i

THE COPPERBELT UNIVERSITY

SCHOOL OF ENGINEERING

MECHANICAL ENGINEERING DEPARTMENT

FINAL YEAR PROJECT REPORT

INVESTIGATION ON LOW LIFT PUMPS FAILURES

(NWSC - Bulangililo Works)

BY

KANYIMBO PENJANI

MASENDEKE DARLINGTON M.

2011

2

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THE COPPERBELT UNIVERSITY

SCHOOL OF ENGINEERING

MECHANICAL ENGINEERING DEPARTMENT

FINAL YEAR PROJECT REPORT

INVESTIGATION ON LOW LIFT PUMPS FAILURES:

A CASE FOR

NKANA WATER AND SEWERAGE COMPANY

Bulangililo Works.

BY

KANYIMBO PENJANI (COMPUTER No: 07074485)

MASENDEKE DARLINGTON M.(COMPUTER No: 07030193)

“Report submitted in partial fulfilment of the requirement for the Degree of Bachelor of

Engineering, Copperbelt University.”

4

2011

DECLARATION

We declare that the work contained herein is our own work and all the works of other people

included in this project has been plainly and accordingly acknowledged and without any doubt,

this project has not been previously presented at this University or any other institution for

similar purposes.

This project was undertaken by:

KANYIMBO PENJANI MASENDEKE DARLINGTON M.

Signature:__________________ Signature:___________________

Date:______________________ Date:_______________________

Supervised by:

The Head of Mechanical Engineering Department

MR. MWABA C.

Signature:______________________

Date:__________________________

Co-Supervisor: External Supervisor

MR. KUCHEBA AM. MR. BUPE R. (NWSC)

Signature:______________________ Signature: _____________________

Date:__________________________ Date:_________________________

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ACKNOWLEDGEMENTS

We would like to whole heartedly acknowledge the help and support rendered to us through the

achievement of this project and the program at large. There are many names we would have

loved to mention but in honour of time and space, we would like to acknowledge the following;

Our supervisor; Mr. Mwaba C. co-supervisor; Mr. Kucheba AM. Mr. Njovu External supervisor;

Mr. Bupe R. Mr Mwenge R.and Mr. Yowela W. from Afred H knight,Mr. Mwaba B. Mr.

Kangwa, Mr.Simukonda D. Mr. Muwowo G. Mr. Lungu from NWSC Mr. Silwamba J. from

Mopani, Mr. Munsaka R. from Sulzer, Mr Langi E. Mr. Bwalya S.from SKF.

We would like to further acknowledge the help and support of our class mates EM 5 and EE 5,

Mr Masendeke A, Nanja A. the Kanyimbo stronghold, the CBU UCZ fellowship, Linah

Chelemu, Zyambo V.,Kachingwe M and many more.

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DEDICATION

KANYIMBO PENJANI

Coming to the end of this first phase of light is humbling and gives deep appreciations of what

worth people are. In the illumination of day and the uncertainty of night my life gives full

witness of how one cannot and should never think one was meant to make it independently. The

necessity of working with friends, family and individuals has modelled in my life simplification

and ability to see in midst of unsettled situations. I therefore, dedicate this work to the unborn,

born and unknown, born and known, dead and known, dead and unknown.

MASENDEKE DARLINGTON

To my beloved mum and Dad- you have always wished the best for me. This has come true. To

my beloved brother, Alvin. You have really been my mentor and my source of inspiration. I

appreciate all that you‟ve done for me. To my beloved late brother, Henry. I always miss you.

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TABLE OF CONTENTS DECLARATION ………………………………………………………………………………………………………………………………………..iv

ACKNOWLEDGEMENT……………………………………………………………………………………………………………………………..v

DEDICATION……………………………………………………………………………………………………………………………………………vi

LIST OF FIGURES…………………………………………………………………………………………………………………………….………..4

LIST OF TABLES…………………………………………………………………………………………………………………………………….

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CHAPTER ONE ....................................................................................................................................... 14

1 INTRODUCTION .............................................................................................................................. 14

1.2 PROBLEM STATEMENT ................................................................................................................. 15

1.3 AIM.............................................................................................................................................. 15

1.4 OBJECTIVES .................................................................................................................................. 15

1.5 PROJECT METHODOLOGY ............................................................................................................ 15

1.6 GANTT CHART .............................................................................................................................. 16

1.7 BACKGROUND ............................................................................................................................. 16

1.8 PUMP DETAILS ............................................................................................................................. 17

CHAPTER TWO ...................................................................................................................................... 20

PUMP DESCRIPTION .............................................................................................................................. 20

2.1 POSITVE DISPLACEMENT .............................................................................................................. 20

2.2 NON POSITIVE DISPLACEMENT PUMPS ........................................................................................ 21

2.2.1 CENTRIFUGAL PUMPS ............................................................................................................... 22

2.2.2 COMPONENTS OF A CENTRIFUGAL PUMP. ................................................................................ 23

2.2.2.1 SHAFT ................................................................................................................................ 24

2.2.2.2 WEARING RINGS ................................................................................................................ 26

2.2.2.3 IMPELLER ........................................................................................................................... 26

2.2.2.4 CASING............................................................................................................................... 29

2.2.2.5 BEARINGS........................................................................................................................... 31

2.2.2.6 LUBRICATION ..................................................................................................................... 35

2.2.2.7 SUCTION NOZZLE ............................................................................................................... 36

2.2.2.8 DISCHARGE NOZZLE ........................................................................................................... 36

2.2.2.9 STUFFING BOX.................................................................................................................... 37

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2.2.2.10 STUFFING BOX COVER ...................................................................................................... 37

CHAPTER THREE .................................................................................................................................... 38

CHARACTERISTICS AND OPERATIONS OF CENTRIFUGAL PUMPS............................................................. 38

3.1 PUMP PERFORMANCE CURVE ...................................................................................................... 39

3.1.1 NET POSITIVE SUCTION HEAD................................................................................................ 40

3.1.2 EFFECIENCY CURVE ............................................................................................................... 41

3.1.3 THE SYSTEM CURVE .............................................................................................................. 42

3.1.4 SELECTION OF PUMP ............................................................................................................. 45

3.1.5 PUMP OPERATING POINT ...................................................................................................... 46

3.2 TYPICAL FORCES ON IMPELLERS ................................................................................................... 47

3.2.1 AXIAL FORCE ......................................................................................................................... 48

3.2.2 RADIAL FORCE……………………………………………………………………………………………………………………….44

3.3 PUMP PIPING ............................................................................................................................... 52

3.4 VIBRATION MONITORING CONCEPT ............................................................................................. 55

3.5 VIBRATION MEASUREMENT AND PARAMETERS ........................................................................... 55

3.5.1 VIBRATION AMPLITUDE......................................................................................................... 55

3.5.2 VIBRATION FREQUENCY ........................................................................................................ 56

3.5.3 VIBRATION PHASE ................................................................................................................. 57

CHAPTER FOUR ..................................................................................................................................... 58

PRACTICAL DATA FROM SITE ................................................................................................................. 58

4.1 BEARINGS .................................................................................................................................... 58

4.2 LUBRICATION ............................................................................................................................... 61

4.3 SHAFT ALIGNMENT ...................................................................................................................... 61

4.4 VIBRATIONS ................................................................................................................................. 63

4.5 MATERIALS USED ......................................................................................................................... 67

4.6 TEMPERATURE MEASUREMENTS ................................................................................................. 67

4.7 PIPE NETWORK ............................................................................................................................ 68

CHAPTER FIVE ................................................................................................................................... 69

RESULTS AND ANALYSIS ................................................................................................................ 69

5.1. BEARINGS ................................................................................................................................ 69

5.2 LUBRICATION .......................................................................................................................... 71

5.3 VIBRATIONS ............................................................................................................................. 71

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5.4 TEMPERATURE ........................................................................................................................ 73

5.5 SHAFT ALIGNMENT ................................................................................................................ 73

5.6 PIPE NETWORK ........................................................................................................................ 74

CHAPTER SIX ..................................................................................................................................... 89

CONCLUSION AND RECOMMENDATIONS .................................................................................... 89

6.1 ROOT CAUSE OF FAILURE ..................................................................................................... 89

6.2 RECOMMENDATION ............................................................................................................... 90

6.3 CHALLENGES ........................................................................................................................... 92

6.4 CONCLUSION ........................................................................................................................... 92

6.5 APPENDIX ................................................................................................................................. 93

6.6 REFERENCES ............................................................................................................................ 94

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LIST OF FIGURES

Fig 1 Gantt chart………………………………………………………………………...10

Fig 2 Low Lift pumps Failure rate……………………………………………….……...11

Fig 3 Position of bearing on the Shaft…………………………………………………...13

Fig 4 Cross section of a centrifugal pump……………………………………………….17

Fig 5 Components of a centrifugal pump………………………………………………..18

Fig 6 Shaft ……………………………………………………………………………….19

Fig 7 Closed Impeller…………………………………………………………………….22

Fig 8 Open Impeller……………………………………………………………………...22

Fig 9 Special Open Impeller…………………………………………………………..…22

Fig 10 Non Clogging Impeller…………………………………………………………...22

Fig 11 Wear Resistant Closed Impeller………………………………………………….23

Fig 12 Wear Resistant Open Impeller………………………………………………...…23

Fig 13 Typical centrifugal pump casing………………………………………………....23

Fig 14 Volute casing……………………………………………………………………..24

Fig 15 Diffuser casing…………………………………………………………………...25

Fig 16 Classifications of Bearings…………………………………………………….. 28

Fig 17 liquid flow direction……………………………………………………………...33

Fig 18 Pump performance curves………………………………………………………..34

Fig 19 Efficiency Curve……………………………………………………………….....36

Fig 20 System Curve……………………………………………………………………..37

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Fig 21 Combination of Pump and system curves………………………………………..39

Fig 22 Pump operating point……………………………………………………………..41

Fig 23 Forces on end suction centrifugal pump………………………………………….42

Fig 24 Axial thrust on rotating impeller…………………………………………………44

Fig 25 Radial force in centrifugal pumps……………………………….……………….45

Fig 26 Radial forces in a double and single volutes design….………………………….46

Fig 27 Misalignment in pipe system…………………………………………….……….47

Fig 28 Design of discharge section………………………………………………………47

Fig 29 design of the suction side…………………………………………………………48

Fig 30 suction length in relation to diameter…………………………………………….48

Fig 31 Vibration signature……………………………………………………………….50

Fig 32 Bearings used for low lift pumps………………………………………………...52

Fig 33 Pump with external water cooling system……………………………………….54

Fig 34 low lift pump……………………………………………………………………..54

Fig 35 Laser Alignment System…………………………………………………………56

Fig 36 Vibration spectra and trends……………………………………………………..59

Fig 37 Pipe network…………………………………………………………………....62a

Fig 38 (a) welded bearing components, (b) Decolorized bearing due to heat…………..64

Fig 39 Shaft with welded inner race……………………………………………………..65

Fig 40 Brittle fractured bearings…………………………………………………………67

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Fig 41 System curve……………………………………………………………………..80

Fig 42 pump curve…………………………………………………………………….…81

Fig 43 combination of system and pump curve………………………………………….82

Fig 44; Impeller with two vanes curved from the eye…………………………………...83

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LIST OF TABLES

Table 1, Bearing operating characteristics……………………………………………...11

Table 2; Categories of Positive Displacement Pumps…………………………………..13

Table 3, classification of centrifugal pumps…………………………………………….14

Table 4; Bearing parameters…………………………………………………………….51

Table 5 Technical specifications and characteristics……………………………………53

Table 6 Alignment results……………………………………………………………….55

Table 7; Vibration readings……………………………………………………………..56

Table 8; diagnosis report on vibrations………………………………………………....58

Table 9; pump and pipe material………………………………………………………..59

Table 10 Temperature reading results…………………………………………………..60

Table 11 Fittings on the suction side……………………………………………………60

Table 12 Fittings on the discharge side…………………………………………………60

Table 13 Head and flow rate values of system curve…………………………………...79

Table 14 Head and flow rate values of pump curve…………………………………….80

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CHAPTER ONE

1 INTRODUCTION A pump is a device for lifting, transferring, or moving fluids by suction from one position

to another. Pumps have a variety of use both at industrial level and domestic level. They

help to move a fluid from one position to another. This may be with respect to elevation or

simply horizontal distance. Their application include; water pumping systems, chemical

process plants, food, beverage and pharmaceutical processing, pulp and paper processing,

sewage/waste water pumping etc.

During their life span pumps encounters failures which may be attributed to various

reasons, which may include; Poor operating conditions, lack of proper maintenance, wrong

application of pump etc. These failures usually manifest in specific forms. Very few

industrial pumps come out of service and go into the maintenance shop because the volute

casing or impeller split down the middle or because the shaft fractured into a number of

pieces. The majority of the pumps are taken out of service because the bearings or the

mechanical seals failed. (1)

This research is dedicated to look into the causes of failure of low lift centrifugal pumps at

NWSC Bulangililo Works. The pumps were installed back in 2002 but have been

experiencing failures since then. The failures are usually in form of broken bearings, (races

or cage), sometimes shaft damage.

The centrifugal pump was developed in Europe in the late 1600‟s and was seen in the last

seventy-five years. Prior to that time, the vast majority of pumping applications involved

positive displacement. The increased popularity of centrifugal pumps is due largely to the

comparatively recent development of high speed electric motors, steam turbines and

internal combustion engines. Centrifugal pumps are relatively high speed machines used in

many applications. Since 1940s, compact and efficient pumps have been able to be made

due to the development of high speed drivers.

Centrifugal pumps are commonly used to move liquids through a piping system; it consists

basically of a rotating device, called an impeller, inside a casing. The fluid enters the pump

impeller along or near to the rotating axis and is accelerated by the impeller, flowing

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radially outward into a diffuser or volute chamber(casing), from where it exits into the

down or upstream piping system. Centrifugal pumps are used for large discharge through

small heads.

1.2 PROBLEM STATEMENT

Frequent failures of Low Lift Pumps at Nkana Water and Sewerage Company- Bulangililo

Works.

1.3 AIM

The aim of this project is to undertake the root cause failure analysis of the pumps.

1.4 OBJECTIVES

The main objective of this project is to conduct a technical analysis of the failures.

1.5 PROJECT METHODOLOGY

The project methodology is as follows;

Site visitation

In our effort to find concrete and reliable information about the pumps, we intend to visit

the Nkana water Bulangililo plant and any other place which may have relevant

information i.e. Nkana East water pump, Mulonga water stations etc.

Literature Review

Literature on pumps is to be reviewed and other specialized companies in the field may be

consulted e.g. SKF, Alfred H Knight etc.

Data collection

Data will be collected by both site visitation and by talking to the operators, fitters,

engineers and any other relevant source of information.

The data to be collected include,

◦ Bearing types used

◦ Materials used

◦ Frequency of bearing failures

◦ Maintenance schedules.

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Data analysis

Conclusion and writing of report.

1.6: GANTT CHART

The following Gantt chart illustrates the pattern of work flow.

Fig1:Gantt chart

1.7 BACKGROUND

Nkana water and Sewerage Company Bulangililo Water Works has for the past nine (9)

years been experiencing frequent breakdowns of their low lift pumps. The pumps were

installed in 2002 in an attempt to rehabilitate the water works however, the pumps have

been failing and this has led to maintenance personnel changing the bearings of the pumps

nearly every month. Whenever there is a failure, it is usually in the form of damaged

bearing. Therefore, after every failure the bearings are replaced with new ones. This has

been the case since the installation of the new pumps. The failure, in addition to the low

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performance of the pumps has led to low efficiency of operations. The water works i.e.

Bulangililo, has four low lift pumps out of which only three are supposed to be running at a

time. This however has not been the case due to failures and general low output which has

led to all the four been usually operated in hope of getting a desirable water inflow.

The following graph in figure 2 gives a summary of failure rate as reflected in the log

books for operation records.

Fig 2:Low Lift pumps Failure rate

The colors in the graph represent a particular pump; the graph shows which pump had a

problem and how many times during that month. For instance, we can tell that in the month

of January 2011, pump 2 failed three times, pump 3 failed twice and pump 4 failed once.

This has been the trend for the pumps since inception.

1.8 PUMP DETAILS

The following are the specification of the low lift pumps in discussion;

MAKE: MATHER AND PLATT- CENTRIFUGAL PUMPS

MODEL: 14/16 BKN

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NAME PLATE NO: 99/1492

RATED DISCHARGE: 1350 M³/HR

RATED HEAD: 21.57 M

RATED SPEED: 985 RPM

PUMP EFFICIENCY: 84%

POWER ABSORBED: 94.40 KW

POWER PROVIDED: 132 KW

BEARINGS TYPE:

◦ NON DRIVE END: NU 218 EC

◦ DRIVE END: 22315

Table 1: Bearing operating characteristics [SKF General Catalogue]

BEARING TYPE

NU 218 EC

22315

NAME OF BEARING

Cylindrical Roller Bearing

Spherical Roller Bearing

LOAD RATING

Dynamic, C = 208 KN

Dynamic, C = 440 KN

Static, Co =220 KN

Static, Co = 475 KN

SPEED RATING

Reference Speed = 4500 rpm

Reference Speed = 3200 rpm

Limiting Speed = 5000 rpm

Limiting Speed = 4300 rpm

FATIGUE LOAD LIMIT

Pu = 27 KN Pu = 48 KN

MASS

2.30 kg

5.55 kg

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Figure 3 shows the location of the bearings and their positioning with reference to the impeller.

The arrangement is that of a cantilever setup and the impeller over hangs on the far end of the

arrangement. The shaft is 0.875m away from the coupling with the bearing being at a distance of

0.28m from the impeller and the second being at a distance of 0.59m from the impeller.

Fig 3: Position of bearing on the Shaft

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CHAPTER TWO

2. PUMP DESCRIPTION A pump is a mechanical device which uses suction to raise or move liquids, compress

gasses, or force air into inflatable objects such as tyres. A pump displaces a volume by

mechanical action; it creates a vacuum in an enclosed space by the removal of a gas.

Pumps are generally classified as follows;

Positive Displacement.

Non-positive Displacement.

2.1 POSITVE DISPLACEMENT

These are inherently pumps of low capacity and high discharge velocity. Positive

displacement pumps are constant flow machines; they produce a constant flow at a given

speed no matter what the discharge pressure. The pump traps a fixed amount of a fluid and

forces it into the discharge pipe.

Although used to pump a great variety of liquids, positive displacement pumps are also

capable of pumping slurries in small volumes or in consistencies that cannot be handled

by centrifugal pumps. Positive displacement pumps are subdivided into three categories:

reciprocating pumps, rotary pumps, and pneumatic (ejector) pumps, as shown in table 2

below.

Table 2: Categories of Positive Displacement Pumps

CATEGORY EXAMPLE

Reciprocating Pumps Plunger

Rotary Pumps Lobe, Progressive cavity, Screw

Pneumatic Pumps Ejector

Some positive displacement pumps work using an expanding cavity on the suction side and

a decreasing cavity on the discharge side. Liquid flows into the pump as the cavity on the

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suction side expands and the liquid flows out of the discharge as the cavity collapses. The

volume is constant given each cycle of operation.

The pumps in focus under this study are centrifugal which fall under non-positive

displacement pumps. We shall therefore base our attention on non-positive displacement

pumps.

2.2 NON POSITIVE DISPLACEMENT PUMPS

Non Positive Displacement Pumps are pumps that raise the energy of the liquid by first

imparting velocity energy to it and then converting this to pressure energy. These are also

called centrifugal pumps and are classified on the basis of characteristic features as shown

in table 3.

Table 3: classification of centrifugal pumps

CLASSIFICATION

METHOD

TYPES DESCRIPTION

1.Type Of Casing Volute pumps

Turbine pump/

diffusion pump

Area of flow in the casing

increases from impeller to

delivery pipe.

Has a series of guide

blades around the

impeller.

2.Working Head Low lift centrifugal

pumps

Medium lift centrifugal

pumps

High lift centrifugal

pumps

Works against heads of up

to 15m

Works with heads as high

as 40m

Works with heads above

40m

3.Liquid Handled Closed impeller pumps

Semi-open impeller

pumps(non-clog pump)

Open impeller pump

Used for pure and debris

free liquids

Used where liquids

contain debris

Used to pump liquid with

suspended solid matter.eg

sewage.

4.Number Of

Impellers Per Shaft

Single stage centrifugal

pump

Has one impeller, usually

low lift pump.

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Multi-stage centrifugal

pumps

Has 2 or more impellers,

used for high working

heads.

5.Number Of

Entrances To The

Impeller

Single suction/entry pump

Double entry/suction

pump

Water is admitted on one

side of impeller.

Water is admitted from

both sides of impeller,

axial thrust is neutralized.

6. Direction Of Flow

Through Impeller.

Radial flow pump

Axial flow pump

Mixed flow pump

Usually used in all

centrifugal pumps

Delivers huge quantities

of water at low heads,

suitable for irrigation

purposes

Used for irrigation

purposes

2.2.1 CENTRIFUGAL PUMPS

The centrifugal pumps is by far the most prolific member of the pump family, they are

widely used because of their design simplicity, high efficiency, wide range of capacity,

head, smooth flow rate and ease of operation and maintenance.

Centrifugal pumps include radial, axial and mixed flow units. A radial flow pump is

commonly referred to as a straight centrifugal pump, the common type of which is the

volute pump. In this type of pump, fluid enters the pump through the eye of the impeller,

which rotates at high speed. The fluid is accelerated radially outward from the casing. If

properly primed a partial vacuum is created that continuously draws more fluid in the

pump. In the axial flow centrifugal pump the rotor is a propeller and fluid flows parallel to

the axis of the shaft. In mixed flow, the direction of the liquid from the impeller acts as an

in between that of the radial and axial flow pumps. Figure 4 shows a centrifugal pump and

the flow of fluid through it.

They have real impellers with a series of blade attached to a rotating disc. The blades

normally slope backwards, away from the direction of rotation. As the impeller rotates the

curvature of the blade pushes the liquid out into the casing that surrounds it.

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Fig 4: Cross section of a centrifugal pump [14]

This collects the displaced fluids and directs it into the piping. The centrifugal force arises

because the radius at which an individual particle leaves the impeller is larger than that at

which it enters. As liquid is displaced out into the casing, more liquid is pushed in to

replace it. This process takes place continuously to create a pumping action.

2.2.2 COMPONENTS OF A CENTRIFUGAL PUMP.

A centrifugal pump has two main components:

Rotating masses comprised of an impeller and a shaft

Stationary masses comprised of a casing, casing cover, and bearings.

The general components, both stationary and rotary, are depicted in figure 5. Each of the

component plays a distinguished role that cannot be over looked.

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Fig 5: Components of a centrifugal pump [14]

Key; 1.Suction Nozzle 2.Impeller 3.Volute 4.Discharge Nozzle 5.Casing 6.Seal 7.Shaft 8.

Single roller bearing 9.Spherical roller bearing 10.Oil rings

2.2.2.1 SHAFT

The basic purpose of a centrifugal pump shaft is to transmit the torque encountered when

starting and during operation while supporting the impeller and other rotating parts. It must

do this job with a deflection less than the minimum clearance between the rotating and

stationary parts. The loads involved are as follows:

torques

weight of the parts

axial and radial hydraulic forces

In transmitting power from the driver to the impeller, the pump shaft must support all radial

and axial impeller forces. A shaft is a rotating member, usually of circular cross section,

used to transmit power or motion to the impeller. It provides the axis of rotation, or

oscillation, of elements such as gears, pulleys, flywheels, cranks and sprockets etc. The

geometry of a shaft is that of a stepped cylinder. The use of shaft shoulders is an excellent

25

means of axially locating the shaft elements and to carry any thrust loads. An example of a

shaft is given in figure 6.

Fig 6: Shaft

A pump shaft is subject to torsion, bending, and tension. Shafts are usually designed to

withstand the stress set up when a pump is started quickly. Good shaft design is needed to

ensure that the endurance limit of the shaft material is not exceeded under any anticipated

operating conditions. Furthermore, the deflection of the shaft must be controlled to avoid

wear ring rubbing and damage to the packing or the mechanical seals. A deflection of

0.05mm (0.002in.or 2 mils) at the shaft seal is often given as a recommended limit.

Deflection at radial wear rings should be compatible with wear ring nominal clearance.

Deflection is not affected by strength, but by stiffness, represented by the modulus of

elasticity, which is essentially constant for all steels. Thus, rigidity cannot be controlled by

material decisions, but only by geometric decisions.

Any object made of an elastic material has a natural period of vibration. When a pump

impeller and shaft rotate at any speed corresponding to the natural frequency, minor

imbalances will be magnified. The speeds at which this magnification takes place are called

critical speeds. The lowest critical speed is called the first critical speed. The next higher is

called the second critical speed, etc.

26

In centrifugal pump nomenclature:

A rigid shaft means one with an operating speed lower than its first critical speed;

A flexible shaft is one with an operating speed higher than its first critical speed.

The shaft critical speed can be reached and passed without danger because frictional forces

(surrounding liquid, stuffing box packing, and various internal leakage joints) tend to

restrain the deflection for a short duration.

2.2.2.2 WEARING RINGS

Wearing rings (for casing or impeller) provide an easily and economically renewable

leakage joint. There are various types of wearing ring designs, and the selection of the

most desirable type depends on the following:

liquid being handled

pressure differential across the leakage joint

rubbing speed

pump design (i.e., sewage vs. clean liquid)

In leakage joints involving a flat-type wearing ring, the leakage flow is a function of the

following:

area

length of the joint

pressure differential across the joint

If the path is broken by relief chambers, the velocity energy in the leakage jet is

dissipated in each relief chamber, thereby increasing the resistance. As a result, with

several relief chambers and several leakage joints for the same actual flow through the

joint, is less resulting in higher pump performance and operating efficiency.

2.2.2.3 IMPELLER

The impeller is the main rotating part that provides the centrifugal acceleration to the fluid.

They are often classified in many ways.

Based on major direction of flow in reference to the axis of rotation

Radial flow

Axial flow

Mixed flow

27

Based on suction type

Single-suction: Liquid inlet on one side.

Double-suction: Liquid inlet to the impeller symmetrically

Closed impellers require wear rings and these wear rings present another maintenance

problem. Open and semi-open impellers are less likely to clog, but need manual adjustment

to the volute or back-plate to get the proper impeller setting and prevent internal re-

circulation. Vortex pump impellers are great for solids and "stringy" materials but they are

up to 50% less efficient than conventional designs. The number of impellers determines the

number of stages of the pump. A single stage pump has one impeller only and is best for

low head service. A two-stage pump has two impellers in series for medium head service.

Open Impeller

An open impeller consists of vanes attached to a central hub without any form of sidewall or

shroud.

Disadvantage of an open impeller

Structural weakness – if the vanes are long, they must be strengthened by ribs or a

partial shroud. Generally, open impellers are used in small inexpensive pumps or

pumps that handle abrasive liquids.

Advantage of an open impeller

It is capable of handling suspended matter with a minimum of clogging.

The open impeller rotates between two side plates, between the casing walls of the volute.

The clearance between the impeller vanes and sidewalls allows a certain amount of water

recirculation, which increases as wear increases. To restore the original efficiency, both the

impeller and the side plates must be replaced. This is a much greater expense than would be

encountered by an enclosed impeller where simple rings form the leakage point.

Semi-Open Impeller

The semi-open impeller incorporates a shroud or an impeller back wall. This shroud may or

may not have “pump-out” vanes, which are located at the back of the impeller shroud.

Function of the “pump-out” vanes

To reduce the pressure at the back hub of the impeller;

28

To prevent foreign matter from lodging in the back of the impeller and interfering

with the proper operation of the pump and the stuffing box.

Enclosed Impeller

The enclosed impeller is used universally in centrifugal pumps that handle clear liquids. It

incorporates shrouds or enclosing sidewalls that totally enclose the impeller “waterways”

from the suction eye to the impeller periphery. This design prevents the liquid recirculation

that occurs between an open or semi-open impeller and its side plates. A running joint must

also be provided between the impeller and the casing to separate the discharge and suction

chambers of the pump. The running joint is normally formed by a relatively short cylindrical

surface on the impeller shroud that rotates within a slightly larger stationary cylindrical

surface. If one or both surfaces are made removable, the “leakage joint” can be repaired

when wear causes excessive leakage. Some examples of impellers are given below;

Fig 7: Closed Impeller Fig 8: Open Impeller

The Closed impeller of figure 7 is used for pumping clean liquids.

The Open Impeller in figure 8 is designed for liquids containing solid particles, abrasive

liquid or stock up to 8%consistency

Fig 9: Special Open Impeller Fig 10: Non-Clogging Impeller

29

The Special Open Impeller in figure 9 is suitable for liquids containing bigger solids

particles and long fibers, abrasive liquids or stock up to 8% consistency.

The Non Clogging Impeller in figure 10 is used for sludge or slurries containing big solid

particles.

Fig 11: Wear Resistant Closed Impeller Fig 12: Wear Resistant open impeller

The Wear resistant closed impeller in figure 11 is used for pumping either erosive and

corrosive liquids or slurries containing solid particles.

Wear Resistant Open Impeller in figure 12 is suitable for liquids with bigger solid particles

and long fibers, abrasive liquids or stock up to 8% consistency.

2.2.2.4 CASING

The pump casing is of single/double volute design, two halves of which are bolted together

along the pump axis. Gaskets are provided between the split flanges of top and bottom

casing. The suction and delivery branches of the pump are case integral with bottom half

casing, which also incorporates the mounting feet. Tapings are drilled in suction and

delivery branches for connecting the pressure gauges and providing casing drain. Tapings

are provided in top half casing for taking water seal connection on both the stuffing boxes.

The typical centrifugal pump casing is shown in figure 13 below.

Fig 13: Typical centrifugal pump casing

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The pump casing is an air tight chamber surrounding the pump impeller encloses and

contains the discharge and, sometimes, the suction nozzles. The casing collects the high-

velocity flow from the impeller and converts some of the velocity energy into pressure. The

casing may be one piece or of axial or radial–split design. It has provision to fix stuffing

box and house packing materials which prevent external leakage. The essential functions of

the casing are to guide water to and from the impeller.

The following three types of casings are commonly employed:

Volute casing

In this type of casing the area of flow gradually increases from the impeller outlet to the

delivery pipe so as to reduce the velocity of flow. This can be clearly seen in figure 14.

Thus, the increase in pressure occurs in volute casings.

Fig 14: Volute casing

Vortex Casing

If a circular chamber is provided between the impeller and the volute chamber, the casing

is known as vortex casing. The circular chamber is known as vortex or whirlpool chamber

and such a pump is called a volute pump with vortex chamber. The vortex chamber

converts some of the kinetic energy into pressure energy. The volute chamber further

increases the pressure energy. Thus, the efficiency of a volute pump fitted with a vortex

chamber is more than that of a simple volute pump.

31

Diffuser Casing

In this type of casing, the impeller is surrounded by a series of guide vanes, mounted

on a ring which is known as the diffuser. The liquid leaving the impeller passes

through the passage having a gradually increasing area between the vanes. The

velocity of flow decreases and the kinetic energy is converted into pressure energy.

Machines with diffuser vanes have maximum efficiency, but are less satisfactory

when a wide range of operating conditions is required. Fig 15 shows a diffuser type of

casing.

Fig 15: Diffuser casing [14]

2.2.2.5 BEARINGS

A pump bearing system consists of bearings to carry radial loads and at least one bearing

to carry axial or thrust loads. For horizontally mounted pumps, there are two basic pump

rotor arrangements: the pumping element overhangs the bearing system or the

pumping element is between bearings. The function of bearings in centrifugal

pumps is to keep the shaft or rotor in correct alignment with the stationary parts under

the action of radial and transverse loads. Bearings that give radial positioning to the rotor

are known as radial or line bearings, and those that locate the rotor axially are called thrust

bearings. In most applications, the thrust bearings actually carry both axial and radial

loads.

The bearings in every process pump accomplish three important functions;

32

They locate the rotating element in its correct position relative to the stationary

parts of the pump.

They allow the shaft to rotate with the least amount of friction to maximize the

pump operating efficiency.

They absorb all the radial and axial loads which are transmitted through the shaft

during the different operating modes.

The thrust bearing of the cantilever-type pump carries not only the entire thrust load but

also its share of the radial load. The thrust bearing must be capable of supporting the axial

loads in both directions because thrust reversal can be expected with most pumps,

especially during the starting process. In same designs, a separate (usually smaller) thrust

bearing is installed for the thrust reversal. Typically, the thrust bearing is pressed on the

shaft and locked in the bearing housing; thus the thrust bearing determines the axial

location of the shaft and the impeller. The radial bearing is also pressed on the shaft but is

free to slide in the frame or in the bearing housing. This protects the bearings from

excessive axial loads generated by unequal thermal expansion of the shaft and frame.

Bearing Arrangement

The Face to Face arrangement and the Back to Back arrangement can both accommodate

axial loads in either direction, but only by one bearing at a time. In view of this similarity,

it is frequently assumed that these arrangements are interchangeable. Not in end suction

process pumps.

The Back to Back arrangement is the only one that can accommodate a tilting moment.

Consequently this one must be used when the thrust bearing is located close to the source

of a tilting moment. As that can occur as a result of shaft misalignment, the Back to Back

arrangement must always be used in an end suction pump where it is close to the shaft

coupling. The thrust bearing in a “double-ender” type of pump is located at the opposite

end of the shaft from the coupling. But if we continue to use the Back to Back

arrangement in this pump style, it will ensure that the maintenance team doesn‟t have to

remember which bearing style goes in which pump. Therefore, every time a double

angular contact bearing is used, it should be mounted in the Back to Back arrangement.

33

To ensure that these bearings provide long-lasting, trouble-free service, it must be

recognized that they are only a part of the total bearing arrangement. Other important

aspects include the support and the protection of these bearings. Such support includes a

strong shaft and housing to minimize the effect of any externally induced stresses or

vibration. A strong shaft that will minimize the effect of vibration is essential to the

support of the pump bearings. Therefore the slenderness ratio of the shaft must be kept at

as low a level as is possible.

Bearing Frame and Housing.

A bearing frame is a member of an end suction pump to which is assembled the liquid end

and rotating element. A bearing housing is a pump component into which the bearings are

mounted. In the case of the end suction pump, the frame also serves as the bearing housing.

Frames or bearing housings serve the purpose of providing bore alignment for the bearings,

an oil or grease reservoir of adequate capacity, heat dissipation by convection or other

means and protective shaft seals to keep out dirt and moisture and keep oil in. In addition,

frames also serve as a mounting means for the wet end and provide a pedestal foot or feet

to the base. Obviously, housings must be sturdy enough to withstand the real life distortion

forces they will be subjected to without distorting to a point that their bearing alignment is

threatened.

Bearing Classification

Bearings for pumps are classified in many ways. Figure 16 shows the classification of

bearings. Line verses thrust describe the bearing‟s ability to handle radial or axial thrust

loads and how the bearing positions the rotor. Cylindrical roller bearings have no inherent

axial thrust handling capability. The outboard bearing is made up of two angular contact

ball bearings, back to back. These bearings can take both radial and thrust loads. The

inboard ball bearing takes up thrust loads imposed in the direction of the driver and the

outboard ball bearing takes up any thrust the inlet of the pump.

The bearings most used on pumps are the single and double deep groove ball, the double

row self-aligning and the single and double row angular contact.

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Single Row Deep Groove Ball – Conrad type: this bearing can carry significant radial

loads and substantial thrust loads in either direction, even at high speeds.

Double Row Deep Groove Ball: similar to single row above but added row allows

substantial increase in radial load.

Self-aligning Ball Bearings: have two rows of balls and a common spherical raceway

which provides the self – aligning capability.

Angular contact Ball Bearings: can carry appreciable thrust loads in one direction either

alone or with combined radial loading.

Fig: 16 Classifications of Bearings

35

Corresponding Cylindrical Roller Bearings can offer higher radial load carrying capacity

than their ball bearing counterparts. Rubber lined, cut less bearings are used in water

service in applications involving sand and grit. They are designed to provide a

hydrodynamic wedge under the shaft and lift it clear of the bearing. They can operate dry

for short start up periods.

Bearing failures are one of the three highest failure areas on centrifugal pumps. These

failures are caused by excessive operation at low flows, excessive axial thrust loads due to

obstructions, improper upstream piping or wear, improper or inadequate lubrication,

misalignment and piping strains, dirt and moisture.

2.2.2.6 LUBRICATION

Grease is the most common lubricant for pump bearings. It requires relatively little

attention and service and usually gives long bearing life. Oil-bath lubricant is used much

less frequently because returning the oil in the bearing housing possesses some technical

difficulties, especially on vertical pumps. Force-fed oil lubrication is used very

infrequently. Although it ensures the lowest bearing operating temperatures, provides the

best protection from bearing contamination, and gives the longest bearing life, it requires

expensive auxiliary equipment, which also needs service and attention.

The lubricant separates the rolling elements and raceway contact surfaces, and minimizes

the effect of friction. The selection of the lubricant is a consideration of its viscosity, and

depends on the operating temperature, the bearing size and its rotational speed.

While the bearing manufacturer can identify the minimum viscosity required for these

conditions, the selected lubricant should provide a higher viscosity than the minimum

identified.

Grease Lubrication

Lubricating grease is essentially a soap thickening agent in mineral or synthetic oil. When

selecting the right grease, the base oil should be able to satisfy the bearing‟s lubrication

requirements. When using grease, two rules must be followed;

◦ Do not apply too much grease to a bearing as it will cause the grease to overheat

and reduce the lubricating effectiveness.

36

◦ Do not mix different types of grease as many of their contents and preservatives

are incompatible.

Oil Lubrication

Mineral oils are still the most common lubricating oil in general service pumps, while

synthetic oils tend to be more resistant to higher temperatures and seem to require less

frequent change in these applications. The oil level in the bearing housing should be

maintained at the centerline of the lowest ball in the race, and the oil must be able to enter

the bearing from both sides. An increasingly popular method of pump bearing

Lubrication is the Oil Mist system where the mist is a collection of atomized oil droplets

that are sprayed in the bearings by compressed air. The Purge Oil Mist system

incorporates a static oil bath for the bearings, while the Pure Oil Mist system comprises

the only form of lubrication used. The latter is gradually taking over almost exclusively as

it has been found that the Purge System is less effective. The Pure Oil Mist System is used

to fill the bearing housing as much as possible with oil in order to minimize the entry of

contaminants.

2.2.2.7 SUCTION NOZZLE

This directs the flow of liquid to the eye of the impeller. A hand hole cover is provided for

cleaning and inspection purposes. A gasket is provided between the hand hole cover and

the suction pipe and also between the suction pipe and casing in order to avoid leakage. To

prevent the entrance of solid particles, debris etc. into the pump is provided with a strainer

at its lower end. The lower end of the pipe is also fitted a non-return foot valve which does

not permit the liquid to drain out of the suction pipe when the pump is not working.

This also helps in priming. The suction cover encloses the suction opening of the casing

and also contains the suction nozzle of the pump. On some designs, the suction cover is an

integral part of the pump casing.

2.2.2.8 DISCHARGE NOZZLE

This is a pipe connected at its lower end to the outlet of the pump and it delivers the liquid

to the required height. A regulating valve is provided on the delivery pipe to regulate the

supply of water.

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2.2.2.9 STUFFING BOX

The primary function of a stuffing box is to prevent leakage at the point where the shaft

passes out through the pump casing. For general service pumps, a stuffing box consists of a

cylindrical recess that accommodates a number of rings of packing seal cage and gland

around the shaft or shaft sleeve. If sealing liquid to the box is desired, a lantern ring or seal

cage is used, which separates the rings of packing into approximately equal sections. The

packing is compressed to give the desired fit on the shaft or sleeve by a gland that can be

adjusted in an axial direction. A small leakage from the stuffing box is required to provide

lubrication and cooling.

2.2.2.10 STUFFING BOX COVER

The stuffing box cover, also called the back head or adapter, encloses the inboard opening

of the casing and also contains the stuffing box.

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CHAPTER THREE

3. CHARACTERISTICS AND OPERATIONS OF CENTRIFUGAL PUMPS. A centrifugal pump is made up of a set of rotating vanes that are enclosed within a housing.

These vanes are utilized to impart energy to a liquid through centrifugal force thereby

building up a velocity head. At the periphery of the pump impeller, the liquid is directed

into a volute which commonly has an increasing cross sectional area along its length so that

as the liquid travels along the chamber and its velocity is reduced. Since the energy level of

the liquid cannot be dissipated at this point, the conversion of energy law (Bernoulli‟s

theorem) requires the liquid loses velocity energy as it moves along the chamber, it must

increase the energy related to pressure. Hence the increase in pressure.

A centrifugal pump has two main parts, a rotating element i.e. an impeller and shaft, and a

stationary element which is made up of a casing, stuffing box and bearings. The pump

transfers the energy provided by a prime mover such as an electric motor, steam turbine, or

gasoline engine to energy within the energy being pumped. The energy within the pump is

present as a velocity energy, pressure energy, or a combination of both. The liquid being

pumped surrounds the impeller and as the impeller rotates, the rotating motion of the

impeller imparts a rotating motion to the liquid. There are two components to the motion

imparted to the liquid by the impeller; one motion is in the radial direction outward the

centre of the impeller. This motion is caused by the centrifugal force due to the rotation of

the liquid, which acts in a direction outward from the centre of the rotating impeller.

Secondly, as the liquid leaves the impeller, it tends to move in a direction tangential to the

outside diameter of the impeller. The actual liquid direction is as a result of the two flow

direction as shown in figure 17.

The amount of the energy being added to the liquid by the rotating impeller is related to the

velocity with which the liquid moves. The energy expressed as pressure energy will be

proportional to the square of the resultant exit velocity.

39

Fig 17: Liquid flow direction.[14]

Were, H=energy, V= velocity, g=acceleration due to gravity.

Form this equation it is evident that any increase in the impeller tip velocity will increase

the energy imparted to the liquid. Additionally, any change in the vane tip velocity will

result in a change in the energy imparted to the liquid that is proportional to the square of

the change in tip velocity

3.1 PUMP PERFORMANCE CURVE

The pump characteristic is normally described graphically by the manufacturer as a pump

performance curve. The pump curve describes the relation between flow rate and head for

the actual pump. The other important information for proper pump selection is also

included - efficiency curves, Net Positive Suction Head Required NPSHr curve, pump

curves for several impeller diameters and different speeds, and power consumption. These

are shown in figure 18.

40

For a pump to properly function, the NPSHA must be greater than the NPSH required by the

pump (obtained from the pump manufacturer). Increasing the impeller diameter or speed

increases the head and flow rate capacity - and the pump curve moves upwards.

Fig 18: Pump performance curves [14]

The head capacity can be increased by connecting two or more pumps in series, or the flow

rate capacity can be increased by connecting two or more pumps in parallel.

3.1.1 NET POSITIVE SUCTION HEAD

Net Positive Suction Head (NPSH) may be defined as the difference between the net inlet

head and the head corresponding to the vapor pressure of the liquid. NPSH may also be

defined as the net head (in meters of liquid) that is required to make the liquid flow through

the suction pipe from the sump to the impeller. It is related to how much suction lift a pump

can achieve by creating a partial vacuum, atmospheric pressure then pushes the liquid into

the pump. This term has significance only when cavitating liquids are handled.

41

NPSH is a dimensional parameter that can be used to check cavitation in pump, the

minimum NPSH depends upon the pump design, its speed and the discharge.

NPSH (a) is the Net Positive Suction Head Available; it is a function of the system in

which the pump operates. It is the excess pressure of the liquid in feet absolute over its

vapor pressure as it arrives at the pump suction, to be sure that the pump selected does not

cavitate.

NPSH(r) is the Net Positive Suction Head Required by the pump, which is read from the

pump performance curve. NPSH required is a function of the pump design and is

determined based on actual pump test by the vendor. As the liquid passes from the pump

suction to the eye of the impeller, the velocity increases and the pressure decrease. There

are also pressure losses due to shock and turbulence as the liquid strikes the impeller. The

centrifugal force of the impeller vanes further increases the velocity and decreases the

pressure of the liquid. The NPSH required is the positive head in feet absolute required at

the pump suction to overcome these pressure drops in the pump and maintain the majority

of the liquid above its vapor pressure. NPSH(r) may be thought of as friction loss caused by

the entry to the pump suction.

The NPSH is always positive since it is expressed in terms of absolute fluid column height.

The term "Net" refers to the actual pressure head at the pump suction flange and not the

static suction head.

3.1.2 EFFECIENCY CURVE

The efficiency of a pump depends on various factors such as the type, the size of the pump,

rotational speed, hydraulic layout, surface finish, fluid properties and specific speed. At

high specific speed efficiency starts to fall off due to an increasing proportion of hydraulic

losses. With reduced specific speed there is above all a considerable increase in the

proportional of disc friction losses, leakage and balancing flow losses. Figure 19 shows the

efficiency curve.

42

Fig 19: The Efficiency Curve.

The actual head capacity curve stipulates the output of the centrifugal pump. From this the

water-kW can be calculated by means of the following formula;

kW = (Q*H)/ 102

Q is in m3 /s

H is in m.

Due to additional mechanical losses, the input of power has to be higher. The efficiency of

the pump therefore is:

Efficiency = kilowatt output/kilowatt input = ( Q * H ) / power-kW

As the water kilowatt is zero at no flow, the efficiency curve starts at zero, rising steadily to

a maximum and then decreasing with increasing flow.

3.1.3 THE SYSTEM CURVE

The system resistance or system head curve (a graphical presentation of the Energy

Equation) is a curve showing the relation between the quantity of liquid flowing through

the pipeline and the friction head loss in the pipeline. Refer to figure 20. It is also defined

as the change in flow with respect to head of the system. It must be developed by the user

based upon the conditions of service. These include physical layout, process conditions,

and fluid characteristics. It represents the relationship between flow and hydraulic losses in

a system in a graphic form and, since friction losses vary as a square of the flow rate, the

system curve is parabolic in shape.

43

Hydraulic losses in piping systems are composed of pipe friction losses, valves, elbows

and other fittings, entrance and exit losses, and losses from changes in pipe size by

enlargement or reduction in diameter.

Fig 20: System Curve. [1]

The system head visualized in the System Curve is a function of the elevation - the static

head in the system, and the major and minor losses. The total head of a system of;

Static Head; this is the height difference between the suction and discharge fluid levels; it is

the difference in the liquid surface level at the suction source or vessel subtracted from the

liquid surface level where the pump deposits the liquid. If the discharge pipe emergences

above the liquid level, then the static head is referred to the center line of the outflow

section.

Pressure Head; this is change in pressure across the system. It is expressed in feet of head

and may or may not exist in all the system. For instance, if both the suction and discharge

vessels have the same pressure then there will be no pressure head.

Velocity Head; this is the energy lost into the system due to the velocity of the liquid

moving through the pipes.

Friction Head; this is the friction losses in the system expressed in feet of head. The friction

head is the measure of the friction between the pumped liquid and the internal walls of the

44

pipe, valves, connections and accessories in the suction and discharge piping. Because the

velocity head and the friction head are energies lost in the system, this energy would never

reach the final point where it is needed. Therefore, these heads must be calculated and

added to the pump at the moment of design and specifications. It is also necessary to know

these values especially when they are significant, at the moment of analyzing a problem in

the pump. The friction head and velocity head can be measured with pressure gauges in an

existing system.

The sum of these four heads is called system head;

HT= Hs + Hp + Hf + Hv

Where

HT is system head, Hs is the Static head, Hp is Pressure head, Hf is Friction head and Hv is

Velocity head.

However, we know that;

Hf = f Lv2 /2gd

Where

f is Darcy‟s frictional coefficient or factor

L is the length of the pipe

v is the average fluid velocity

d is the inside pipe diameter

g is the acceleration due to gravity.

Darcy‟s equation can be applied for calculating the head loss to friction, for both laminar as

well as turbulent flows. The only difference will be in the evaluation of the frictional

coefficient „f‟.

Frictional losses in Laminar flow has the following friction factor;

45

f = 64/Re

Where Re is Reynolds number.

Substituting this in Hf we have;

Hf = 64 L v2 /Re 2gd

This is called the Hagen- Poiseuille equation.

For the frictional losses in turbulent flow the friction factor is dependent on both the

Reynolds number and the relative roughness of the pipe. The relative roughness is given

by;

Relative roughness = Pipe inside surface roughness/Pipe inside diameter = € /D

Where D is diameter and € is absolute roughness.

3.1.4 SELECTION OF PUMP

A pump can be selected by combining the System Curve and the Pump Curve as illustrated

in figure 21.

Fig 21: Combination of Pump and system curves. [1]

46

The operating point is where the system curve and the actual pump curve intersect.

The system curve shows the complete picture of the dynamic system. This permits the

purchase, Installation and maintenance of the best pump for the system. The system curve

is most useful when mated with the pump family curve. This is why the family curves are

the most useful to the design engineer, the maintenance engineer, and purchasing

personnel. The pump always operates at the intersection of the system curve and the pump

curve.

The main criteria of the selection of the type of pump are the values of discharge (Q), Head

(H) and speed (N). From these values the specific speed of the pump is calculated and

subsequently the type of motor can be decided. When the specific speed is low and it is

possible to increase the pump speed, it is better to use multistage pumps, head and the

number of stages is decided on the basis of the head and the type of pump to be used.

The type of impeller is another aspect of pump selection.

Impeller shrouded type; used for pumping fresh clean water.

Impeller-unshrouded or propeller type; used for pumping solid-liquid mixture or near

plastic material.

Mixed flow impeller with diffuser vanes; used for deep well or submersible pumps.

Axial flow pumps are employed for very low heads of about 6m and for large discharges.

Radial flow pumps are used when the head is high.

3.1.5 PUMP OPERATING POINT

When a pump is installed in a system the effect can be illustrated graphically by

superimposing pump and system curves as shown in figure 22. The operating point will

always be where the two curves intersect.

If the actual system curve is different in reality to that calculated, the pump will operate at a

flow and head different to that expected. For a centrifugal pump, an increasing system

resistance will reduce the flow, eventually to zero, but the maximum head is limited as

shown in figure 22. Even so, this condition is only acceptable for a short period without

causing problems.

47

Fig 22: Pump operating point

An error in the system curve calculation is also likely to lead to a centrifugal pump

selection, which is less than optimal for the actual system head losses. Adding safety

margins to the calculated system curve to ensure that a sufficiently large pump is selected

will generally result in installing an oversized pump, which will operate at an excessive

flow rate or in a throttled condition, which increases energy usage and reduces pump life.

3.2 TYPICAL FORCES ON IMPELLERS

The principal forces on impellers are primarily exerted during operation when the impeller

or group of impellers is turning within the volute at the design speed. Although other

forces, such as pressure or head, can be imposed on the interior and exterior surfaces of

impellers, these forces are generally imposed equally to all of the surfaces of the impeller

during a static or non-operating state.

The primary forces exerted on impellers during operation are axial forces (the forces

normal or parallel to the impeller eye) and radial forces (those forces that are exerted

perpendicular to the impeller eye like suction). The application of these two common

48

forces for a typical end-suction centrifugal pump is shown in the figure 23.

Fig 23: Forces on end suction centrifugal pump. [1]

Although each of these applied forces is generally provided for by the pump manufacturer

in the original design of the pumping unit, it is important for system designers to have a

basic understanding of the forces.

3.2.1 AXIAL FORCE

The axial force, also known as thrust, on an impeller is the sum of the unbalanced forces

acting on the impeller in the axial direction, or parallel to the pump shaft. Although the

maximum axial force is typically applied only to one side of the impeller, it can reverse

direction or even alternate should a change in the pump‟s operating conditions occur.

Axial forces are the result of one or more of the following conditions:

pressure distribution at the mechanical seal or packing leakage zone

pressure distribution within the leakage zone of the hub

pressure difference at opposite ends of the shaft

momentum of the incoming or outgoing fluid flow

49

pressure distribution at the hub in the fluid passages (between vanes)

Pressure difference between both sides of the vanes (in the axial direction).

Depending on the origin of the source, each of these forces is further defined as

mechanical or fluid-induced. Mechanical forces are those resulting from a mechanical

influence, such as rotor imbalance or vibrations. Fluid-induced forces result due to the

direct influence from the working fluid.

Most axial forces imposed in a horizontal centrifugal pump are fluid-induced forces.

Axial forces also act in vertical turbine and submersible pumps. However, in these cases it

is usually referred to as down thrust or up thrust with the combined effect of hydraulic and

mechanical forces due to the weight of the rotating impeller stack and drive shaft.

As seen in Figure 22, the application of axial thrust on a rotating impeller is generally

unbalanced, with the resulting value dependent on the lower pressure region within the

suction eye versus the higher pressure resulting from the impeller discharge exerted on the

outer surfaces of the impeller. To prevent lateral movement of the rotating assembly, this

force differential is typically resisted in a single thrust bearing or multiple thrust bearings

if the load is too high for a single bearing to withstand. The thrust bearing is usually

located in either the pump frame or the driver e.g. Electric motor.

In addition to lowering the axial forces, a hydraulic balance line also offers inherent

advantages for installations with sandy conditions, as these contaminants are continuously

flushed out of the packing or mechanical seal zone to avoid abrasive wear on the running

surfaces of the seal/packing and the shaft sleeve. In many high head or capacity end-

suction centrifugal pumps, the differential of axial forces between the discharge and

suction components of the impeller can cause high loading and premature failure of the

thrust bearing. In these cases remediation methods, such as a hydraulic balance line as

shown in Figure 24, are often used to provide a path for circulating water from the higher

pressure region of the impeller discharge back to the suction port.

Another method commonly used to reduce axial thrust loads on high head centrifugal

pumps is through the drilling of small bypass holes through the impeller, providing a path

50

of recirculating fluid between the suction eye and the higher pressure region of pressure

behind the impeller

Fig 24: Axial thrust on rotating impeller. [1]

This method should be used with extreme care, as decreased efficiency and performance is

often a penalty paid for this fix.

A final method involves the installation of wear rings on the back side and front side of

the impeller. Should a system designer wish to provide an installation with reduced axial

thrust, use of a double-suction pump may also be considered. Although not an automatic

solution, a double-suction pump properly configured and installed provides an alternative

to the typical single-suction centrifugal pump for applications with projected high axial

force loads and may be a valid consideration for installations using existing drivers with

limited thrust bearing capacities.

51

3.2.2 RADIAL FORCE

In addition to the axial forces within an operating pump, most centrifugal pumps also

work with a component of radial force, or thrust, which is also shown in Figure 23 as a

force exerting perpendicular to the shaft and acting on the periphery of the impeller from

the discharge water impacting the volute. Radial forces that exist around the edge of a

single impeller are proportional to the pump‟s total head (in feet), the impeller diameter,

and the vane width of the impeller.

Radial forces are typically more of a concern for high head and high speed volute type

centrifugal pumps, especially those with large diameter impellers, and are not generally as

much of a concern with vertical turbine or submersible pumps due to the difference of

design and construction between the respective unit

Fig 25: Radial force in centrifugal pumps. [14]

As shown in Figure 25 an impeller within a volute type of centrifugal pump is surrounded

by a volute with unequal distances around the periphery between the edge of the impeller

and the volute casing. This differential of space results in differing values of radial forces,

while the diffuser type of pump is encircled with a bowl and diffuser providing a relatively

uniform spacing between the impeller discharge and the diffuser, resulting in a more

uniform value of the radial force surrounding the impeller.

52

Unbalanced radial force usually results in increased shaft deflection, which in turn can

lead to premature seal or packing failure, bearing and wear ring damage and failure, and

premature motor bearing failure for those units with direct driven electric motors.

In extreme cases, prolonged operation with extreme unbalanced radial forces can result in

shaft breakage.

Fig 26: Radial forces in a double and single volutes design. [15]

In response to the problems associated with radial forces on high head volute pumps,

many manufacturers have developed a modified volute pattern, called a double volute, as

shown in Figure 26. A double volute style of pumps includes an added internal discharge

vane to a standard volute pump, thus providing a reasonably uniform distance between the

impeller periphery and the volute around the impeller‟s entire circumference. This added

feature greatly lowers the unbalanced radial forces within the pump and extends the life of

all rotating components.

3.3 PUMP PIPING Flange bolts should not be used to unite misaligned piping to the pump as shown in

figure 27. This damages the flange faces and stresses the pump casing.

53

Fig 27: Misalignment in pipe system. [1]

The discharge pipe must not have any turn in the first section after the pump which

is 10 times diameter of pipe.ie. If pipe diameter is 0.6m, then the first 6m of the

discharge pipe must be straight. This is illustrated in figure 28.

Fig 28: Design of discharge section

The suction pipe must not have concentric reducer (shown in figure 29a) as they

create air traps which cause cavitation. It must have an eccentric pipe reducer

instead as illustrated in figure 29(b).

54

(a)

(b)

Fig 29: Design of the suction side. [1]

The length of the suction end must be 10 times the diameter of the pipe and must

not have any turns within this distance. This is diagrammatically represented in

figure 30.

Fig 30: suction length in relation to diameter. [1]

55

3.4 VIBRATION MONITORING CONCEPT

Mechanical vibration may be simply defined as the oscillatory motion of a machine or

machine component about a position of rest. This arises as a result of normal transmission

of cyclic forces within the structure. These forces may be due to inherent design factors, or

may be due to developing and deteriorating defects.

Developing and deteriorating defects give rise to changing vibration characteristics. Thus

all rotating machines do vibrate.

Nonetheless, depending on the characteristics of the vibration, the levels of vibration may

be acceptable or not. These characteristics when measured, recorded and trended can be

analyzed for instantaneous severity, and type of defect. Thus, a vibration-based machine

health monitoring program will always include diagnosis of machinery defects through

analysis of the measured vibration characteristics.

3.5 VIBRATION MEASUREMENT AND PARAMETERS

Vibratory motion of rotating machines is normally measured by means of a transducer

connected to a processing instrument such as Fast Fourier Transform analyzer. The

transducer may be mounted on the machine bearings or mounted pointing at the rotating

shaft. The transducer picks up the vibratory signal for subsequent processing.

Depending on the complexity of the rotating machine and the anticipated type of defects,

the type and location of vibration transducers may be decided. Furthermore, the appropriate

type and level of signal processing and subsequent analysis method can be adopted.

An extreme method might require complicated signal filtering, time averaging, frequency

analysis, re-synthesis and then statistical distribution analysis. For general condition

monitoring, the processed vibration signal is normally quantified in terms of amplitude,

frequency and phase.

3.5.1 VIBRATION AMPLITUDE

Vibration amplitude is normally expressed as an assessment of the overall severity of the

vibrations. It may be expressed in displacement (microns), velocity (mm/s) or acceleration

(g), depending on the type of information required and the type of defects being monitored.

The measurements may be expressed as Root Mean Square (RMS), Peak (pk) or peak to

56

peak (pk-pk) depending on the type of instrumentation and the standards being applied.

The overall vibration amplitude is often a measure of the average vibration status. Since all

machines vibrate, every machine will have an acceptable level of vibration. This is often

the value set as Alarm Level in vibration monitoring.

In setting the alarm level, one may use;

Machinery manufacturer‟s specification .

International standards such as ISO 2372, BS 4675, VDI2056, etc.

Machinery maintenance history or experience from other similar machines.

3.5.2 VIBRATION FREQUENCY

This is a measure of the number of complete cycles of a vibration signal pattern expressed

as cycle per minutes (CPM) or cycle per second (Hz). Most vibration signals are complex,

comprising more than one frequency. Using Fast Fourier Transform (FFT), a complex

signal can be separated into a spectrum of individual sinusoidal components, each with its

own frequency and amplitude.

Every machine, running in good condition, will have its particular frequency spectra, often

referred to as the signature, as shown in figure 31. Analysis of the spectra ultimately aids in

defect diagnosis (or understanding the cause of vibrations)

Fig 31: Vibration signature. [19]

57

3.5.3 VIBRATION PHASE

This is a measure of the relative position of a vibrating part at a given time with respect to

another or a fixed point. The information is useful in diagnosing machinery defects.

Various defects in a rotating machine cause cyclic forces, and depending on the response of

the machinery structure, the resultant vibrations generated will exhibit particular

characteristics.

In general, the rotating element (rotor), as a mass rotates with some residual imbalance

(irrespective of how well balanced it may be, such as to ISO 1940). This generates the

centrifugal force that strains against the shaft, bearings and support system, which together

form the stiffness/damper system. Amongst the defects that generate vibrations in rotating

machines are;

Unbalanced rotors

Cracked shafts

Misaligned rotors

Bent rotors

Worn bearings

Damaged gears

Some defects are more aggravated by others. Furthermore, high vibrations cause more

defects, resulting into a cause and effect vicious cycle, which would rapidly lead to failure

if not checked.

58

CHAPTER FOUR

4. PRACTICAL DATA FROM SITE The following is the information gotten about the operations and maintenance of the pumps

under discussion;

The maintenance of the low lift pumps is done on a preventative basis. Preventative

maintenance is time driven type of maintenance and the tasks are based on elapsed time or

hours of operation. The workers inspect the pumps daily to notice any irregularity. This

program assumes that machines will degrade within a time frame depending on its

particular classification.

4.1 BEARINGS

The bearings used are the 22315 spherical roller bearings on the drive end and NU 218 EC

single roller bearing on the non-drive end. The spherical roller bearing is the out board

while the single roller bearing is the in board. On the other three pumps, the NU 218 EC

has been replaced with 22218 single roller bearing and their lubrication system has been

changed to oil lubrication in pursuit of better performance. Figure 32 shows the spherical

and single roller bearings.

22315 spherical roller bearings NU 218 EC single roller bearing

Fig 32: Bearings used for low lift pumps.[18]

For specifications of NU 218 EC refer to table1.

The specifications of the bearings are as shown in table 4;

59

Table 4: Bearing parameters [18]

BEARING

22135-Drive end

22218-Non Drive end.

TYPE OF ROLLING

ELEMENT

SPHERICAL ROLLER SPHERICAL ROLLER

PRINCIPAL DIAMETER: D= 160mm, d= 75mm

B= 55

D=160mm, d= 90mm

B=40

BASIC LOAD RATING C= 490 kN

Co= 475 kN

C= 325Kn

C0=375Kn

FATIGUE LOAD RATING

Pu= 48 kN

Pu= 39 KN

REFERENCE SPEED

3 200 r/min

3 800r/min

MASS

5,55 kg

3,40 kg

The bearing modification however, has not met the expectation for which it was intended.

The failure rate still remains unchanged and has brought about oil contamination on the

environment. The oil at time leaks due to improper sealing and mixes with water that flows

back into the river. This poses the danger of the same oil being sucked in by the pumps

thereby causing more damage on the pipe lining. The bearings on the drive end are fitted

with the aid of a hydraulic machine whilst the non-drive end bearing is manually fitted. The

single roller bearing supports the axial load while the double spherical roller bearing

supports both the axial and radial loads. In case of failure, the double spherical roller

bearing is the one that gets damaged often. In addition, the bearing house gets heated up

that they have even made a water cooling system for it externally by means of a pipe

running water over the housing as illustrated in figure 33.

60

Fig 33: Pump with external water cooling system

Without the incorporation of the external water cooling system, the bearing would fail

within a day. Figure 34 is the diagrammatic representation of the low lift pump;

Fig 34: Low lift pump

61

Key;

1. Suction nozzle. 2. Inspection box. 3. Impeller. 4. Casing. 5. Single roller bearing.

6. Shaft. 7. Spherical roller bearing. 8. Bearing housing. 9. Seal

4.2 LUBRICATION

The pumps use two types of lubricants, namely; GX 350 oil and Energrease LS3-NLGI 3

(multipurpose industrial grease) lubricants. Table 5 shows the technical specifications and

characteristics of the Energrease LS3-NLGI 3.

Table 5: Technical specifications and characteristics.

CLASS

Energrease LS3-NLGI 3

DESCRIPTION

All purpose industrial and automotive

BASE OIL VISCOSITY

At 40oC--120mm

2/s

At 100 oC--12mm

2/s

TEMPERATURE LIMITS LTL1)

-30oC

HTPL2)

+120

THICKENER/ BASE OIL Lithium soap/mineral oil

Initially all the pumps used grease lubricant until the bearing housing on three pumps

where modified by introducing a seal to allow the use of oil as a lubricant. One pump with

the bearing types 22315 double spherical roller bearing on the drive end and NU 218 EC

single roller bearing on the non-drive end is grease lubricated while the other three with

bearing types 22315 double spherical roller bearing and 22218 single roller bearings are oil

lubricated. The lubrication is done at least once every two days.

4.3 SHAFT ALIGNMENT

Before the pumps are commissioned, it is ensured that the motor shaft and the pump shafts

are properly aligned and that no damage is caused when the pump shaft is turned. The

coupling is flexible type of coupling and the halves are free to move relative to each other.

62

The shaft can be misaligned angularly or radially. Radial misalignment (also known as

offset or parallel misalignment) refers to the distance between two shaft centerline and is

generally measured in thousandth of an inch. Theoretically, radial misalignment is

measured at the coupling centerline.

Angular misalignment refers to the condition when the shafts are not parallel but are in the

same plane with no offset. For this type of misalignment it is possible for the mating shafts

to be in the same plane at the coupling face intersection but to have an angular relationship

such that they are not collinear.

The alignment is checked with the aid of a laser alignment device which offers the

advantage of eliminating mechanical linkage and run out gauges. This greatly increases the

accuracy and repeatability of the data obtained. Laser alignment system incorporates a

micro processing unit, which eliminates recording errors. Figure 35 shows a Laser

Alignment System.

Fig 35: Laser Alignment System. [15]

Table 6 shows alignment readings for the low lift pump before commissioning it;

63

Table 6: Alignment results.[Alfred H Knight- alignment report].

Position

Direction

Readings in mm

Vertical

Parallel

0.037

Angular

0.025/100

Horizontal

Parallel

-0.050

Angular

0.010/100

4.4 VIBRATIONS

Before the commissioning of the pumps, the vibrations are checked and it is ensured that

they are within accepted range. A vibration analyzer measures the amplitude, frequency

and phase of vibration. Also when vibration occurs at several frequencies, it separates one

frequency from another so that each individual vibration characteristic can be measured.

The vibration pickup senses the velocity of the vibration and converts it into an electrical

signal. The analyzer receives this signal, converts it to the corresponding amplitude and

frequency. The amplitude is measured in terms of peak-to-peak displacement in mils (1 mil

= .001") and is indicated on the amplitude meter. Vibration monitoring as a condition

monitoring technique, is the continuous or regular measurement and trending of

mechanical vibrations of a machine for the purposes of ascertaining the machine‟s dynamic

behavior.

64

Therefore a vibration based machine health monitoring program would consist of:

Vibration measurement and recording (data collection).

Analysis of the vibration severity, characteristics and trend (vibration analysis).

Diagnosis of the machinery defects.

Table 7 shows some of the vibrations readings gotten for the low lift pumps.

Table 7: Vibration readings.[ Alfred H Knight- vibrations report ].

Measurement Points

Vibration Levels in mm/s pk

Bearing

Location

Direction Alarm

Level

JAN APR MAY JUNE JULY AUG

1

V 4.0 3.4 2.2 2.1 1.9 2.1 2.0

H 4.0 2.3 2.5 1.5 1.7 1.5 1.2

A 4.0 3.1 2.9 2.0 2.0 2.2 2.2

2

V 4.0 4.3 3.0 3.1 2.7 3.8 2.6

H 4.0 3.3 2.8 2.0 2.2 2.8 2.4

A 4.0 2.1 3.4 1.3 1.2 2.3 1.5

3

V 8.0 6.6 6.5 3.5 3.8 5.1 5.4

H 8.0 6.1 5.6 3.8 3.8 5.0 4.6

A 8.0 3.5 5.4 3.8 3.6 4.8 3.2

Spike

Energy

1.0 0.22gSE 0.45gSE 0.37gSE 0.55gSE 0.21gSE 0.18gSE

4

V 8.0 7.1 5.3 2.6 3.5 5.3 3.5

H 8.0 3.5 3.8 2.7 2.9 3.7 3.2

A 8.0 6.2 4.1 3.8 4.5 3.1

Spike

Energy

1.0 0.62gSE 0.64gSE 0.46gSE 0.49gSE 0.58gSE 0.44gSE

Abbreviations:V-Vertical; H-Horizontal ; A-Axial ; NDE-Non Drive End ; DE-Drive End

65

The Vibration Spectra and Trends are as follows;

Fig 36: Vibration spectra and trends. [Alfred H Knight - vibrations report].

From the vibration readings, a report is prepared which summarizes and gives the

diagnosis/recommendations based on the results. This is as shown in table 9;

66

Table 8: diagnosis report on vibrations [Alfred H Knight- vibrations report ].

Machine

Train

Current

Vibration

Severity

Diagnosis/Recommendations

Low Lift

Pump

No.1

Fair

Fair vibration levels with a fairly stable trend. No

significant defect at the moment.

Continue periodic tests, monthly, to monitor the

trend.

Always ensure adequate and frequent lubrication of

the bearings.

Ensure correct bearing sealing to prevent ingress of

water

Low Lift

Pump

No.2

Fair

Fair vibration levels with a fairly stable trend. No

significant defects at the moment.

Continue periodic tests, monthly, to monitor the

trend.

Always ensure adequate and frequent lubrication of

the bearings.

Ensure correct bearing sealing to prevent ingress of

water

Low Lift

Pump

No.3

Good

Pump overall vibrations still good with no significant

defects.(The spike energy on the pump bearing is still

low)

Continue periodic tests, monthly, to monitor the

trend.

Always ensure adequate and frequent lubrication of

the bearings.

Ensure correct bearing sealing to prevent ingress of

water

Indications of deterioration in the mechanical

67

Low Lift

Pump

No.4

Slightly Rough

looseness on the pump DE hold down bracket.

The hold down bolts on the pump DE hold down

bracket, often works loose.

This must be re-tightened frequently. The current

fastener should be replaced with one that has only

one strong washer. The excess washers and nuts

should be removed. Note that if this is not attended

to, the bracket would continue loosening and the

vibration levels might deteriorate again.

4.5 MATERIALS USED

The materials used for the pump and pipe are as shown in table 9.

Table 9: pump and pipe materials

Component Material

Impeller Bronze

Casing Cast iron

Shaft Steel

Pipes Mild Steel

4.6 TEMPERATURE MEASUREMENTS

The bearing - housing temperatures measured using an Infra-Red Thermometer, are as tabulated

below;

Table 10: Temperature reading results

PUMP No.

TEMPERATURE READING ( 0C )

1

46.8

2

40.8

3

38.8

4

29.0

68

4.7 PIPE NETWORK

The pipe layout is key to the life of the pump‟s proper operation. Pipe layout influences

pressure exerted on the bearings of the pump. With the installation of new pumps, the pipe

network was also changed.

The new pipe network has two flanges, a suction valve and an eccentric increaser on the

suction side. It further has two flanges, ten turns and two valves on the discharge side. The

pipe network has a vertical height of 14m with a horizontal length of 93m. The pipe

network is shown in figure 37.

Table 11 and 12 show the number of fittings on the pipe network and their corresponding

K- values.

Table11: Fittings on the suction side.

Type of fitting Quantity K-value

Flange 1 0.04

Check valve 1 0.9-1.7

Eccentric increaser 1 1

Table12: Fittings on the discharge side.

Type of fitting Quantity K-value.

Flange 2 0.04

Valve 2 0.9-1.7

90° elbow 4 0.25

45° elbow 5 0.18

T Junctions 2 0.75

Concentric increaser 1 1.16

Concentric reducer 1 1.16

69

CHAPTER FIVE

5. RESULTS AND ANALYSIS

5.1. BEARINGS

The bearing parameters are within the working parameters of the pump. They are the right

application and have an appropriate speed range. The type of bearing used i.e. spherical

roller bearing can withstand both axial and radial forces. This makes them a suitable

application for the low lift pumps.

However, the bearings are under a risk of being over heated all the time that an additional

cooling system of a water pipe has been improvised. Without which they would fail within

a day. Heat is caused by friction of the moving components of the pumps. This is

minimized by lubrication which prevents body to body contact of rotating elements.

Lubrication creates a smooth motion of bearings with a minimum clearance at maximum

load. However, if the lubricant is lost or overheated its effect is absent thereby leading to

increase in friction which produces a lot of heat and burns the bearing. Some of the effects

of heat on the lubricant are;

Decrease in viscosity; heat reduces the viscosity of the lubricant which reduces its ability

to support the load. This reduction in viscosity means more asperity to asperity contact

which leads to more friction and more heat being produced.

Formation of a "varnish" residue and then "coke" at the elevated temperature. This

"coking" will destroy the ability of the grease or oil to lubricate the bearing. It will also

introduce solid particles into the lubricant creating a three body abrasion and causing a

rise in temperature.

In addition to the heat generated by overloading, overheating of bearings also comes about

as a result of oil leakage. The pumps are lubricated using oil which comes out due to loose

seal fittings that close the bearing housing. This usually occurs in the night when the

pumps are un-monitored. The deficiency in lubricant is evidenced by the state of bearing

when failure occurs; they are burnt out with no lubricant and at times weld to each other

and to the raceways. This is shown in figure 38.

70

(a)

(b)

Fig 38: (a) welded bearing components, (b) Decolorized bearing due to heat

The bearing is decolorized due to excessive heat which comes about as a result of friction

from body to body-contact of the bearings rolling elements with the races and usually

components get welded to each other i.e. bearing parts or inner race to shaft. This is

illustrated in figure 39.

Besides losing of lubricant, water also mixes with the lubricant due to improper design of

the modified lubricant housing. Because the bearing seal is not tight, water gets inside the

bearing or the oil leaks out.

71

Fig 39: Shaft with welded inner race

The presence of water is evidenced by the white emulsions formed when the water reacts

with the lubricant; this reaction makes the lubricant of no effect. This is due to the fact that

water and oil emulsion does not provide a good lubricating film and easily gets evaporated

leaving the bearings to run dry and burn out. At times this is noticed before failure and the

white emulsions are found and removed, however before long, the same situation occurs

and may go unnoticed as it has been the case.

5.2 LUBRICATION

The frequency of lubrication which is on a daily basis has contributed to the overheating

of the bearings. According to the maintenance manual the bearings are supposed to be re-

lubricated after every 1000hrs, which is 41days of operations. Initially, the pumps were

designed to be grease lubricated but this was later on changed to oil lubrication by

modifying the bearing housing. This was in an attempt to reduce on overheating.

However, this has resulted in barely any improvement and has led to other problems such

as oil leaks and contamination of the river. In additional, the housing seal allows entrance

of water into the bearing, making the lubricant to be of no effect. This contributes largely

to the failure of the bearings as the bearings are starved of the lubricant.

5.3 VIBRATIONS

Vibrations are continued periodic oscillations relative to a fixed reference point, it is

moving back and forth very rapidly of a machine. This is one of the main side effects of

72

wrong impeller application and hence the cause of failure. Vibrations come about as a

result of cavitation, which causes unbalanced rotation of the impeller. The number of

vanes on the impeller in clear water pumps i.e. four and above, helps to create a smooth

flow for the water and share the radial thrust evenly. The more the number of impeller

vanes, the easier it is to overcome the radial thrust in the casing without it being further

transmitted to the rotating elements of the pump i.e. shaft and bearing. Therefore, wrong

impeller selection contributes to vibrations and reduced bearing life. Pump vibrations at

Bulangililo water plant are amplified by the worn out mounting bolts. This leads to

reduced holding capability of the mounting of the pump. Consequently, bearing failure

ensues owing to overloading due to misalignment between the pump and its driver.

At the point of commissioning, the pump is well aligned and firmly held but not for long

due to constant vibrations which are amplified with time thereby leading to a brittle

fracture type of failure of bearings. The incapability to hold by the mounting bolts is

evidenced in the laser alignment tests where the major recommendation is frequent

tightening of the pump base and constant monitoring. In addition, the vibration levels are

further evidenced by the physically felt impact on the pipe system and the noise produced.

This is shown in the type of failure in figure 40 (a) and (b).

(a).

73

(b).

Fig 40.Brittle fractured bearings

From the visual inspection of the damaged bearings, the following were observed;

The failure is not gradual as there was no sign of continuing wear, the failure was

sudden and lead to a brittle failure.

The surfaces are shiny.

The surfaces had chevron marks at the point where the fracture started from.

This kind of failure is either due to brittle material or due to sudden impact loading.

5.4 TEMPERATURE

As long as the lubricant is there and the external pipe for cooling is running water, the

temperatures are within acceptable range as shown in table 12. This however, is prone to

changes due to lose of lubricant and ingress of water in the lubricant. This leads to body

to body contact which results in increased friction and rise in temperature. If uncontrolled

this could lead to burning of bearings and welding together of bearing components as

shown in figure 38 (a) and figure 39.

5.5 SHAFT ALIGNMENT

On commissioning, the motor and pump shaft are well aligned as indicated by the shown

readings in table 7. However, because of the loose mounting bolts, the firmness of the

pump is lost leading to increased vibrations. These vibrations lead to sudden

74

misalignment which results into brittle fracture of the bearing. Refer to figures 40 (a) and

( b).

5.6 PIPE NETWORK

The pipe network is used to come up with the system curve. The system curve is a graph

that is used for determining the correct pump to use by superimposing with the pump

performance curve. The following are the calculations for the system head.

CALCULATING THE ACTUAL SYSTEM HEAD

(i) System Head When Q=1350m3/hr. = 0.375 m

3/s

The system head is obtained from;

H =

Where = Static head.

= Frictional losses in the pipe.

= Velocity head loss.

= Other losses

1. Velocity head loss

HV = +

=

+

Where = suction velocity

= discharge velocity.

Now

Vs =

=

=

=0.0995 m2

Thus,

Vs =

75

= 3.769 m/s

And

=

Note: the discharge has two diameters such that = .

Hence, = 3.769m/s

=

= 0.203m

2

=

=

=1.847 m/s.

Thus,

=

=

= 0.724 m.

=

=

+

=0.8979 m.

= + = 0.724 + 0.8979 = 1.6219 m.

2. Frictional losses in the pipe

=

=

= 0.16921 m

=

=

+

76

=0.6833 + 1.9058

=2.589 m

= +

= 0.1692 + 2.589

= 2.7582 m

3. Other losses

= +

= for 2-T junctions, 4,900 elbows, 5,45

0 elbows.

=

has two components, 0.724 m and 0.174 m. The 2-T junctions are in

the region with 0.724 m, 3,90

0 elbows are also in the region

0.724 m

while the others are in the region 0.174 m.

Thus,

= 2×0.75 ×0.724 +3×0.25×0.724 +0.25× 0.174 +5×0.18×0.174

= 1.086 + 0.543 + 0.0435 + 0.1566

= 1.8291 m

= +

= 0.9×0.724 + 2×0.90×.724

= 0.6516 + 1.3032

=1.9548 m

= +

= +

= 0.04×0.724+ 0.04× 0.724 + 0.04×0.174

= 0.02896 + 0.02896 + 0.00696

77

= 0.0649 m

=

= 1.16 × 0.174

= 0.2016 m

=

= 1.16 × 0.174

= 0.2016 m

=

= 1×0.724

= 0.724 m

Entrance losses; =

=

= 0.5 × 0.724

= 0.362 m.

Exit losses; =

=

= 1 × 0.174

= 0.174 m.

Thus, 1.8291 +1.9548 + 0.0649 + 0.2016 +0.2016 + 0.724 + 0.362 + 0.174

= 5.512 m

4. Static head

= 14 m

78

The system head is;

H =

= 14 + 2.7582 +1.6219+5.512

= 23.8921

H = 23.89 m

ii. CALCULATING HEAD AT FLOW RATE OF Q = 2025M3/HR = 0.5625M

3/s

1. Velocity head Loss.

Hv = Hvsuction + Hvdischarge.

=

Now

=

= 0.0995m2

Thus

Vs=

=5.653m/s

And

Vd1=

where ad1=as

Vd1= 5.653 m/s

79

Vd2 =

ad2 =

= 0.203 m2

Vd2 =

= 2.771 m/s

Hv= Hvs + Hvd1 + Hvd2.

=

=

= 1.62877 + 1.62877 + 0.39136

= 3.6488978

Hv= 3.6489m

2. Other Losses.

Hf elbows =Hfsuct.elbows + Hfdisch.elbows

Hf for 2 T-junctions, 4, 90oelbows, 5, 45

o elbows

= Ks.Hvds + Kd1.Hvd1 + Kd2.Hvd2

=

= 4.11491m

Hf valves= Hf.suct.valves+ Hf.disch.valves

=Ks HVs + Kd Hvd

=

80

=1.46592 + 2.93184

=4.39776m

Hf.flanges= Hf.suct+ Hf.disch

=Ks Hvs + Kd Hvd

=

=0.065152 + 0.065152 + 0.015656

=0.14596m

Hf. concentric increaser = Kd2. Hvd2

=

=0.454024m

Hf concentric reducer = Kd2 . Hvd2

=

=0.454024

=0.454m

Hf eccentric increaser = Ks .Hvs

=

=1.6288m

Entrance Losses; hen = K

=K. Hvs

=

81

=0.8144m

Exit Losses; hex = K

=K. Hvd2

=

=0.3914m

HO= Hf.elbows+Hf.valves+Hf.flanges+Hf.concentric increaser+Hf.concentric reducer +Hf.eccentric increaser

+ hen + hex

=4.1149+4.3978+0.14596+0.454+3.2576+0.8144+0.3914

=12.40126m

3. Frictional Losses in the Pipe.

Hf =

=

= 0.35137 +1.53726 + 4.64699

= 6.53562m

4. Static Head

Hs = 14 m

H = HS + Hf +Hv + Ho

=14 + 6.53562 + 3.6489 + 12.4013

=36.5858.

Head = 36.5858m.

82

iii. CALCULATING SYSTEM HEAD WHEN Q=675M3/HR = 0.1875 M

3/S

The system head is obtained from;

H =

Where = Static head.

= Frictional losses in the pipe.

= Velocity head loss.

= Other losses

2. Velocity head loss

HV = +

=

+

Where = suction velocity

= discharge velocity.

Now

Vs =

=

=

= 0.0995 m2

Thus,

Vs =

= 1.884 m/s

And

=

Note: the discharge has two diameters such that = .

83

Hence,

= 1.884m/s

=

= 0.203m2

=

=

= 0.924 m/s.

Thus,

=

=

= 0.1809m.

=

=

+

= 0.2244 m.

= +

= 0.1809 + 0.2244

= 0.4053m.

2. Frictional losses in the pipe

=

84

=

= 0.1707466 m

=

=

+

= 0.1707466 + 0.5167049

= 0.6875 m

= +

= 0.1707 + 0.6875

= 0.8582 m

2. Other losses

= +

= for 2-T junctions, 4,900 elbows, 5,45

0 elbows.

=

has two components, 0.1809 m and 0.0435 m. The 2-T junctions are

in the region with 0.1809 m, 3,900 elbows are also in the region 0.1809 m

while the others are in the region 0.0435 m.

Thus,

= 2×0.75 ×0.1809 +3×0.25×0.1809 +0.25× 0.0435 +5×0.18×0.0435

= 0.27135 + 0.135675 + 0.010875 + 0.03915

= 0.45705 m

= +

= +

= 0.9×0.1809 + 2×0.9×0.1809

85

= 0.16281+ 0.32562

= 0.48843 m

= +

= +

= 0.04×0.1809+ 0.04× 0.1809 + 0.04×0.0435

= 0.007236 + 0.007236 + 0.00174

= 0.016212 m

=

= 1.16 × 0.0435

= 0.05046 m

=

= 1.16 × 0.0435

= 0.05046 m

=

= 1×0.1809

= 0.1809 m

Entrance losses; =

= = 0.5 × 0.1809 = 0.09045 m.

Exit losses; =

= = 1 × 0.0435 = 0.0435 m.

Thus, 0.45705 + 0.48843 + 0.016212 + 0.05046 + 0.05046 + 0.1809 + 0.09045 +

0.0435

= 1.377462m

4. Static head

86

= 14 m

The system head is;

H =

= 14 + 0.8582 + 0.4053+ 1.377462

= 16.640962

H = 16.64 m

5.2 DRAWING OF SYSTEM CURVE

From the obtained values of head versus flow, the system curve was drawn as shown in

figure 41.

Fig 41: System curve

87

This was drawn based on the values from table 13.

Table 13: Head and flow rate values of system curve

Head

values(y-axis)

14 16.64 23.89 36.59

Flow rate

Q(x-axis)

0 674 1350 2025

From the pump manual book, the following was the given pump characteristic curve i.e.

figure 42.

Fig 42: pump curve

This is drawn from the values of heads and flow rates values given in table 14;

Table 14: Head and flow rate values of pump curve.

88

Head

(m)

32 30 25 23.5 21.5 20 18

Flowrate

(m3/hr)

0 200 850 1140 1350 1500 1660

In order to know the actual head and flow rate of the pump, we superimpose the pump and

system curves and take note of the point of intersection. This is shown in figure 43.

Fig 43: combination of system and pump curve

89

CHAPTER SIX

CONCLUSION AND RECOMMENDATIONS

Having considered all the available parameters as was necessary, the causes of the failures

of the low lift pumps could be summarized as follows;

6.1 ROOT CAUSE OF FAILURE.

Principally, the design of the pump with specific reference to the impeller is not suitable

for the low lift clear water pumping. It is for this reason that the pump has been failing

time and again. In normal operations and correct application of a pump, pumps are able to

run for over six months without giving any problem. However, this has not been so for the

low lift pumps which have failures every month. From the technical information on

pumps, the design of the pumps is that of a slurry pump. A slurry pump is a device that

impels a mixture of water and insoluble solid material such as clay, sewer or cement. A

slurry pump design is clearly distinguished from a clear water pump by its impeller

design. The following properties qualify an impeller as a slurry impeller;

Clear water impellers are designed with vanes starting from four (4) going upwards,

on the other hand slurry pumps, depending upon capacity incorporate from one

(1)to four (4) vanes.

Fig 44; Impeller with two vanes curved from the eye

90

The vanes are not straight, but describe a smooth curve that begins at the impeller‟s

eye and extends to its periphery. They may also be curved upward at their entry.

These details are evident in the impeller as shown in figure 40. This is the impeller

used at the low lift pumps of Nkana water and sewerage company.

Comparing the current pumps with the previous pumps gave a further large contrast

because the old pumps had eight (8) vanes and their performance was

recommendable. In additional, it is worth noting that the sewage works in Nkana

East actually uses a similar design of pumps which has two vanes and works just

well for the sewage application.

The throughlet size of the impeller suits that of a slurry impeller. Throughlet size is

defined as the open internal passage through the impeller that, ultimately

determines the largest diameter solid that can pass. All impellers regardless of their

design has a throughlet size, but in order to maximize throughlet size, solids

handling impellers limit the number of vanes so that the passage between them can

be as large as possible.

Small to medium sized sewage pumps are often referred to as non-clogs and their

impeller are designed to live up to that name. The pumps under discussion came

with an operation and maintenance manual for „Horizontal End Suction Non-clog

Centrifugal pumps.‟ it is therefore evident that the design falls under slurry

classification of pumps.

From figure 42, it is evident that the head against which the pump of 21.57m head

is working is 22.5m. This further proves that the pump is not also the correct pump

size for this application. The pump is meant to operate at a flow rate of 1350m3/hr

but is operating at a flow rate of 1230 m3/hr. These two parameters confirm the

wrong application of the pump.

6.2 RECOMMENDATION

Having conducted this research on the failures of the low lift pumps and considering the

findings, we would like to recommend the following;

The current pumps should be replaced with clear water pumps which have more

than four impeller vanes. It should be ensured that the pumps have a capacity which

91

matches with the system requirement. The pumps should be selected based on how

well the system curve matches with the pump performance curve. The pump

operating point will be identified as the point, where the system curve crosses the

pump performance curve when they are superimposed on each other. The pump

must have the following specifications;

o Head-25m-30m

o Flow rate 1350m3/hr

o Speed >1450rpm

Mounting bolts should be secured by introducing spring washers to improve on

anchorage or by any other appropriate means. Vibration monitoring program and

more frequent planned maintenance should be instituted to avoid catastrophic

failures.

Many bends in the pipe network should be avoided to minimize frictional losses and

it should be ensured that the pipe discharges water from the top of the discharge

reservoir.

The bearing must not be over lubricated as is the cases at present were lubrication is

being done on almost a daily basis. Lubrication must be done after every 1000hours

of operations which is 41 days or one month and eleven days. Over lubrication

leads to overheating whose effects is totally undesired. As for type of lubricant to

use, we recommend grease to be used instead of oil. Grease has the advantage of

being easily retained in the bearing arrangement and contributes to sealing of the

arrangement against contaminants, moisture or water.

Oil lubrication is used in applications which have high speed while grease

lubrication is used in application with lower speeds like our case at low lift pumps,

Bulangililo. This means that for modification to have fully been compatible with

system, the speed should have been changed as well.

We recommend that the company uses a predictive type of maintenance. Predictive

maintenance is a condition driven preventative maintenance program. It uses direct

monitoring of mechanical conditions, pump efficiency and other indicators to

determine the actual loss of efficiency for the pumps in the plant. The predictive

maintenance ensures the maximum interval between repairs and minimizes the

92

number and cost of unscheduled outages created by pump failures. A

comprehensive predictive maintenance management program utilizes a combination

of the most cost effective tools, which is vibration monitoring, thermography,

tribology etc.

Bearings should be installed in a dry, dust-free room away from metalworking or

other machines producing dust. To protect the bearing and mounting position from

contamination by dust, dirt and moisture, bearings should be covered or wrapped

with waxed paper or foil.

6.3 CHALLENGES

The following were the challenges faced in the research;

The plant does not have most measuring instruments such as tachometer, flow-

meters and pressure gauges. This made it hard for collection of the actual

operating parameters.

The pumps under investigation had insufficient design literature and did not

give all the required information for operation such as type of lubricant to be

used.

Effort to contact the pump manufacturers was futile as there was no email

address provided and their website offered no credible means of communicating

with them.

6.4 CONCLUSION

In conclusion, having considered the information concerning centrifugal pumps and

the operating conditions of Bulangililo water works, the current low lift pump are

unsuitable as they have the impeller type which is typical characteristic of

slurry/sewer pumps. From the superimposition of the pump and system curves the

actual head and flow rate are different from the pump rating, proving that the pump is

not the correct size as while. The miss application of the pumps has been the core

reason of the high rate of failure besides operational errors like oil leakage. We

sincerely hope the pumps will be worked on as soon as possible for the benefit of the

company, the workers and the community at large.

93

6.5 APPENDIX

PROBLEM POSSIBLE CAUSE

PUMP OPERATES WITH NOISE OR

VIBRATIONS

Misalignment between pump and driver

Rotating parts rubbing against stationary parts

Worn bearings

Incorrect direction of rotation

Damaged impeller or incorrectly mounted

Impeller or casing partially filled with solid matter

System requirements too far out on head-capacity

curve

Suction strainers filled with solid matter

Incorrect layout of suction sump

Air enters pump during operation

Piping imposes strain on pump

Rotating elements not balanced

Excessive radial forces on rotating parts

Too small distance between impeller outer diameter

and volute tongue

Bent shaft misalignment of pump parts

Faulty shape of volute tongue

Improper installation of bearings

Improper lubrication of bearings

Obstruction to flow in suction or discharge piping.

BEARINGS FAIL PREMATURELY Damaged impeller

Rotating elements not balanced

Excessive axial and radial forces on rotating parts

Bent shaft

Misalignment between pump and driver

Pump operates for prolonged time at low flow rate

Cracked or damaged bearing housing

Improper base plate or foundation

Faulty lubrication system

Excessive grease in the bearings

Dirt or water ingress into bearing housing

Excessive suction pressure

Inadequate cooling of bearing

PUMP OVERHEAT OR SEIZES Pump allowed to run dry

Simultaneous operation of poorly matched pumps

Worn or damaged bearings

Poor lubrication

Internal misalignment from too much pipe strain,

poor foundations, or faulty repair work

94

6.6 REFERENCES

1. L. Bachus and A. Custodio (2003), Know and Understand Centrifugal Pumps,Elservier Ltd,

Britain.

2. Ron palgrave (2003), Troubleshooting Centrifugal Pumps and their systems, Elsevier Ltd,

Great Britain.

3. R.Rayner (2006), Pump Users Handbook, 4th Edition, Elsevier Advanced Technology Ltd,

Great Britain.

4. PareshGirdhar and Octo Moniz (2005), Practical Centrifugal Pumps, Software services Pvt.Ltd,

Netherlands.

5. A.S.Rangwala (2005), Turbo-Machinery Dynamics, McGraw Hill Companies, United States of

America.

6. E.T.George (2006), Handbook of Lubrication and Tribology, 2nd

Edition, Taylor and Francis

Group, United States of America.

7. Nestor Perez (2004), Fracture Mechanics, kluwer Academic Publishers, United States of

America.

8. Michael Clifford (2010),An introduction to Mechanical Engineering, Hodder Education,

United Kingdom.

9. John Moubray (1997), Reliability – Centered Maintenance, 2nd

Edition, Butterworth –

Heinemann, United Kingdom.

10. KugelfischerGeog (1989), FAG Rolling Bearings – Standardprogramme, Catalogue WL

41510/2EA, Germany.

11. Winterthur (2007),Sulzer centrifugal Pump Handbook, 2nd

Edition, Elsevier Ltd, Switzerland.

12. Robert L. Sanks, Pumping Station Design, 2nd Edition, Butterworth Heinemann, USA.

13. Myles. K (1996), Pumps Principles and Practice, 3rd Edition, Sulzer, Perskor printers Ltd,

South Africa.

14. www.cheresources.com/centrifugal pumps. 20th October, 2011. 23:00hrs

15. John Piotrowski(2007), Shaft Alignment Handbook, 3rd

Edition, Taylor and Francis group,

United State of America.

16. www.applitudexchange.com/bearings in centrifugal pumps.20th October, 2011.23:30hrs.

17. Keith Mobley(1999), Vibrations Fundamentals, Butterworth-Heinemann,United States of

America.

18. SKF General Catalogue, (2005),Germany.

95

THE PIPE LAYOUT OF THE PUMPING SYSTEM FOR THE LOW LIFT PUMPS.

The pipe network has a vertical distance (static head) of 14m, it has a horizontal distance of 93m and the

total length of the discharge pipe is 123m.

Fig 37 The pipe network of the Bulangililo water works.