in a lower pressure secondary

circuit, evaporating the secondary coolant to saturated steam in most designs 6.2 MPa

(60 atm, 900 psia), 275 C (530 F) for use in the steam turbine. The cooled primary

coolant is then returned to the reactor vessel to be heated again.

1.3.1.2 Moderator

Pressurized water reactors, like all thermal reactor designs, require the fast fission

neutrons to be slowed down (a process called moderation or thermalization) in order to

interact with the nuclear fuel and sustain the chain reaction. In PWRs the coolant water is

used as a moderator by letting the neutrons undergo multiple collisions with light hydrogen
http://en.wikipedia.org/wiki/Bar_(unit)http://en.wikipedia.org/wiki/Bar_(unit)http://en.wikipedia.org/wiki/Megapascalhttp://en.wikipedia.org/wiki/Atmosphere_(unit)http://en.wikipedia.org/wiki/Psighttp://en.wikipedia.org/wiki/Pressurizerhttp://en.wikipedia.org/wiki/Nucleate_boilinghttp://en.wikipedia.org/wiki/Watt#Megawatthttp://en.wikipedia.org/wiki/Psiahttp://en.wikipedia.org/wiki/Thermal_reactorhttp://en.wikipedia.org/wiki/Neutron_moderatorhttp://en.wikipedia.org/wiki/Neutron_moderatorhttp://en.wikipedia.org/wiki/Thermal_reactorhttp://en.wikipedia.org/wiki/Psiahttp://en.wikipedia.org/wiki/Watt#Megawatthttp://en.wikipedia.org/wiki/Nucleate_boilinghttp://en.wikipedia.org/wiki/Pressurizerhttp://en.wikipedia.org/wiki/Psighttp://en.wikipedia.org/wiki/Atmosphere_(unit)http://en.wikipedia.org/wiki/Megapascalhttp://en.wikipedia.org/wiki/Bar_(unit)


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atoms in the water, losing speed in the process. This "moderating" of neutrons will happen

more often when the water is denser (more collisions will occur). The use of water as a

moderator is an important safety feature of PWRs, as an increase in temperature may cause

the water to expand, giving greater 'gaps' between the water molecules and reducing the

probability of thermalization; thereby reducing the extent to which neutrons are slowed down

and hence reducing the reactivity in the reactor. Therefore, if reactivity increases beyond

normal, the reduced moderation of neutrons will cause the chain reaction to slow down,

producing less heat. This property, known as the negative temperature coefficient of

reactivity, makes PWR reactors very stable. This process is referred to as 'Self-Regulating',

i.e. the hotter the coolant becomes, the less reactive the plant becomes, shutting itself down

slightly to compensate and vice versa. Thus the plant controls itself around a given

temperature set by the position of the control rods.

1.3.1.3 Advantages

1. The water, which is used as coolant, moderator and reflector, is cheap and available inplenty.

2. The reactor is compact and has high power density (65 KW/ liter).3. PWR reactors are very stable due to their tendency to produce less power as temperatures

increase; this makes the reactor easier to operate from a stability standpoint.

4. It allows to reduce the fuel cost extracting more energy per unit weight of fuel as PWR isideally suited to the utilization of fuel designed for higher burn-ups.

5. PWR turbine cycle loop is separate from the primary loop, so radioactive materials do notcontaminate water in the secondary loop.

6. PWRs can passively scram the reactor, in the event that offsite power is lost, toimmediately stop the primary nuclear reaction. The control rods are held by

electromagnets and fall by gravity when current is lost; full insertion safely shuts down

the primary nuclear reaction.
http://en.wikipedia.org/wiki/Temperature_coefficienthttp://en.wikipedia.org/wiki/Temperature_coefficient


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1.3.1.4 Disadvantages

1. The coolant water must be highly pressurized to remain liquid at high temperatures. Thisrequires high strength piping and a heavy pressure vessel and hence increases construction

costs. The higher pressure can increase the consequences of a loss-of-coolant accident. The

reactor pressure vessel is manufactured from ductile steel but, as the plant is operated,

neutron flux from the reactor causes this steel to become less ductile. Eventually

the ductilityof the steel will reach limits determined by the applicable boiler and pressure

vessel standards, and the pressure vessel must be repaired or replaced. This might not be

practical or economic, and so determines the life of the plant.

2. Following shutdown of the primary nuclear reaction, the fission products continue togenerate decay heat at initially roughly 7% of full power level, which requires 1 to 3 years

of water pumped cooling. If cooling fails during this post-shutdown period, the reactor

can still overheat and meltdown.

3. Additional high pressure components such as reactor coolant pumps, pressurizer, steamgenerators, etc. are also needed. This also increases the capital cost and complexity of a

PWR power plant.
http://en.wikipedia.org/wiki/Loss-of-coolant_accidenthttp://en.wikipedia.org/wiki/Loss-of-coolant_accidenthttp://en.wikipedia.org/wiki/Ductilityhttp://en.wikipedia.org/wiki/Ductilityhttp://en.wikipedia.org/wiki/Decay_heathttp://en.wikipedia.org/wiki/Decay_heathttp://en.wikipedia.org/wiki/Ductilityhttp://en.wikipedia.org/wiki/Loss-of-coolant_accident


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Chapter 2

Arrangement of

different pumps

in a reactor


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2. Arrangement of different pumps in PWR reactor

Figure 2.1


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2.1 Functions of different type of pumps in the reactor

2.1.1 Primary System

2.1.1.1 Primary Reactor Coolant Pump

Pump for continuous re-circulation of nuclear reactor coolant (light water) to the

steam generator to cool the thermal energy generated by the nuclear power

2.1.1.2 High-pressure injection pump

At the time of the trouble of the primary cooling water loss or when the main

steam pipe is broken, the reactor core is urgently cooled and boric acid is urgently

injected. As a result, the reactor core is cooled and the fuel rod is kept in the

allowable temperature.

2.1.1.3 Filling pump

This pump supplies make-up water to maintain the volume when the primary

reactor coolant has contracted due to decrease of load. Also it supplies sealing

water to the sealed part of the primary reactor coolant pump.

2.1.1.4 Residual heat removal pump

This pump removes the heat of the primary reactor coolant after the reactor is

stopped and lowers its temperature. Also, it prevents expansion of the trouble at

the time of the trouble of the coolant loss by injecting boric acid of the fuel

replacement water pump into the reactor core.

2.1.1.5 Containment spray pump

This pump has the role of minimizing the leakage of the radioactive material fromthe containment vessel at the time of the accident of nuclear power. At the time of

such accident the pump sprays boric acid water into the vessel and eliminates

contamination of the fissionable material.

2.1.1.6 Nuclear reactor component cooling water pump

This pump has the intermediate characteristics of two lines, the primary reactor

coolant and the seawater pump, and even if the primary reactor coolant leaks, it

prevents radioactive water from being discharged outside.


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2.1.1.7 Seawater pump

This pump sends water to each of the following equipment and cools them.

1. Nuclear reactor component cooler2. Diesel generator3. Refrigerating machine for air-conditioning

2.1.1.8 Motor driven auxiliary feed water pump

This is the horizontal type turbine driven pump for sending water of the

condensing tank (or the de-aerator storage tank) to the steam generator to remove

the heat of nuclear reactor cooling line when starting or stopping. Since it is

necessary to supply feed water even when electricity stops, two pumps are

provided; one is the motor driven pump that can be operated from the electric

source of the diesel generator and the other is the turbine driven pump that can be

operated with main steam.

2.1.1.9 Turbine driven auxiliary feed water pump

This is the horizontal type turbine driven pump for sending water of the

condensing tank (or the de-aerator storage tank) to the steam generator to remove

the heat of nuclear reactor cooling line when starting or stopping. Since it is

necessary to supply feed water even when electricity stops, two pumps are

provided; one is the motor driven pump that can be operated from the electric

source of the diesel generator and the other is the turbine driven pump that can be

operated with main steam.

2.1.2 Secondary system

2.1.2.1 Feed water pump

This pump takes from the de-aerator storage tank feed water pressured up by the

booster pump and pushes it into the steam generator through the high-pressure

heater. Accordingly, the main feed pump must be high temperature and high-

pressure pump since it requires the head larger than the pressure inside the steam


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generator. Also, from the viewpoint of assuring feed water of the steam generator,

this is one of the important auxiliary machinery of the plant.

2.1.2.2Feed water booster pump

This pump takes feed water from the de-aerator storage tank and increases NPSHa

(net positive suction head). Namely, the purpose of installation of this pump is to

assist suction of the main feed water pump. In the case of the plant in which the

de-aerator is not installed, the feed water booster pump is not required and not

installed since the head of the condensate pump is designed very high.

2.1.2.3 Condensate booster

This booster takes condensate from the condenser hot well and sends it to the

condensate booster pump through the grand condenser and the condensate de-

mineralizer.

2.1.2.4 Condensate booster pump

The purpose of installation of this pump is to pressure up the condensate sent from

the condensate pump and to send it to the de-aerator through the low-pressure

heater. This pump is not installed in the plant in which the condensate de-

mineralizer is not provided.

2.1.2.5 Moisture separator drain pump

This pump is installed to suction seawater from the intake, to exchange heat while

passing through the condenser and to discharge water to sea from the water

discharge outlet. Usually two units are installed and there is no spare. It is

determined from the viewpoint of the construction cost and economy whether the

pump should be of the fixed pitch vane type or the variable pitch vane type.

2.1.2.6 Low pressure feed water pump

This pump has the hydraulic variable pitch vane structure. Its merits are smaller

consumption of motive power and smooth start of operation (start of operation

with the vane fully closed). Whether the pump should be of the fixed pitch vane

type or the variable pitch vane type is determined from the viewpoint of

construction cost and economy.


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2.1.2.7 Circulating water pump

The steam discharged from the high-pressure turbine is condensed to drain. Drain,

after collected in the drain tank, is sent to the de-aerator by the moisture separator

drain pump.


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Chapter 3I.A.E.A.

Specifications


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3. I.A.E.A. specifications for an ACR-1000 PWR1. Circulation type : Forced2. Pump type : Vertical, Centrifugal, Single suction, Double

discharge

3. Number of pumps : 44. Pump speed : 1785 rpm5. Rated head : 226m6. Flow at rated head : 4.3 m3/s7. Pump suction pressure : 11.9 MPa8. Pump delivery pressure : 13.1 MPa9. Pump operating temperature: 275 C10. Motor : Squirrel cage Induction motor

Inferred parameters

Density of water at operating temperature is 760 kg/m3

[refer any steam table]


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Chapter 4Design of

Impeller


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4. Calculation of pump parameters4.1 Specific speed

Specific speed (nq) is defined as the speed of a geometrically similar

pump which consumes 1 kW and develops 1 m of total head, the pumping liquid being

water under normal temperature of 4C & at atmospheric pressure of 1.03325 bar.

Type of Impeller

Radial Semi-axialAxial

Specific Speed

nq= n *

1040 40 - 110 110 onwards

~ 22.5 ~ 1.151.5

~ 0.50.8

nq == = 63.5

As nq is in the range (40 rpm - 110 rpm) using, the pump impeller is of a mixed flow

& centrifugal type, single stage (helicoidal). Semi-axial impellers are superior to pumps with

radial impellers since the discharge flow from radial impellers becomes more non-uniformwith increasing specific speed due to the flow deflection in the flow in the meridional section

& rowing blade span. Hence the ratio d2/d1 = 1.3 to 1.5.

4.2 Efficiencies

Theoretically, all the energy supplied to the pump by the prime mover, in the form of

mechanical energy, should be converted into fluid energy. Owing to manufacturing

inaccuracies and entirely different flow conditions prevailing in pump, entire energy input

(mechanical energy) is not converted into fluid energy.


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mech= 0.900

Hydraulic efficiency (h) is the ratio between, actual head to the theoretical head.

hydr= 0.920

Volumetric efficiency (h) is the ratio between actual quantity and theoretical quantity.

vol= 0.950

overall = mech x hydr x vol = 0.787

4.3 Shaft diameter ds

Motor capacity [P] required is determined by the following equation,

[P] =

=

= 9.75 MW

Assuming an overload of 30% for safety, [P] = 9.75 x 1.3 = 12.68 MW.

The nearest motor capacity available is 14 MW manufactured by ABB, hence we proceed by

taking a motor of 14MW capacity and speed 1785 rpm.

The pump shaft is a stressed member for during operation it can be in tension,

compression, bending, and torsion. As these loads are cyclic in nature, the shaft failure is

likely due to fatigue. The shaft design depends on the evaluations of either the torsion shear

stress at the smallest diameter of the shaft or a comprehensive fatigue evaluation taking into

consideration the combined loads, the number of cycles, and the stress concentration factors.

Selecting shaft material as AISI-329 stainless steel having the following mechanical

properties:

1. Density = 7970 Kg/2. = 724 MPa3. = 560 MPa

Power (P) transmitted by shaft = 14 MW.


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T ==

=74.896 kNm

Hence

[] = T

[] = = 70 MPa [by M.S.S.T & FOS (factor of safety) =4]Therefore, ds min = 175.97 mm

As the application demands we shall use spline shaft of specification 20 x 180mm x 200mm

and stepped to 240mm at the conjunction with the mechanical seal.

4.4 Semi-axial impeller design

4.4.1 Pressure co-efficient

= = 0.742 [nq,ref =100]

Where is the pressure coefficient, the optimum value is selected using thepressure coefficient against the specific speed nq curve. Using this graph as a reference,

appropriate value of is selected. is a function of the specific speed & is alwaysbased on the outlet diameter at the outerstreamline.

Graph 4.1


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4.4.2 Impeller outlet diameter d2a

= 84.6 = /

= 827.15 mm

Using the above equation as obtained from Ref.No.1; the impeller outlet diameter is

calculated which also matches with that calculated from the equations given in Ref.No.2.

4.4.3 Mean Impeller outlet diameter d2m & diameter of inner streamline d2i

=

= = 0.9563

. =

=

. Figure 4.1

The above formulae are referred from Ref.No.1. pg 375.

4.4.4 Impeller inlet velocity

= 0.23 x [Ref.No.3: pg. 134]= 15.31 m/s


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4.4.5 Flow co-efficient = (0.18 to 0.27)( )

[Ref.No.1: pg. 375]

= 0.2542 [nq,ref = 200]

The impeller inlet properties are calculated according the required suction specific

speed. The blade inlet angles are generally in the range of 12 to 18. The flow co-efficient in

the design point may be selected using the above equation.

Also, =

= 59.4 m/s = 636.26 mm

Again, ()

Where is the hub diameter.

= 205.73 mmLet = 330 mm

Now, = 1.3, which should be in the range 1.15 to 1.5. Also hence the design is

correct.

The standard factor, ()

= 0.89544

4.4.6 Normalized suction specific speed & suction specific speed = = 233.106

Referring to the graph for normalized suction specific speed as a function of theapproach flow angle () at at the outer streamline (flow rate referred to flow rate atbest efficiency point i.e. ); nq,ref = 27. The following relation is applicable topumps with semi-axial impellers in the range nq = 10 to 160. The standard deviation is +14%.


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Graph 4.2

= (

)

233.106 =

(

)

[nq,ref = 27]

= 260Now,

=

= 14.26

This is verified from the Graph 4.2 for normalized suction specific speed as afunction of the approach flow angle () at at the outer streamline (flow rate referredto flow rate at best efficiency point i.e. ).

4.4.7 Outlet width With given values of outlet angle & the blade number, the head increases with the

outlet width & the Q-H curve becomes flatter. The outlet recirculation intensifies with

increasing & consequently, the power consumption rises as well.A sufficiently large outlet width is required to achieve a stable Q-H curve. Conversely,

the non-uniformity of the flow at the impeller outlet grows with the width of the impeller,

increasing the turbulent dissipation losses in the collector as well as pressure pulsations &

excitation forces. It should also be ensured to maintain

.


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The relative outlet width = is commonly selected from empirical data. Torender the impeller discharge flow at the best efficiency point as uniform as possible & to

avoid unnecessary turbulent dissipation losses, should be selected as low as ispermissible with respect to the stability of the Q-H curve, which is given by:

= ( )( ) ( )

[nq,ref= 100]

= 0.153 126.55 mm

Now, = = = 215.265 mm > Thus the design is acceptable.

4.4.8 Axial extension ()Favorable flow conditions are achieved by moving the leading edge forward into the

impeller eye. In this way low blade loadings and correspondingly moderate low-pressure

peaks are obtained and cavitation is reduced.

= ) ( )

[nq,ref=74]

= 162.05 mm

The selection of the blade number ZLa depends on various criteria:

1. To reduce pressure pulsations and hydraulic excitation forces,2. The hydrodynamic blade loading should be in an optimum range. If the loading is too

low, unnecessarily high friction losses must be expected. If the loading is too high, the

turbulent dissipation losses increase due to uneven flow distribution. The blade

loading can only be verified after completing the impeller design.

3. Less than 5 blades are unfavorable for high heads per stage since the impeller outletflow becomes very non-uniform over the circumference due to the large blade spacing.

The consequence would be unnecessarily high pressure pulsations, noise and

vibrations.


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4. In case of even number of blade i.e. 6 or 8, the wrap angle for the twin volute, wouldhave to be reduced from 180

to about 165

, in order to avoid the pulsating effect.

5. To avoid further complications in design, and reduce the effect of pulsations, wechoose the optimum number of blades to be,

Cm2 = 0.18 * = 12 m/s [For Km2, Ref.No.3, pg 375]

= =

= 77.30 m/s

= Hth ) = 226 x (1+

)

where = 0.55 + 0.6sin(2m) , r = r2m, z = Zla, Mst = 0.5 ( r22-r1

2)

[Plfeiderers semi-empirical relation]

= [Ref.No.3 pg.94]

= 38.07 m/s

= 39.23 m/s

Figure 4.2 Figure 4.3

4.4.9 Blade angles at inlet and outlet of the impeller at outer, meridional and inlet streams

The incidence should be selected between 0 and 4. The flow rate of shock-less entry

according to is thus commonly slightly above the best efficiency point. The highest

efficiencies are obtained when the blade outlet angle on the mean streamline is selected as

2B,m = 20 to 26.


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() = 17

; ; ; ;

With specific speeds above nq 40 to 60 the blade outlet angle should not be made

constant over the impeller outlet width. The outer streamline is unloaded so that

2B,a < 2B,m < 2B,i. The head is calculated on the basis of the mean angle 2B,m.

4.4.10 Blade thickness e and angle of incidence i

Requirements in terms of casting ability and mechanical strength determine the blade

thickness. Experience shows that these requirements are met when e/d2 = 0.016 to 0.022 is

selected. The upper range applies to high pressure impellers with more than 600 m of head per

stage. The lower range applies to low heads and low specific speeds. With increasing blade

width (i.e. also with growing specific speed) the blade stresses rise at given head; the blade

inlet angle is obtained by adding the incidence i1 to the flow angle 1 (with blade blockage)

from following equation.

( )-1

e = 0.022 x d2a= 18.85 mm

And 1B = 1 + i1

Substituting the values as derived and solving, by trial and error, we select the angle of

incidence to be 1 for

4.4.11 Blade leading edge profile

Unfavorable leading edge profiling generates local excess velocities and a

correspondingly intense low-pressure peak which impairs the cavitation behavior and may

even affect the efficiency. Designing the leading edge as a semi-circle is very unfavorable in


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this regard and only acceptable for small pumps or applications with low requirements.

Elliptical inlet profiles provide favorable pressure distributions.

Let e1denote the thickness of blade at inlet,

e1 = 0.75 * e = 14.14 mm

The blade leading edge and trailing edge profile width be e2= 0.5 x e in order to

reduce the width of the wake, turbulent dissipation losses and pressure pulsations

e1 = 14.14 mm and e2 = 7.07 mm Figure 4.4

4.5 Impeller blade shape (profile)

Two consecutive blades define the shape of an impeller passage. The length of the

blade and hence the passage length, can be different for same diameters d 1 and d2, the same

angles 1 and 2 and same number of blades. In short passages the angle of divergence may

be excessively large, which increases separation and forms harmful eddies. If the passages are

very long, and the angle of divergence small, the losses due to separation are reduced, while

the frictional losses are increased. The sum of the losses should be minimum for highest

efficiency. Hence it is necessary to make some compromises. We apply the point by point

method of drawing a single curvature blade.

It is based on the assumption that the transition of 1 into 2 depends on the radius rand the central angle, for a given r. The values of r and constitute polar co-ordinates

for a given point on the blade. After determining a series of points, a smooth curve is drawn

through them giving the central line of the blade.


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Let us consider an elementary right angle triangle as shown. For an infinitesimally

small angle , it can be derived that,

Figure 4.5

The following table, Table 4.1 constitutes different radii at its corresponding central angle.

The different velocity components are also computed.


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Sample calculation

The values of r, and Cm are linearly interpolated as initial and ending values are

known. Interpolation ensures that the flow velocity changes gradually with minimum losses.

The rest of the values are computed as follows.

Considering the 3rd

point of outer stream:

1. u= = 63 m/s2. tan() =

tan(14.707) =

wu = 55.38 m/s

3. cu = uwu = 6355.38 = 7.614 m/s4. w = = 57.2645 m/s5. 6. In = = = 11.30377. In-1 = = = 11.838. = 0.0095 ( 11.3037 + 11.83)/2 = 0.109889. Cumulative sum of upto the current location = 0.2248910.

The point 3 is plotted at a distance of 0.337m and at an angle of 12.883 with the reference

axis.

The other points are plotted accordingly.


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Chapter 5

Blade analysis


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5. Blade loading and stress analysisThe blades are supported at one end, fixed to the back shroud, the blades act as

cantilever beams under a uniformly distributed load starting from the inlet of the blade at

static pressure of 11.9 MPa and goes up to the outlet of the blade at 13.1 MPa. The blade can

thus be divided into equally spaced congruent beams by forming suitable number of elements.

For sake of convenience, we shall divide the entire blade into 10 equal parts.

Assuming the pressure is uniformly increasing over the face of the blade, the

elemental blade can be assumed to be under an average constant load. Hence we shall now

proceed to test the previously calculated blade thickness for its strength under the moment of

the force. The length of the blade is found to be 429.95 mm at the inner streamline and 261.43

mm at outer streamline. Refer figure 5.1.

Elemental dimensions

1. Cross-section of cantilever beam = 42.995(b) x 18.85(d) x 26.14(h) mm2. Length of the beam at exit of the impeller (2x)= 130.14 mm3. Effective force on the blade F = (13.04 -11.96) N/mm2 x (130.14 x 26.14) = 3674

N

4. Moment of force M = F xx = 239068.6 Nmm5. Section modulus for the beam Z = bd2/6 = 2546.18 mm36. Bending stress induced b = = 93.89 N/mm2

Based on the above calculations, we can see that the blade thickness is sufficient to

sustain the bending moment due to the pressure developed. However, we shall formulate a

table to determine the thickness required at different points on the blade to sustain the bending

moment, as shown in Table 5.1. The tensile strength of the Stainless Steel is 140 N/mm2.

Let the section modulus be Z = bd2/6, where b = 42.995mm. Now for the beam to have

sufficient strength to sustain the bending moment,

= = 4.201 mm


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Table 5.1

l1 l2 h Area avg p force x M d min edesign

125.66 134.62 26.14 3401.86 0.08 272.149 65.070 17708.7 4.201 7.07

134.62 143.58 26.14 3636.09 0.18 654.496 69.550 45520.5 6.735 10.99

143.58 152.54 26.14 3870.32 0.3 1161.09 74.030 85956.6 9.255 14.92

152.54 161.50 26.14 4104.54 0.42 1723.9 78.510 135345.7 11.614 18.85

161.50 170.46 26.14 4338.77 0.54 2342.93 82.991 194443.1 13.921 18.85

170.46 179.42 26.14 4573.00 0.72 3292.56 87.471 288005.0 16.942 18.85

179.42 188.38 26.14 4807.23 0.6 2884.33 91.951 265219.6 16.258 18.85

188.38 197.34 26.14 5041.45 0.48 2419.9 96.431 233355.5 15.250 18.85

197.34 206.30 26.14 5275.68 0.36 1899.24 100.912 191657.0 13.821 16.49

206.30 215.26 26.14 5509.91 0.24 1322.37 105.392 139368.7 11.785 14.14

mm mm mm mm2

N/mm2

N mm N-mm thk of blade (mm)

Thus it is clear that the design by using the empirical formulae provides feasible

results in practice.

Figure 5.1

A

x

P

d

b


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Chapter 6

Design ofVolute Casing


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6. Volute casing designIn the volute casing the kinetic energy present at the impeller outlet is converted into

static pressure with as few losses as possible. The volute casing then directs the fluid into the

discharge nozzle or, in multistage pumps, into the following stage.

Single volutes are the least expensive solution in terms of manufacturing costs; they

are best accessible for the dressing of the cast channels. The radial forces generate bearing

loads, bending stresses in the shaft and shaft deflection which can threaten the reliability of

the machine. The heads up to which single volutes can be sensibly used depends on the

specific speed, the design of the pump (especially of the bearing housing), the shaft thickness

and the bearing span or the overhung of the impeller. When pumping water, the limit is about

Hopt = 80 to 120 m at nq < 40. At high specific speeds a double volute may be indicated

already for heads above Hopt = 60 to 80 m. When pumping liquids with densities considerably

lower than water, the limit is correspondingly higher since the radial forces are proportional to

gH.

Double volutes are employed when bearing loads, shaft stresses and shaft deflection

become impermissibly high without measures for radial thrust reduction so that their control

would require an excessively high design effort and cost. The rib between the inner volute and

the outer channel also reduces the casing deformation under internal pressure, which

facilitates the casing design at high specific speeds.

Twin volutes differ from double volutes mainly in that both partial volutes end in

separate channels and not in a common discharge nozzle as is the case with double volutes.

Apart from special designs they are found in multistage volute pumps or vertical pumps

where the partial volutes end in a central column pipe.

Hence we shall design a double discharge twin volute, to suit our requirements of low

velocity of discharge, and high capacity.


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Parameters to be considered for designing the volute

6.1 Wrap angle of partial volutes

In the case of double or twin volutes the wrap angle of the partial volutes generally

amounts to sp = 180. In this case an even impeller blade number should be avoided in order

to reduce pressure pulsations.

6.2 Casing design flow rate QLe

To ensure that the actual best efficiency flow coincides with the design flow rate, the

volute must be designed for Qopt. The design flow rate must be augmented by possible

leakages which may flow through the volute, note that the leakage losses at the impeller inlet

do not flow through the volute.

QLe = Qopt(1.05 ~ 1.10) [Assuming 5 ~ 10 % leakage losses]

6.3 Inlet velocity

The circumferential component of the absolute velocity at the impeller outlet is

calculated. Downstream of the impeller it develops in accordance with the conservation of

angular momentum as per c3u =

. With some types of pumps, a diffuser or stay vanes can

arranged between the impeller and the volute. In this case the circumferential velocity c4u at

the outlet of these components must be used as inlet velocity to the volute.

6.4 Cutwater diameter dz

Between impeller and cutwater a minimum clearance (gap B) must be maintained to

limit pressure pulsations and hydraulic excitation forces to allowable levels. The ratio of the

cutwater diameter dz* = is calculated from the formula below.

[Ref.No.1. Pg.567]


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= 1.20 [nq,ref=40; ref=1000 kg/m3; Href= 1000m]

dz = 1.0201 m [some extra allowance is added for alignment of impeller inside the

casing]

Shape of the volute cross sections

The shapes must be selected so as to suit the pump type, taking into account casing stresses

and deformations as applicable. Figure shows some design options for the volute cross

section. When designing the casing, also the requirements of economic manufacture of

patterns and castings must be taken into account. With the rectangular and trapezoidal basic

shapes, for instance, all corners must be well rounded for casting reasons.

Figure 6.1

In practice it is found that constant velocity design gives higher efficiency than free

vortex design for pumps and vice versa for hydraulic turbines, due to increase in area of flow


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Constant velocity of flow through all volute sections CVis determined as CVR = Cu2R2, where

R is the radius of the centre of gravity of the last volute cross-section.

The moment of all surface forces acting on the elementary volume of fluid in the spiral casingis zero i.e., Mz = 0.

Thus the moment of momentum remains constant in the elementary fluid

With the increase in radius R in spiral passage the tangential velocity decreases,

correspondingly the pressure energy increases.

Applying equation for a circular cross section volute design with Cu,r=constant.

) for 180 half of a twin voluteWhere Mm = Cu3r2 = 19.4176 m

2/s and Q = 2.278 m

3/s

Uncorrected radius of volute is given by,

= where C =

= 9640.82 units

At =15, = = 0.037431 m = 37.431 mm

And correction due to hydraulic resistance is given by,

= = 0.0008571 m = 0.8571 mm

The total radius of the circular cross-section of the volute is therefore,

tot= + = 38.2883 mm


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So we can now substitute different values of angle and generate a table for interpolated

values and then interpolate the steps by drawing the curve path using a suitable CAD package.

Figure 6.2

Table 6.1

final = +

15 37.431 0.857 38.2883

30 53.847 1.714 55.5613

45 66.805 2.572 69.3769

60 77.974 3.429 81.4027

75 87.999 4.286 92.2848

90 97.211 5.143 102.3545

105 105.808 6.001 111.8085

120 113.917 6.858 120.7751

135 121.628 7.715 129.3434

150 129.006 8.572 137.5783

165 136.099 9.429 145.5286

180 142.946 10.287 153.2324

Degrees Mm mm mm

6.5 Casing thickness

The pressure inside the volute casing is 13.1 MPa, hence we shall proceed to

determine the casing thickness by principle for thick cylinder by Lamis Equation. According

to Lamis equation, there will be two important principle stresses for every elemental ring.

a. Radial pressurepxb. Hoop stressfx

Radius final

rz+final


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Where by derivation, it is seen that

Radial pressurepx =

And Hoop stress fx =

The parameterx will vary from r1to r2 across the thickness of the cylinder r2 - r1 = tcylinder.

Herepx at r2 is atmospheric pressure andpx at r1 is 13.1 MPa i.e. 13.1 N/mm2

By computing the value of constants a and b for different radii and the corresponding

thicknesses, we formulate a table to compute the value of hoop stresses at the radii r1and r2.

Figure 6.3

Table 6.2

thkcylinder fxat r1 fx at r2 b a

40 42.6654 16.4724 50470.7 8.24

45.4545 48.9235 22.7240 115961.0 11.36

50.9090 52.3242 26.1216 188984.9 13.06

56.3636 54.3018 28.1032 266725.2 14.05

61.8181 55.4968 29.2986 347871.5 14.65

67.2727 56.1979 30.0002 431607.2 15

72.7272 56.6002 30.3981 517550.1 15.2

78.1818 56.7746 30.5770 605120.5 15.29

83.6363 56.8058 30.6081 694381.9 15.3

89.0909 56.7610 30.5614 785143.7 15.28

94.5454 56.6421 30.4408 877262.4 15.22

100 56.4632 30.2643 970513.1 15.13

mm N/mm2

N/mm2

r2

r1

px

fx


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As the pump is of critical application, we shall select the casting material for the volute casing

to be High Grade Spheroidal Graphite Cast Iron 600/3 or 700/2, having mechanical properties

as follows.

Table 6.3

Material EN-GJS-600-3 EN-GJS-700-2

Tensile Strength N/mm2

600 700

Proof Stress N/mm2

370 420

Hardness BHN 190-270 230-300


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Chapter 7

Axial Thrust


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7. Determination of axial thrustThe unbalanced axial thrust on the back shroud is expressed by:

Where,

is the area corresponding to the diameter of the impeller wearing ring in mm2 is the area of shaft sleeve through the stuffing box, in mm2

is the head acting over the whole unbalanced area is the peripheral velocity at the impeller wearing ring diameter is the peripheral velocity at shaft sleeve diameter is the specific weight of the fluidA1 = 0.31769m

2at Dr = 0.636m HL = 160.95m Us = 22.43m/s at dD = 240mm

As = 0.04524m2 at Ds = 0.240m Ur = 93.46Dr = 7455.6 kg/m2s2The pressure distribution in the space between the impeller shrouds is based on the

assumption that the angular velocity of rotation of the liquid in this space is equal to one half

that of the impeller.

To balance the axial thrust , radial ribs are provided on the back shroud. With theseribs closely fitted to the casing walls the liquid will rotate approximately with full impeller

angular velocity. This will further reduce the pressure on the impeller back shroud over the

area , determined by the diameter of the radial ribs .


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The reduction of the axial forces on the back shroud is given by

For a complete balance, . From this relationship, the diameter of the radial ribs canbe determined. The number of ribs varies from four for small pumps to six for large pumps.

Frequently, radial ribs are used to reduce the pressure on the stuffing box.

First we consider a 100% balance of the unbalance using radial ribs,

By appropriate calculations, we get,

But as the Dr, is greater than the outer diameter of the impeller d2i, it is not geometrically

possible. Hence, we shall balance the axial thrust force partially, therefore for a 75% balance,

we get,

+ Fweight Axial load on the shaft due to static pressure acting on the hub is

Fax + Hence Fax = 412kN, where d = dhub = 0.21m Figure 7.1

Balancing the forces on a free body diagram Figure 6.1 of the impeller and shaft as a single

body,

Fres = Fax + Fhyd unbalance + Fweight = 412.17kN - 251.38 +188. 54 - 3.2kN = 346.13 kN

Therefore the resultant axial force has to be borne by the bearings.


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Chapter 8

Bearing

and

Seals


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8. Bearing and Seal selection

As the load on the shaft is purely axial in nature, and the operating temperature is

275C, it is only appropriate to choose heavy duty tapered roller thrust bearings for usage. We

referred to a commercial bearing manufacturer RBC bearings catalogue for the data, and

worked out the following calculation.

8.1 Bearing specifications:

1. Static Capacity Co = 4427000 lbf2. Dynamic capacity C = 1304000 lbf3. Inner diameter Di= 9.254. Outer cage diameter Do= 21.55. Width of bearing = 5

C = 1,304,000 lbf = 591484.45048 kgf

Load P = (XFr + YFa) Sfkt

= Y Fa Sf [Let Kt = 1.2 and Sf= 1, as it is heavy duty brg]

Fa = 346.13 kN

Therefore,

P = 1346.131031.2

= 415356 N

= 41535.6 kgf

C = P

L90 = (

= 3775.41 m-r

L90 =

Lh = 35251.28 hrs

= 1470 days = 4.1 years


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Selection of seals

The seal is subjected to following conditions and has the given dimensional

constraints:

1. Temperature = 275C2. Pressure = 2 x Pop = 262 bar3. Shaft outer diameter = 240 mm4. Speed at the surface of the shaft = 22 m/s

Considering the above requirements the most suitable seals are Gas-packed tandem sealswhich have the following features,

1. High pressure services with 425 bar (6200 psi) dynamic and static pressure ratings.2. Suitable for high speed applications with peripheral velocity up to 250 m/s and speeds

up to 50,000 rpm.

3. High reliability with more than five years means time between repairs or overhaul.4. Narrow and short axial designs for small seal chambers with greater ease of retrofit

and reduced influence on rotor dynamics.

5. Interchangeability seal parts and test fixture.6. Ease of serviceability.

The gas-packed tandem seals are designed using either the bi-directional T-Groove

technology or the Advanced Pattern Groove (AGP) technology. Both these proven lift-off

patterns have high film stiffness and damping capabilities that maintain the gas film under

slow roll conditions as well as high speeds up to 250 m/s.

8.2 Bi-directional face pattern:

The T- groove provides increased protection with unique bi-directional T-groove face

design and can operate in clockwise as well as counter-clockwise direction of rotation. This

attribute provides optimum protection from reverse rotation.

During shaft rotation in either direction, gas flows into the symmetrical T-grooves and

is pumped circumferentially toward the edge of the groove. Stagnation of the gas flow at the

edge builds pressure and results in hydrodynamic lift-off, even at low peripheral velocities.


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The unique design of the T-groove provides lower leakage than the conventional designs

patterns.

8.3 Uni-directional face pattern:

The APG incorporates specially designed tapered grooves that progressively more

shallow as they reach the circumferential groove. The APG design outperforms traditional

spiral groove designs allowing lift-off at lower speeds, low-pressure hydrostatic lift, and

better film stiffness performance.

During dynamic operation, any face pattern creates additional pressure to separate

faces so they are non-contacting. The APG extend farther across the face than conventional

uni-directional patterns, providing early lift-off and better performance at lower speeds. The

tapered groove depth of the APG design allows the faces to rapidly adjust, providing stable

operation during changing process conditions. The deeper grooves at the face periphery pump

the sealed medium toward the centre dam, developing pressure to cause hydrodynamic lift.

The APG operates with non-contacting seal faces, thus keeping parasitic horsepower

requirements low.

Considering the above requirements and features, there are two options available.

Their commercial names and specifications are as follows.

1. Gaspac T2. Gaspac L

8.4.1 Gaspac T

This tandem seal provides full pressure breakdown across the primary seal faces. The

secondary seal faces normally operate under low pressure. In the event of primary seal failure,

the secondary seal acts as an installed spare. The process gas has controlled leakage across

both sets of seal faces.

The standard operating limits are:

Pressure = 250 bar / 3600 psi


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Temperature = -100C to 230C Speed = 1 to 250 m/s Shaft size = 40 mm to 360 mm

Figure 8.1

8.4.2 Gaspac L

These are the tandem seals with inter-stage labyrinth used to eliminate process gas

leakage to the atmosphere. This is accomplished by introduction of an inert buffer gas to the

secondary seal. With slightly higher inert buffer gas pressure, labyrinth keeps the process gas

from migrating to the secondary seal faces. The inter-stage labyrinth provides a low pressure

solution to controlling emissions across a gas seal.

The standard operating limits are:

Pressure = 250 bar / 3600 psi Temperature = -100C to 230C Speed = 1 to 250 m/s Shaft size = 40 mm to 360 mm


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As both seals mentioned above have the same standard operating limits, the Gaspac L

type is selected considering its added feature that is the elimination of the process gas leakage

to the atmosphere.

Figure 8.2


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9. ConclusionThe coolant re-circulating pump is the heart of a nuclear reactor. The failure of the

pump can cause immediate shut down of the reactor or even a fatal accident. Keeping this in

mind, the pump is designed for a high factor of safety and longer life than usual, causing

some of the dimensions of components to look inflated. However, the final dimensions of the

pump obey the general thumb rules of hydraulic design. The geometrical model of the pump

shows no signs of interference. It is thus assured that the pump is geometrically correct. As far

as the achievement of actual discharge or head or practical application of the pump is

concerned; the model that has been generated may be analyzed in software like CFX. The

fluid dynamic analysis of the pump model by generation of dynamic mesh is beyond ourB.Tech. curriculum, which is why we have excluded it in this work.


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10. References1. Guilich, Johann Friedrich (2010), Centrifugal pumps, Springer Heidelberg Dordrecht,

London & New York.

2. Lobanoff, Val. S., Ross, Robert. R. (1992), Centrifugal Pumps Design andApplication, Gulf Publishing Company, Houston, Texas.

3. Lazarkiewicz, S. and Troskolanski, A. T. (1965), Impeller Pumps,ISBN 0080111726,Elsevier.

4. Status Report No.69, for Advanced CANDU Reactor, International Atomic EnergyAgency, P.O. Box 100, Wagramer Strasse 5, 1400 Vienna, Austria.

5. ACR-1000 Technical Description (January 2010), Atomic Energy of Canada Limited,2251 Speakman Drive, Mississauga, Ontario, Canada L5K 1B2.

6. RBC Bearings Catalogue.7. Flowserve Company Catalogue on Mechanical Seals.

Websites

www.google.com/images/

www.iaea.org
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