Pump

2PREFACE

This training manual on pumps is intended to be used for basic

skill training for Scientific Assistant trainees. This manual is prepared

with a view of imparting training effectively on Function, Construction,

Material of parts, and maintenance aspects of the pumps. This training

manual will also be used for Tradesman trainees. It has been prepared in

view the important aspect of the subject and bring the competency in the

new entrants so that they can do right maintenance at the first attempt.

I express my sincere thanks to Shri Prashant Puri, STO for

giving his valuable suggestions during preparation of this manual and to

Shri C.M.Mishra, ENC(MT) for his guidance and encouragement to

complete this task. I owe my sincere gratitude to Shri N.Nagaich,

Training Superintendent, RAPS-1 to 4 for his kind cooperation and

motivation in preparation of this manual.

R P SainiSO/E

Nuclear Training Center

3CONTENTS

S. NO. DESCRIPTION PAGENO.

Chapter-1 Pumps1.0 Introduction 31.1. Classification of Pumps 3

Chapter-2 Centrifugal Pumps2.0 Working Principal 42.1. Main Parts of a Centrifugal Pump 42.2. Single and Multi Stage Design: 12

Chapter-3 Terms and Characteristics of Centrifugal Pumps3.1. Suction Head and Suction Lift 143.2. System Head 143.3. Pump Characteristic 173.4. System Head Curves 193.5. Evaluation of Pump Characteristics 203.6. Centrifugal Pump Characteristic Relations 213.7. Specific Speed 213.8. Net Positive Suction Head 223.9. Cavitation 233.10. Parallel Operation of Pumps 243.11. Series Operation of Pumps 25

Chapter-4 Thrusts and its Balancing in Centrifugal Pumps4.1. Radial Thrust in Centrifugal Pump 264.2. Axial Thrust in Centrifugal Pump 284.3. Axial Thrust in Multistage Pumps 36

Chapter-5 Operations of Centrifugal Pumps5.1. Initial Start 395.2. Pipe Cleanliness 395.3. Alignment 405.4. Rotation 405.5. Pump Bearings 405.6. Pump Drive 415.7. Priming 415.8. Position of The Valves 415.9. Pump Warm Up 41

5.10. Starting the Pump 425.11. Operating Checks 435.12. Pump Shutdown 43

Chapter-6 Maintenance of Centrifugal Pumps6.0 General 446.1. Rotating Parts 446.2. Stationary Parts 516.3. Maintenance Records 526.4. Maintenance Intervals 52

4Chapter-7 Rotary Pump7.1. General 557.2. Gear Type Pump 557.3. Screw Pump 577.4. Sliding Vane Type Pump 577.5. Lobe Pumps 58

Chapter-8 Rotary Pump Maintenance8.1. General 598.2. External- Gear Pumps 598.3. Internal - Gear Pumps 608.4. Screw Pumps 648.5. Sliding Vane Pump 65

Chapter-9 Reciprocating Pump9.1. Introduction 679.2. Working of a Reciprocating Pump 679.3. Main Parts of a Reciprocating Pump 689.4. Classification of Reciprocating Pumps 719.5. Air Vessels or Accumulators 72

Chapter-10 Centrifugal Pumps- Trouble Shooting 74Chapter-11 Positive Displacement Pumps- Trouble Shooting 78Chapter-12 Maintenance Procedures

12.1 Centrifugal Pump 8212.2 Gear type Rotary Pump 8512.3 Screw type Rotary Pump 8712.4 Sliding Vane Pump 8912.5 Reciprocating Pump 91

5CHAPTER - 1

PUMPS

1.1 INTRODUCTION

The hydraulic machines, which convert the mechanical energy intohydraulic energy, are called pumps. The hydraulic energy is in the form ofpressure energy. Most process in industries involves the transportation ofliquids, or their transfer from one level of pressure or static energy toanother. The pump is the mechanical means for achieving this transport ortransfer, and thus becomes an essential part of all processes.

1.2 CLASSIFICATION OF PUMPS

Pumps may be classified on the basis of application they serve, the materialfrom which they are constructed, the liquid they handle, and even theirorientation in the space. All such classification, however, are limited in scopeand tend to substantially overlap each other. A basic system of classification isby which energy is added to the fluid. Following tree diagram gives pumpsclassification.

PUMPS

PositiveDisplacement Pump

RotodynamicPump

CentrifugalReciprocatingPump

RotaryPump

Plunger Piston Diaphragm Axial orpropeller

pump

Radial Mixedflow

pump

Gearpump

Screwpump

Vanepump

Radial pistonpump

External Internal

6CHAPTER - 2

CENTRIFUGAL PUMPS

The mechanical energy is converted into pressure energy by means ofcentrifugal force acting on the fluid, the hydraulic machine is calledcentrifugal pump.

2.0 WORKING PRINCIPAL

The centrifugal pump works on the principle of forced vortex flow, whichmeans that when a certain mass of liquid is rotated by an external torque, therise in pressure head of the rotating liquid takes place. It consists of a rotatingelement including impeller & a shaft and a stationary element made up of acasing, stuffing box. Vanes of impeller impart energy to the fluid throughcentrifugal force. The fluid is forced into impeller due to differential pressurebetween pressure at water surface in suction tank and the pressure at suctioneye of impeller. Fluid is discharged through impeller outlet at higher pressure& velocity. The velocity is converted into pressure by means of volute or setof stationary diffuser vanes surrounding impeller.

7Fig. 2.1

2.1. MAIN PARTS OF A CENTRIFUGAL PUMP

The following are the main parts of a centrifugal pump:

1. Shaft 2. Impeller.3. Casing. 4. Bearings5. Stuffing box 6. Wearing rings6. Suction pipe with a foot valve and a strainer. 7. Delivery pipe.

All the main parts of the centrifugal pump are shown in Fig. 2.2 of a singlestage, end suction centrifugal pump as given below.

Fig. 2.22.1.1. SHAFTS: Pump shaft depending upon the type of motor used, the pump shaft

may be a part of the motor rotor, or may be independent and coupled to themotor shaft. Pumps shafts are normally made of stainless steel or othercorrosion-resistant material. Corrosion resistant materials are used because ofthe expense involved when replacing shafts. It is more economical to install ahigh quality shaft despite high initial cost.

The shaft may be independently bearing supported, or may be supported bythe motor bearings. The method used will depend upon the design of thepump.

2.1.2. IMPELLER: The rotating part of a centrifugal pump is called Impeller. Itconsists of a series of vanes. Impeller vanes direct the flow of fluid within thepump. The impeller is mounted on a shaft, which is connected to the shaft ofan electric motor. This applies high energy force to the fluid to give it velocityand momentum. The open area in the centre of the impeller is called theimpeller eye and partially determines the pump capacity.

The pump impeller is the most critical part of a pump because the size, shape,and speed determine the pump capacity under all head conditions.

Impeller shrouds enclose the blades of the impeller and contain the flow offluid in the impeller area. Suction point, where fluids enter the pump, is

8normally located near the centre of the casing. The diameter of the passage atthis point partially determines the rate at which the unit can pump.

The solid center portions of the impeller mounts on the shaft and is called theimpeller hub. Depending upon the design and size of the impeller, the hub canbe quite large or small. If the impeller is of the double suction type, the hubusually is open in the center and two separate hubs will be mounted on theshaft.

Types of Impellers:

Based on flow:

Radial flow : Used for high head and low flow.Mixed flow : Used for medium head & flow.Axial flow : Used for high flow and low head.

Based on construction:

Open type : Used for fluids having debris.Semi-open type: Used for clean fluids. Single shroud provides strength.Close type : Used for clean fluid.

Based on type of vanes:

Forward vane: Not normally used due to high consumption, lessefficiency and drooping characteristics.

Radial vane : Not normally suggested due to improper guiding offluid.

Backward vane : Generally used (reason discussed under pumpcharacteristic)

Impellers are further classified as:

1. Single-suction with a single inlet on one side.2. Double-suction with water flowing to the impeller symmetrically from

both sides, which benefits axial balancing.Several types of impellers are shown in Fig 2.3:.

9Fig. 2.3

The pump impellers can be made of cast iron, cast steel, fabricated steel,bronze, brass, molded rubber, fiberglass or any other material that will becompatible with the material being.

2.1.3. CASING

The casing of a centrifugal pump is an air-tight passage surrounding theimpeller and is designed in such a way that the kinetic energy waterdischarged at the outlet of the impeller is converted into pressure energybefore the water leaves the casing and enters the delivery pipe. The casingencloses the pump impeller, shaft and packing gland. Usually casings are ofthe volute or increasing diameter type.

The following three types of the casings are commonly adopted:

a. Volute casing.b. Diffuser vane casingc. Regenerative Casing

a. Volute Casing:

Fig 2.4 shows the volute casing, which surrounds the impeller. It is of spiraltype in which area of flow increase gradually. The increase in area of flowdecreases the velocity of flow. The decrease in velocity increases the pressureof the water flowing through the casing. It has been observed that in case ofvolute casing, the efficiency of the pump increases slightly as a large casing,large amount of energy is lost due to the formation of eddies in this type ofcasing.Converts kinetic energy of fluid into pressure energy.Designed such as to produce an equal liquid velocity around circumference ofimpeller. When operating at other than its peak efficiency, unbalanced forcesoccurred around circumference of impeller. This results in radial thrust.To overcome above problem of radial thrust double volute casing is used.Double volute casing is similar to two vane diffuser casing.

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Fig. 2.4

b. Diffuser Vane Casing

If a circular chamber is introduced between the casing and the impeller asshown in Fig 2.5, the casing is known as Vortex casing. By introducing thecircular chamber, the loss of energy due to the formation of eddied is reducedto a considerable extent. Thus the efficiency of the pump is more than theefficiency when only volute casing is provided.Compared to volute it is more efficient in converting kinetic energy intopressure energy, because of better guidance of fluid.Less efficient when operating at other than peak efficiency.Hydraulically balance the radial loading of motor.Less changes in impeller size is permissible in the given casing compared tovolute casing.

Fig. 2.5c. Regenerative Casing

This casing is shown in fig in which the impeller is surrounded by a series ofguide blades mounted on a ring, which is known as diffuser. The guide vanesare designed in such a way that the water from the impeller enters the guidevanes without shock. Also the area of the guide vanes increases, thus reducingthe velocity of flow through guide vanes and consequently increasing thepressure of water. The water from the guide vanes then passes through thesurrounding casing, which is in most of the cases concentric with the impelleras shown in fig 2.6.

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Fig. 2.6Types of Casing based on Construction:

1. Horizontally split casing e.g. BFP#1,2 & 3; process water booster pumpetc.

2. Vertically split casing e.g. primary coolant pump. Moderator pump, &BFP#7 etc.

The Horizontal or Axial split is made on the centre line of the shaft andallows the upper half of the casing to be removed completely, giving access tothe pump shaft, impellers and bearings. The term horizontal Does Not refer tothe horizontal Position of the pump of the split casing. If the pump wasmounted in a vertical position, it could still have a horizontal split casing-thename indicates the method of the split rather than the position.

Pump casings are also split Radial. Originally, this was called a vertical split,meaning perpendicular to the centre line of the shaft and not directional.However, the term was confusing, and it is now more commonly referred to asa radial split.

When manufactured with a radial split, the casing can be made up in segments(especially for multistage pumps) and bolted together to form one pump. Thiseliminates costly construction and casting that would otherwise be involved inmaking a multistage pump.

The pump casings are made of many different materials, such as cast iron,steel or bronze, as well as many non-corrosive and abrasion-resistant alloys.

2.1.4. WEARING RINGS:

There is a small clearance between impeller hub & casing to reduce inter-stageleakage. This leakage point allows small amounts of water to recirculate in thepump. The leakage joint is generally provided with renewable wearing rings,which avoids costly build up of worn parts by welding. These are designed tobe adjustable so the proper clearance between the impeller and the casing canbe maintained. Wearing rings can be put either on impeller hub or on thecasing or on both. Generally when clearance double, the wearing ring are to be

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renewed. Fig. 2.7 explains different types of wearing ring & their mountingprocedure.

Fig. 2.7

Most manufacturers make the wearing rings of bronze or brass. They can beinstalled either in the casing or on the impeller, or both, as shown above.. Thetype of wearing ring and the method of installing it are determined by thepump manufacturers.

It should be remembered, however, that wearing rings are subjected to wearfrom gritty or abrasive materials and eventually have to be replaced. Thisreplacement may be necessary in six months, a year, or longer depending uponthe application and the material being pumped.

Wearing rings are not always included in the design of a pump. However,when they are included, the wearing rings are usually replaceable. Their mainfunction is to allow some fluid leakage between the impeller and the casing.By allowing a small amount of leakage between the impeller and the casing inthe suction area, a hydraulic seal is set up, helping the pump to operate moreefficiently. In addition, a small amount of liquid is recirculated from thedischarge to the suction side of the pump.

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Fig. 2.8

2.1.5. STUFFING BOX:

Stuffing boxes have the primary function of protecting the pump againstleakage at the point where the shaft passes out through the pump casing.Function of the stuffing box varies depending on whether the pump handlessuction lift or suction pressure. If the pump handles a suction lift and thepressure at the interior stuffing box end is below atmospheric, the stuffing boxfunction is to prevent air leakage into the pump. If this pressure is aboveatmospheric, the function is to prevent liquid leakage out of the pump.Normally gland packing is used. If When the pressure at interior of stuffingbox is negative, packing are used with lantern ring and water injection toprevent ingress of air into pump through stuffing box. Following are thevarious cases, when lantern ring & sealing liquid are used:

1. A suction in excess of 5meters.2. A discharge pressure under 7 meters.3. Hot water (over 1200C) being handled without adequate cooling.4. Muddy, sandy, or gritty water being handled.5. The liquid being handled is other than water such as acid, juice or

sticky liquids.

A packing gland seals the fluid flow in the pump. The packing gland may useeither a loose or a braided type packing that is replaceable and adjustable orthe mechanical seal.

2.1.6. BEARINGS:

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The function of bearings in pumps is to keep the shaft or rotor in correctalignment with stationary parts under the action of radial and transverse loads.

Shaft bearings can be of sleeve, single or double row ball bearing, or rollerbearing types. In addition, some pumps use special thrust bearings to eliminatedevelopment of high end thrust. Thrust bearings are normally used onmultistage pumps, which have high pressures and capacities. On smallhorsepower pumps, the impeller and shafts frequently overhang the motorbearings. In this case, all radial and axial thrust must be taken up by the motorbearings.

2.1.7. SUCTION PIPE WITH A FOOT VALVE AND A STRAINER:

A pipe whose one end is connected to the inlet of the pump and other end dipsinto water in a sump is known as suction pipe. A foot valve, which is a non-return valve or one way type of wave is fitted at the lower end of the suctionpipe. The foot valve opens only in the upward direction. A strainer is alsofitted at the lower end of the suction pipe.

2.1.8. DELIVERY PIPE:

A pipe whose one end is connected to the outlet of the pump and other enddelivers the water at a required height is known as delivery pipe.

Single stage, single suction type pumps have served for simple illustrations ofpump parts. Centrifugal pumps are also made in double suction and multistagedesigns. The type of pump used will be determined by the amount of fluid tobe pumped and the pressure required.

Multistage Pumps are available in either single or double suction. Whenmultistage pumps are used, water is discharged from one stage to the nextthrough internal passages in the pump casing. Each stage builds up thevelocity of the fluid until the desired head is reached.

Impellers discussed so far have been only for single impeller pumps. Inmultistage pumps, impeller design must be worked out carefully to avoidincreasing radial and thrust loads at each stage. To offset this, impellers arefrequently mounted back-to-back even though they are several inches apart.Some pump impellers are equipped with small vanes on the back of the shroudto pump any water, which gets behind the impeller. This action also helpscounteract the thrust from the normal impeller loading.

Many centrifugal pumps are made with end suction and have a vertical orhorizontal discharge point. Because smaller pumps are usually cast with asupporting base, the direction of flow is pre-determined by the pump design.When the pump casing is mounted independent of the supporting framework,the discharge point could be positioned in several different ways. Thedirection or location of the discharge point around the circle depends on thenumber of bolts that hold it in position on the frame.

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On double suction pumps the intake and discharge are normally on the samehorizontal plane, below the pump centerline, and on opposite sides of thepump. Some double suction pumps, as shown in Fig., are available withbottom suction and a side discharge port. Most of the double suction pumpshave horizontal split casings.

As was pointed out previously, the sealing mechanism may be either themechanical seal or the packing gland type. Selection depends on pumpconstruction and the material being handled, as well as the manufacturer'spreference.

On pumps handling hot fluids the packing gland is frequently provided withsome type of cooling. This can be cool water, gas, or other liquid, dependingon the material being pumped. In some applications, the pump casing and anysupporting bearings may also need to be cooled.

2.2. SINGLE and MULTI STAGE DESIGN:

A pump in which the head is developed by a single stage is called single stagepump. Often the total head to be developed requires the use of two or moreimpellers operating in series, each taking its suction from the discharge of thepreceding impeller. For this purpose two or more single stage pumps may beconnected in series or all impellers may be interposed in a single casing. Theunit is then called a multistage pump. If an electric motor operating at about3450 rpm is supposed to deliver 600ft of head at a capacity of 200gpm, forsuch pump specific speed will be around 400, indicating use of extreme typeof radial type of impeller. Such impeller will be difficult to design. If twostages are used, head per stage will be 300ft, and specific speed is 600, whichrepresents much less difficult design. For such reasons multi stage centrifugalpump are used.

The water from suction pipe enters the 1st impeller at inlet and is discharged atoutlet with increased pressure. The water with increased pressure from theoutlet of the 1st impeller is taken to the inlet of the 2nd impeller with the help ofa connecting pipe as shown in fig 2.8. At the outlet of the 2nd impeller, thepressure of water will be more than the pressure of water at the outlet of the 1st

impeller. Thus if more impellers are mounted on the same shaft, the pressureat the outlet will be increased further.

Let n = Number of identical impellers mounted on the same shaft,H = Head developed by each impeller.

Then total head developed

= n x H

The discharge passing through each impeller is same.

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Chapter-3

TERMS AND CHARACTERISTICS OF CENTRIFUGAL PUMP

3.1 Suction head and Suction lift

Suction head is the term used to describe a pump condition where the level ofthe source of supply is located above the centre line of the pump. Suction lift isthe term used to describe a situation where the level of the source of supply islocated below the centre line of the pump.

3.2 System head

In strict terms, a pump can only operate within a system. To deliver a givenvolume of liquid through that system, a pump must impart energy to theliquid, made up of the following components, which applies to all types ofpumps:

Static head. Friction head. Entrance and exit losses.

3.1.1. Static head

The static head refers to a difference in elevation. Thus, the total static headof a system is the difference in elevation between the liquid levels at thedischarge and the suction points of the pump (Fig.3.1). The static dischargehead is the difference in elevation between the discharge liquid level and thepump centerline. The static suction head is the difference in elevationbetween the suction liquid level and the pump centerline. If the static suctionhead has a negative value because the suction liquid level is below the pumpcenter line, it is usually spoken of as a "static suction lift". If either the suctionor discharge liquid level is under a pressure other than atmospheric, thispressure can be considered either as part of the static head or separately as anaddition to the static head.

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Fig- 3.1

Fig 3.1 c3.1.2. Friction head

The friction head is the head (expressed in feet of the liquid being pumped)that is necessary to overcome the friction losses caused by flow of liquidthrough piping, valves, fittings and any other elements such as heatexchangers. These losses very approximately as the square of the flow throughthe system. They also very with the size, type and surface condition of thepiping and fittings and with the character of the liquid pumped. The resultingcurve is called " system friction curve" as shown in Fig. 3.2.

Fig. 3.2

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In calculating friction losses, we must consider that they will increase as thepiping deteriorates with age. It is usual to base the losses on data establishedfor average piping that is 10 or 15 years old.

3.1.3. Entrance and exit losses:

If the supply of a sump originates in a reservoir, tank or in take chamber,losses occur at the point of connection of the suction piping to the source ofsupply. The magnitude of these losses depends on the design of the pipeentrance. A well-designed bell mouth provides the lowest possible loss.Similarly, on the discharge side of the system where the discharge lineterminates at some body of liquid, the velocity head of the liquid is entirelylost and must be considered as part of the total friction losses of the system.

3.2. System Friction and System Head Curves

As mentioned earlier, friction, entrance and exit losses vary approximately asthe square of the flow through a system. For solving pumping problems, it isconvenient to show the relationship, between capacity and friction head lossesgraphically. These losses are therefore calculated at some predetermined flow,either expected or design, and then calculated for all other flows, using thesquare of the flow relationship. The resulting curve is called the systemfriction curve as shown above.

When we combine the static heads, pressure differences and friction headlosses of any system and plot them against the capacity, the resulting curve(Fig 3.3) is called the system head curve. Superimposing a pump headcapacity curve at constant speed on this system head curve (Fig3.4) willpermit us to determine the capacity at the point where the two curves interact.This is the capacity that will be delivered into the system by that pump at thatparticular speed.

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Fig. 3.3 Fig. 3.4

For systems having varying static heads or pressure differences, it is possibleto construct curves corresponding to the minimum and maximum conditions,as shown in fig 3.5. The corresponding intersections with the pumps headcapacity curve will then determine the minimum and maximum flows that thepump will deliver into the system.

3.3. PUMP CHARACTERISTIC:

Characteristic curves of centrifugal pumps are defined those curves which areplotted from the results of a number of tests on the centrifugal pump. Thesecurves are necessary to predict the behavior and performance of the pumpwhen the pump is working under different flow rate, head and speed. Thefollowings are the important characteristic curves for pumps:

1. Main characteristic curves.2. Operating characteristic curves

3.3.1. MAIN CHARACTERISTIC CURVES:

The main characteristic curve of a centrifugal pump consists of variation ofhead (manometric head, Hm), power and discharge with respect to speed. Forplotting curves of manometric head versus speed, discharge is kept constant.For plotting curves of discharge versus speed, manometric head (H4) is keptconstant. And for plotting curves of power versus speed the manometric headand discharge are kept constant. Fig 3.5 shows main characteristic curves of apump.

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Fig. 3.5

3.3.2. OPERATING CHARACTERISTIC CURVES:

If the speed is kept constant, the variation of manometric head, power andefficiency with respect to discharge gives the operating characteristics of thepump. Fig 3.6 show the operating characteristic curves of a pump.

Fig.3.6

A centrifugal pump operating at certain speed can deliver any capacity fromzero to a maximum depending on pump size, design & suction condition. Theinterrelations of capacity, head, power & efficiency are called the pumpcharacteristics, Fig. shows ideal characteristic of various types of pumps.

Actual characteristic is different because of following losses:

Circulatory flow. Friction loss. Turbulence.

Pumps head-capacity curves can be classified as follows:

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1. Steady rising head characteristic- Rising headcapacity characteristic,meaning a curve in which the head rises continuously as the capacity isdecreased. (Fig 3.7a )

2. Steep rising head characteristic-A rising head-capacity characteristic inwhich there is a large increase in head between that developed at designcapacity and that developed at shutoff. (Fig 3.7b )

3. Drooping head characteristic- Drooping head characteristic, indicatingcases in which the head-capacity developed at shutoff is less than thatdeveloped at some other capacities. This is also known as a droopingcurve. (Fig 3.7c )

4. Flat head characteristic. A head capacity characteristic in which the headvaries only slightly with capacity from shutoff to design capacity. Thecharacteristic might also be either drooping or rising. All drooping curvehave a portion where the head developed is approximately constant for arange in capacity, called the flat portion of the curve. (Fig 3.7d )

Fig. 3. Fig. 3.

Fig. 3.7a Fig. 3.7b

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Fig. 3.7c Fig. 3.7d

3.4 SYSTEM HEAD CURVES:

To see the best suitability of pump it is necessary to identify system headcapacity curve. This is done most frequently for centrifugal pump. Fig 3.8shows system head curve which constitute friction head & total static headcomponents for a given capacity. Friction head is the friction losses in the pipeline, valves etc. friction head curve will change on variation in systemresistance e.g. throttling of valves etc.

Fig. 3.8

3.5 EVALUATION OF PUMP CHARACTERISTICS

Some process operations such as those requiring precise flow controls need ahead that varies markedly with a change in capacity. Pumps with rising headcharacteristic best fill these services. Other process, such as cooling watersupply system, need a head that remains relatively constant for a widevariation in capacity should have pumps with flat head characteristic. There isno clear-cut distinction between flat head & rising head characteristics.However as a rule of thumb, curves that shows a 150% increase in head fromthe capacities of peak efficiency & shut off are called steep rising curve;those showing a 10 to 25% rise are called steady rising curves & those with nomore than 10% increase are called flat.

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Pumps with drooping characteristic give two flows at certain head leading tounsuitability in regulation. Hence such pumps should not be operated in unsaferegion.

On comparison of actual H-Q and power curves of forward, radial andbackward vane impellers. We can see following advantages of backward vaneimpeller:

* Stable H-Q characteristic.* Limiting Power Characteristic.

Fig.3. 9 indicates characteristics of radial flow, mixed flow and axial flowpumps.

Fig. 3.9

3.6. CENTRIFUGAL PUMP CHARACTERISTIC RELATIONS

Certain relations exist that allow the performance of a centrifugal pump to bepredicted for a speed other than that for which the pump characteristic isknown. Similarly prediction of performance of a pump is possible for reduceddiameter of the impellers from the characteristics obtained at the largediameters

Q N Q D H N2 H D2

P N3 P D3

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Where N is speed in RPM and D is diameter.

3.7. SPECIFIC SPEED

The principle of dynamic similarity expresses the fact that two pumpsgeometrically similar to each other have similar performance characteristics. Inorder to have some basic comparison among various types of centrifugalmachines it became necessary to evolve a concept, which linked three mainfactors of performance characteristiccapacity, head, & speed into single term.The term specific speed is such a concept. Specific speed is the speed of ageometrically similarly pump, which delivers one unit capacity against one unitof total head. In its basic form, the specific speed is an index number, expressedas:

N QNS = ---------

H3/4

HereN = rpm of pump.Q = flow at optimum efficiency in gpm.H = head in ft. (head per stage for a multistage pump)

Specific speed for various pumps:

Radial flow pump : below 4000Mixed flow pump : 4000 to 9000Axial flow pump : above 9000

Specific speed is a useful tool in determining the NPSH required by a givenimpeller. Those that produce high total dynamic head have lower specificspeed and reverse is also true. At a particular head & capacity, a pump oflower specific speed will operate safely at lower NPSH than one with a higherspecific speed.

3.8. NET POSITIVE SUCTION HEAD

Adequate NPSH is essential for working with any pump whether centrifugal,rotary or reciprocating pump, to prevent cavitation. Cavitation occurs, whenabsolute pressure of liquid reaches equal to its vapour pressure- this causesformation of vapour bubbles in the fluid system, which are carried along untilpressure increases again. When these vapour bubbles collapse, void informeddemand immediate replacement of fluid. This result formation of pressurewave causing damage to impeller. Cavitation occurs at point such as the pumpimpeller entrance, trailing edge of valve gates other loss pressure points.

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Cavitation can remove considerable amount of metal form impeller surface orother affected surfaces. Cavitation does not mean immediate failure of hepump. But is should not be tolerated.

Net positive suction head can be expressed as follows:

NPSH, ft.= Ha Hv + Hs Hf

Where:

Ha : Pressure on fluid in suction tank, ft.Hv : Vapour pressure of fluid at pumping temperature ft.Hs : Static head or lift from fluid level to pump centre line for

suction head this term will be positive for suction lift this termwill be negative.

Hf : Friction loss for the designed flow rate at the pump suction, ft.

NPSH for existing installation can be measured as follows:

NPSH, ft = Pa + Ps + (Vs2/2g) Hv

Where :

Pa : Atmosphere pressure, ft.Ps : Gage pressure measured at pump suction flange

corrected to datum, ft.,V2/2g : velocity head, ft.Hv : Vapour pressure of fluid, ft.

All pumps require certain NPSH to permit the liquid flow in the pump casing.This is determined by pump designer, and influenced by speed of rotation, theinlet area, or eye area of impeller in a centrifugal pump, the type and numberof vanes in impeller, etc. on reciprocating pumps it largely function of speedand valve design. Available NPSH must be greater than required NPSH; effectof cavitation on centrifugal pump characteristic is shown in fig.3.10

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Fig. 3.10

3.9. CAVITATION

Cavitation is defined as the phenomenon of formation of vapour bubbles of aflowing liquid in a region where the pressure of the liquid falls below itsvapour pressure and the sudden collapsing of these vapour bubbles in a regionof higher pressure. When the vapour bubbles collapse, a very high pressure iscreated. The metallic surfaces, above which the liquid is flowing is subjectedto these high pressures, which cause pitting action on the surface. Thuscavities are formed on the metallic surface and also considerable noise andvibrations are produced.

Cavitation includes formation of vapour bubbles of the flowing liquid andcollapsing of the vapour bubbles. Formations of vapour bubbles of the flowingliquid take place only whenever the pressure in any region falls below vapourpressure. When the pressure of the flowing liquid is less than its vapourpressure, the liquid starts boiling and vapour bubbles are formed. Thesevapour bubbles are carried along with the flowing liquid to higher pressurezones where these vapours condense and bubble collapse. Due to suddencollapsing of the bubbles on the metallic surface, high pressure is producedand metallic surfaces are subjected to high local stresses. Thus the surfaces aredamaged.

PRECAUTION AGAINST CAVITATION:

The following precautions should be taken against cavitation:

i.) The pressure of the flowing liquid in any part of the hydraulic systemshould not be allowed to fall below its vapour pressure. If the flowingliquid is water, then the absolute pressure head should not be below 2.5mof water.

ii.) The special materials or coatings such as aluminum bronze and stainlesssteel, which are cavitation resistant materials, should be used.

EFFECTS OF CAVITATION:

The following are the effects of cavitation:i.) The metallic surfaces are damaged and cavities are formed on the

surfaces.ii.) Due to sudden collapse of vapour bubble, considerable noise and

vibrations are produced.

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3.10 PARALLEL OPERATION OF PUMPS

Adding flow of each pump at common head can draw combined characteristic.If the individual pump characteristic is identical, capacity shared by eachpump will be equal at any load. But if the pumps have different characteristiccapacity sharing will be unequal and at certain capacity the pump with lowershut off head may go out of line. Operating point is indicated at the locationwhere system characteristics is cut to combined pump characteristics. Forsmooth operation pumps should not have unstable characteristic. (Referfig.3.11)

Fig. 3.11

3.11 SERIES OPERATION OF PUMPS

For higher head requirement series operation of pump may be opted.Combined characteristic can be draw by adding head of each pump at commonflow. (Refer fig.3.12) Operating point is indicated at the location where systemcharacteristics is cut to combined pump characteristics.

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Fig.3.12

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Chapter-4

THRUSTS AND ITS BALANCING IN CENTRIFUGAL PUMPS

4.1 Radial Thrust In Centrifugal Pump

The internal water action in volute type casings tends to set up unbalancedradial forces as shown in fig.4.1. To reduce this unbalanced force, somepumps use a double volute casing as shown in Fig.4.2a & 4.2b. This doublevolute adds another guiding vane to the water. It also splits and balances theinternal radial forces of the liquid on the pump impeller shaft and bearings.

Fig. 4.1 Fig. 4.2a

Fig. 4.2b

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The double volute guide vanes also help to strengthen the casings internally,reducing some of the stress load on the casing walls. The double volute guidevanes can be supplied for both solid and split type casings without muchdanger of improper matching because of the casting and pouring techniquesdeveloped in recent years. Double volute guide vanes are also used onmultistage pumps to balance the discharge head and guide the fluid to thesuction side of the next stage.

In a single-volute pump casing design, uniform or near uniform pressures acton the impeller when the pump is operated at design capacity (which coincideswith the best efficiency) fig. 4.3. At other capacities, the pressures around theimpeller are not uniform, and there is a resultant radial reaction (F). Agraphical representation of the typical change in this force with pump capacityis shown in Fig.4.4 note that the force is greatest at shut-off head.

Fig. 4.3 Fig. 4.4

For any percentage of capacity, radial reaction is a function of total head, andof the width and diameter of the impeller. Thus, a high-head pump with alarge-diameter impeller will have a much greater radial reaction force atpartial capacities than a low-head pump with a small-diameter impeller. Azero radial reaction is not often realized; the minimum reaction occurs atdesign capacity. In a diffuser-type pump, which has the same tendency forover capacity unbalance as a single-volute pump, the reaction is limited to asmall arc repeated all around the impeller, with the individual forces cancelingeach other.

The application of the double-volute design principle to neutralize reactionforces at reduced capacity is illustrated in Fig.. Basically, this design consistsof two 180-deg volutes; a passage external to the second joins the two into acommon discharge. Although a pressure unbalance exists at partial capacity

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through each 180. deg arc, forces F 1 and F 2 are approximately equal andopposite, thereby producing little, if any, radial force on the shaft andbearings. Although the double-volute casing design principle has been knownfor some time, it was once a serious manufacturing problem, especially withhorizontal-discharge axially split double-suction pumps. This was because thedivision wall spanned the split (Fig. 4.5).

Fig. 4.5

The double-volute design has many "hidden" advantages. For example, inlarge- capacity medium- and high-head single- stage vertical pumpapplications, the rib forming the second volute and separating it from thedischarge waterway of the first volute strengthens the casing .

4.2 Axial Thrust In Centrifugal Pump:

The pressures generated by a centrifugal pump exert forces on both itsstationary and rotating parts. The design of these parts balances some of theseforces, but separate means may be required to counterbalance others.Axial hydraulic thrust is the summation of unbalanced impeller forces actingin the axial direction. As reliable large-capacity thrust bearings are nowreadily available, axial thrust in single-stage pumps remains a problem only inlarger units.Asymmetric pressure distribution on suction side & back shroud in singlesuction impeller causes axial thrust. Fig. 4.6 depicts pressure distribution.

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Fig.4.6

Methods of axial thrust balancing:

1. Use of double suction impeller.2. Balancing hole in impeller.3. Pump out vane in impeller.4. Hydraulic balancing.5. Staggering of stages in multi stage pump

4.2.1. Double Suction Impeller

Theoretically, a double-suction impeller is in hydraulic axial balance with thepressures on one side equal to and counterbalancing the pressures on theother (Fig.4.7). In practice, this balance may not be achieved for the followingreasons:

1. The suction passages to the two suction eyes may not provide equal oruniform flows to the two sides.

2. External conditions, such as an elbow being too close to the pump suctionnozzle, may cause unequal flows to the suction eyes.

3. The two sides of the discharge casing may not be symmetrical, or theimpeller may be located off-center. These conditions will alter the flowcharacteristics between the impeller shrouds and casing, causing unequalpressures on the shrouds.

4. Unequal leakage through the two leakage joints will tend to upset thebalance.

Combined, these factors create definite axial unbalance. To compensate forthis, all centrifugal pumps, even those with double-suction impellers,incorporate thrust bearings.

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Fig. 4.7The ordinary single-suction radial-flow impeller with the shaft passingthrough the impeller eye (Fig.4.8), is subjected to axial thrust because aportion of the front wall is exposed to suction pressure, thus exposingrelatively more back wall surface to discharge pressure. If the dischargechamber pressure were uniform over the entire impeller surface, the axialforce acting towards the suction would be equal to the product of the netpressure generated by the impeller and the unbalanced annular area.

Fig. 4.8

Actually, pressure on the two single-suction impeller walls is not uniform. Theliquid trapped between the impeller shrouds and casing walls is in rotation,and the pressure at the impeller periphery is appreciably higher than at theimpeller hub. Although we need not be concerned with the theoreticalcalculations for this pressure variation, Fig. describes it qualitatively.Generally speaking, axial thrust towards the impeller suction is about 20 to 30percent less than the product of the net pressure and the unbalanced area.

4.2.2. Balancing Holes

To eliminate the axial thrust of a single suction impeller, a pump can beprovided with both front and back wearing rings. To equalize thrust areas, theinner diameter of both rings is maintained in a chamber located on theimpeller side of the back-wearing ring by drilling so called balancing holesthrough the impeller. Leakage past the back-wearing ring is returned into thesuction area through these holes.

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Fig. 4.9

However, with large single stage single suction pumps balancing holes areconsidered undesirable because leakage back to the impeller suction opposesthe main flow, creating disturbances. In such pumps, a piped connection to thepump suction replaces the balancing holes.

4.2.3. Pump Out Vanes

Another way to eliminate or reduce axial thrust in single-suction impellers isby use of pump-out vanes on the back shroud. The effect of these vanes is toreduce the pressure acting on the back shroud of the impeller (Fig. 4.10). Thisdesign, however, is generally used only in pumps handling gritty liquids,where it keeps the clearance space between the impeller back shroud and thecasing free of foreign matter.

Fig. 4.10

4.2.4. Balancing Drum

The balancing drum is illustrated in Fig.. The balancing chamber at the backof the last-stage impeller is separated from the pump interior by a drum that iseither keyed or screwed to the shaft and therefore rotates with the shaft. Thedrum is separated by a small radial clearance from the stationary portion of thebalancing device, called the "balancing drum head," that is fixed to the pumpcasing.

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The balancing chamber is connected either to the pump suction or to the vesselfrom which the pump takes its suction. Thus, the back pressure in thebalancing chamber is only slightly: higher than the suction pressure, thedifference between the two being equal to the friction losses between thischamber and the point of return. The leakage between the drum head is, ofcourse, a function of the differential pressure across the drum and of theclearance area.The forces acting on the balancing drum in Fig.4.11 are the following:

1. Towards the discharge end-the discharge pressure multiplied by the frontbalancing area (area "B") of the drum.

2. Towards the suction end-the back pressure in the balancing chambermultiplied by the back balancing area (area "C") of the drum.

The first force is greater than the second, thereby counterbalancing the axialthrust exerted upon the single-suction impellers.

Actually, the drum diameter can be selected to balance axial thrust completelyor within 90 to 95 percent, depending on the desirability of carrying any thrustbearing loads.

Fig 4.11

It has been assumed in the preceding simplified description that the pressureacting on the impeller walls is constant over their entire surface and that theaxial thrust is equal to the product of the total net pressure generated and theunbalanced area. Actually, this pressure varies somewhat in the radialdirection because of the centrifugal force exerted upon the water by the outerimpeller shroud (see Fig.). Further more, the pressures at two correspondingpoints on the opposite impeller faces (D and E) may not be equal because of

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variation in clearance between the impeller wall and the casing sectionseparating successive stages. Finally, pressure distribution over the impellerwall surface may vary with head and capacity operating conditions.

This pressure distribution and design data can be determined by test quiteaccurately for anyone fixed operating condition, and an effective balancingdrum could be designed on the basis of the forces resulting from this pressuredistribution. Unfortunately, varying head and capacity conditions change thepressure distribution, and as the area of the balancing drum is necessarilyfixed, can destroy the equilibrium of the axial forces. The objection to this isnot primarily the amount of the thrust, but rather that the direction of the thrustcannot be predetermined because of the uncertainty about internal pressures.Still, it is advisable to predetermine normal thrust direction, as this caninfluence external mechanical thrust bearing design. Because 100 per centbalance is unattainable in practice and the slight but predictable unbalance canbe carried on a thrust bearing, the balancing drum is often designed to balanceonly 90 to 95 per cent of total impeller thrust.

Balancing drum modifications

To reduce the balancing drum leakage, a series of steps along the drum withsmall relief chambers at each step is sometimes provided. The drum surface isalso frequently serrated.

Experience indicates that the most successful simple balancing drum designsare of relatively long length. The length reduces the pressure drop per linearunit and thus decreases the rate of wear. This design, however, suffers fromthe fact that it substantially increases the pump shaft span.

Figure 4.12 illustrates a modification of the balancing drum that incorporates a"labyrinth" construction with concentric pressure-reducing passages. Theexample illustrated provides approximately 18 inches of effective drum lengthin only 7 1/2 in. of axial length.

The balancing drum satisfactorily balances the axial thrust of single-suctionimpellers and reduces pressure on the discharge side stuffing box. It lacks,however, the virtue of automatic compensation for any changes in axial thrustcaused by varying impeller reaction characteristics. In effect, if the axial thrustand balancing drum forces become unequal, the rotating element will tend tomove in the direction of the greater force. The thrust bearing must thenprevent excessive movement of the rotating element. The balancing drumperforms no restoring function until such time as the drum force again equalsthe axial thrust. This automatic compensation is the major feature thatdifferentiates the balancing disk from the balancing drum.

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Fig. 4.12

4.2.5. Balancing Disks:

The operation of the simple balancing disk is illustrated in Fig.4.13. The diskis fixed to and rotates with the shaft. It is separated from the balancing diskhead installed as a casing part, by a small axial clearance. The leakage throughthis clearance flows into the balancing chamber and from there either to thepump suction to the vessel from which the pump takes its suction. The back ofthe balancing disk is subject to the balancing chamber back pressure whereasthe disk face experiences a range of pressures. These vary from dischargepressure at its smallest diameter to back pressure at its periphery. The innerand outer disk diameters are chosen so that the difference between the totalforce acting on the disk face and that acting on its back will balance theimpeller axial thrust.

Fig. 4.13

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If the axial thrust of the impellers should exceed the thrust acting on the diskduring operation, the latter is moved towards the disk head, reducing the axialclearance is reduced so that the friction losses in the leakage return line arealso reduced, lowering the back pressure in the balancing chamber. Thisautomatically increases the pressure difference acting on the disk and moves itaway from the disk head, increasing the clearance. Now, the pressure buildsup in the balancing chamber, and the disk is again moved towards the diskhead until an equilibrium is reached.

To assure proper balancing disk operation, the change in back pressure in thebalancing chamber must be of an appreciable magnitude. Thus, with thebalancing disk wide open with respect to the disk head, the back pressure mustbe substantially higher than the suction pressure to give a resultant force thatrestores the normal disk position. This can be accomplished by introducing arestricting orifice in the leakage return line that increases back pressure whenleakage past the disk increases beyond normal. The disadvantage of thisarrangement is that the pressure on the stuffing box packing is variable acondition that is injurious to the life of the acting and therefore to be avoided.The higher pressure that can occur at the packing is also undesirable.

4.2.6. Combination disk and drum:

For the reasons just described, the simple balancing disk is seldom used. Thecombination balancing disk and balancing drum (Fig. 4.14) was developed toobviate the shortcomings of the disk while retaining the advantage ofautomatic compensation for axial thrust changes.

The rotating portion of this balancing, device consists of a long cylindricalbody (that turns within a drum portion of the disk head. This rotating partincorporates a disk similar to the one previously described. In this design,radial clearance remains constant regardless of disk position, where as theaxial clearance varies with the pump rotor position. The following forces acton this device:

1. Towards the discharge end-the sum of the discharge pressure multiplied byarea A, plus the average intermediate pressure multiplied by area B.

2. Towards the suction end-the back pressure multiplied by area C.

Whereas the "position-restoring" feature of the simple balancing disk requiredan undesirably wide variation of the back pressure, it is now possible todepend upon a variation of the intermediate pressure to achieve the sameeffect. Here is how it works. When the pump rotor moves towards thesuction end (to the left, in Fig. ) because of increased axial thrust, the axialclearance is reduced, and pressure builds up in the intermediate reliefchamber, increasing the average value of the intermediate pressure acting onarea B. In other words, with reduced leakage, the pressure drop across theradial clearance decreases, increasing the pressure drop across the axialclearance.

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Fig. 4.14

The increase in intermediate pressure forces the balancing disk towards thedischarge end until equilibrium is reached. Movement of the pump rotortowards the discharge end would have the opposite effect of increasing theaxial clearance and the leakage and decreasing the intermediate pressureacting on area B.

Figure 4.15 illustrates the pressure distribution in a combination balancingdisk and drum. No attempt is made to describe the exact manner in which thepressure decreases between any two points, although this curve is notnecessarily a straight line. Also, this illustration is not quantitatively correct. Itonly serves to show that changes in the balancing device position vary theinternal pressure distribution without altering the back pressure. The onlypossible variation may be caused by pressure changes at the point where thebalancing device leakage is returned to the system. An orifice may still belocated in the return line. Its function now, however, is not that of changingback pressure but rather of gagging the volume of leakage flow. This flowshould not be throttled outside the balancing device; the orifice pressure dropis negligible, ranging from about 2 to 20 psi.

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Fig.4.15

4.3 Axial thrust in multistage pumps:

It might seem that the advantages of balanced axial thrust and greater availablesuction area in a double-suction impeller would warrant applying suchimpellers to multistage pumps. But, there are definite shortcomings to thispractice. The average multistage pump has relatively low capacity whencompared to the entire range covered by modern centrifugal pumps. It isseldom necessary, therefore, to use double- suction impellers just to reduce thenet positive suction head required for a given capacity. Even if a doublesuction impeller is desirable for the first stage of a large capacity multistagepump, it is hardly necessary for the remaining stages. As to the advantage ofthe axial balance it provides, it must be considered that a certain amount ofaxial thrust is actually present in all centrifugal pumps and the necessity of athrust bearing is therefore, not eliminated.

Most important, the use of double suction impellers in a multistage pump addsneedless length to the pump shaft span. Additional space is required for theextra passage leading to the second inlet of each successive stage. In a pumpwith four or more stages (Fig.4.16), this increase becomes quite appreciableand causes additional casting difficulties. If shaft diameter is increased tocompensate for the longer span so as to maintain reasonable shaft deflection,the impeller inlet areas are correspondingly reduced. The result is that theadvantage of superior suction conditions usually offered by double-suctionimpellers is considerably reduced. Finally, as it is impractical to arrange thevarious double suction impellers in any but the ascending order of the stages,the impeller at one end of the casing becomes the last stage impeller and thepressure acting on the adjacent stuffing box becomes the discharge pressure ofthe next-to-last stage. To reduce this pressure, a pressure-reducing bushingmust be interposed between the last-stage impeller and the stuffing box, andthis bushing further increases the over-all length. The result of all theseconsiderations is that most multistage pumps are built with single-suctionimpellers.

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Fig.4.16

Axial thrust balancing in multistage pump are done as follows:

1. Several single-suction impellers may be mounted on one shaft, eachhaving its suction inlet facing in the same direction and its stagesfollowing one another in ascending order of pressure (Fig.). The axialthrust is then balanced by a hydraulic balancing device.

2. An even number of single-suction impellers can be mounted on one shaft,one half of these facing in an opposite direction to the second half. Withthis arrangement, axial thrust on the one half is compensated by the thrustin the opposite direction on the other half (Fig.). This mounting of single-suction impellers back-to-back is frequently called "opposed impellers."

An uneven number of single-suction impellers may be used with thisarrangement, provided the correct shaft and inter stage bushing diameters areused to give the effect of a hydraulic balancing device that will compensate forthe hydraulic thrust on one of the stages.

It is important to note that the opposed impeller arrangement completelybalances axial thrust only under the following conditions:

1. The pump must be provided with two stuffing boxes.2. The shaft must have a constant diameter.3. The impeller hubs must not extend through the inter stage portion of the

casing separating adjacent stages.

Except for some special pumps that have an internal and enclosed bearing atone end, and therefore only one stuffing box, most multistage pumps fulfil thefirst condition. But because of structural requirements, the last two conditions

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are not practical. A slight residual thrust is usually present in multistageopposed-impeller pumps, unless impeller hubs or wearing rings are located ondifferent diameters for various stages. Because such a construction wouldeliminate axial thrust only at the expense of reduced interchange ability andincreased manufacturing cost, this residual thrust, being relatively small, isusually carried on the thrust bearing.

4.3.1. Hydraulic Balancing Devices

A single-suction impeller is subject to axial hydraulic thrust caused by thepressure differential between its two faces. If all the single-suction impellersof a multistage pump face in the same direction, the total theoreticalhydraulic axial thrust acting towards the suction end of the pump will be thesum of the individual impeller thrusts. The thrust magnitude (in pounds) willbe approximately equal to the product of the net pump pressure (in pounds persquare inch) and the annular unbalanced area (in square inches). Actually theaxial thrust turns out to be about 70 to 80 per cent of this theoretical value.

Some form of hydraulic balancing device must be used to balance the axialthrust and to reduce the pressure on the stuffing box adjacent to the last-stageimpeller. This hydraulic balancing device may be a balancing drum, abalancing disk, or a combination of the two.

44

Chapter-5

OPERATIONS OF CENTRIFUGAL PUMPS

5.1 INITIAL START

Starting a centrifugal pump for the first time can be a trouble some experienceunless the plant crew has made a thorough check of the unit during & afterinstallation. Factors to be considered in starting any centrifugal pump includepipe cleanliness, pump alignment, rotation, lubrication, position of valves,stuffing box leakage, effect of speed changes, bypass quantities, throttling ofthe discharge & performance checks.

Factor to be checked in first start of any centrifugal pump are as follows:

1. Pipe cleanliness.2. Alignment and freeness of rotating assembly.3. Rotation: Proper direction of rotation.4. Pump bearings: should be inspected, cleaned & lubricated. Oil level

checked.5. Lube oil system: cleaned, components commissioned, cooling system

checked.6. Pump exterior: clean external surfaces of the pump and this driver. Check

tightness of all joints & foundation bolts. Ensure area around the pump setis clean.

7. Pump drive: Check the prime over that it lubricated and ready to operate.If possible run the drive independently of the pump to see it is in goodoperating order.

8. Priming: Prime the pump by any appropriate method. Take care to see thatsuction line is full of liquid and there is enough reserve in the supplysystem to keep the line full, while pump operates.

9. Position of valves: open the suction valve wide open. Position of dischargevalve is base on power characteristic of the pump and should be asfollows:

Radial flow pump: start with discharge valve closed since it takes lowpower at low flow, then open the valve, when pump attains full speed.

Mixed flow pump: Often requires greater power, when operated withclosed discharge, hence start with open discharge valve.

Axial flow pump: always take more power when started with dischargevalve closed. So that this pump with discharge valves open.

10. Pump warm up: with pump handling hot liquid, the casing, rotor, and otherparts must be brought to within a temperature of 0 to 100 0F of the liquidbefore unit is started.

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5.2 PIPE CLEANLINESS

Multistage pumps and many single-stage pumps have close clearance runningparts which must be protected from abrasive particles often found in newpiping systems.

5.3 ALIGNMENT:

Bring the pump to operating temperature by admitting liquid to the casing.Check the alignment as explained earlier. Turn the pump over by hand. Itshould turn freely, without binding, scraping or making any noise. Inspect thepump footings to see that any device for expansion of the casing are free & isgood working condition.

5.4 ROTATION

The pump must be run in the direction indicated by an arrow on the casing,which is always toward the discharge nozzle. You will note that the impellerrotates in the direction away from the vane curvature. Check the driver &pump rotation. Touch the motor starter just long enough to make the motorturn a few revolutions. The pump shaft should turn in the direction of thearrow on the casing, Fig a shows how to determine the direction of rotation ofa horizontal centrifugal pump. Stand at the driver end, facing the pump. If thetop of the shaft revolves from left to right, when viewed from this position, therotation is clockwise.

5.5 PUMP BEARINGS

Before any pump is started its bearings must be carefully inspected, cleaned &lubricated, with oil lubricated sleeve bearings, remove the cap, thrust shoes (ifused) & drain plug. Flush out the housing, oil piping, cooler pump and sumptank with kerosene, carbon tetrachloride or safety solvent. Wash the bearingparts thoroughly & reassemble them in the housing. Flush the entire systemwith lube oil and allow it to drain to waste. This will ensure removal of anydirt, metallic or waste particles present in the bearings or lube system. Replacethe drain plug, caps & other parts & fill the bearing as directed by themanufacturer.

Grease-lubricated ball, roller & needle bearings are usually packed with greaseat the factory before the pump is shipped. So no lubrication may be necessarybefore starting the pump. Check the condition of the greases by removing thebearing housing cover. See that there is grease in the bearing. An over greasedbearing may over heat soon after the pump is started.

Never start a pump equipped with kingsbury type thrust bearings without firstpouring oil into the bushing to protect the thrust shoes. Extreme care must be

46

exercised with all types of bearings on pumps to see that they have enoughclean lubricant.

When there is a forced feed lube oil system exists, fill the supply line to thebearings so it will be full when the pump start & here will be no delay in theoil reaching the bearings.

Bearing housings with oil both lubrication or with an oil sump which is to befilled to a given level, ordinarily are equipped with oil gauges or sight gauges.Oil is added when the oil level, due to loss, has dropped below the establishedlow limit. In general, the oil level should never reach higher than the centre ofthe lowest rolling element when the bearing is not rotating.

For regreasing, in a new or used anti friction bearing, remove the bearing fromthe shaft. Using a brush, wash the race, balls & other parts with kerosene orcarbon tetrachloride. Soak the bearing in one of these solvents until the greaseoff the bearing parts.

5.6 PUMP DRIVE

Check the motor, turbine, engine or other drive to see that it is lubricated &ready to operate. Whenever possible, run the drive independently of the pumpto see that it is in good operating order. See the manufacturers instructions fordrive operation.

5.7 PRIMING:

Prime the pump by one of the methods stated earlier. Take care that thesuction pipe is full of liquid & that there is enough reserve in the supplysystem to keep the line full while the pump operates. Reduced flow into thesuction pipe of a centrifugal pump can lead to overheating & extensivedamage to the pump.

5.8 POSITION OF THE VALVES :

Open the suction valve wide. Never use it as a throttling device for pump flow.With a medium or high head centrifugal pump it is best to start with thedischarge gate valve closed this is because the pump requires less power inputwhen primed & operated at full speed with the discharge valve closed.

Mixed flow type centrifugal pumps often require greater power input whenstarted with the discharge valve closed than when it is open. Axial flow typecentrifugal pumps almost always take more power when started with thedischarge valve closed. So it is a common practice to start these two typeswith the discharge gate valve open. But to the certain, check with the pumpmanufacture.

5.9 PUMP WARM UP

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With pumps handling hot water or other how liquids, the casing, rotor & otherparts must be brought to within a temp. of 500 to 1000 F of the liquid before theunit is started. This prevents unequal expansion, with the possibility of contactbetween the moving & stationary parts.

Open the vent valve on top of the casing & admit warm liquid to the pump.Use one or more casing drains to increase the liquid flow from the pump,thereby reducing the time required for warm up. In pumps like boiler feedpumps. This is not economical, instead, a jumper line around the dischargecheck valve is used. Hot liquid flows through this line, into the pump & outthe suction pipe.

5.10 STARTING THE PUMP

The following steps are usually suitable for starting a centrifugal pump ingood operating condition:

i. Turn on the cooling water system for the pump bearings, stuffingboxes & mechanical seals, if these parts are liquid cooled.

ii. Start the auxiliary lube oil pump, if one is fitted & check the oil flow tothe bearings & other parts of the pump.

iii. One the suction gate valve & close or open the discharge gate valve,depending on the starting procedure to be followed.

iv. Close all the drains in the casing & suction & discharge piping.

v. Prime the pump.

vi. Open the warm up valve if the pump is not at the right temp.

vii. Open the recirculating valve.

viii. Start the driver & bring the pump into speed.

ix. As soon as the pump is upto rated speed, open the discharge gate valveslowly.

x. Check the leakage from the stuffing boxes.

xi. Adjust sealing liquid flow to ensure packing lubrication. A flow of 1 to2 gpm. to each stuffing box is usually sufficient.

48

xii. Check the pump bearings for lube oil flow.

xiii. When there is sufficient flow through the pump, close the recirculatingvalve.

xiv. Check the pump suction discharge, lube oil, cooling water & sealingwater pressures & temps.

Bearing temp generally should not exceed 1500F during pump operation.

The above steps are suitable for almost centrifugal pumps. Some steps may beomitted with smaller units having separate cooling & oils systems. If thepumps show any signs of trouble while being started, overheated bearings orpacking, excessive vibration or noises stop the unit immediately. Inspect thepump for the cause of the trouble & take corrective action before starting thepump again.

5.11 OPERATING CHECKS:

While the pump runs make the hourly checks listed below bearing temp,suction & discharge pressure, lube oil temp & pressure, leak off flow,discharge flow meter, stuffing box leakage, cooling water suction & dischargetemp & pressure in put to pump driver & the oil level in the pump & driverbearings. Keep an hourly record of all these readings, using the logbook.

5.12 PUMP SHUTDOWN

When the liquid supply is above the pump centre line, close the discharge gatevalve, then the suction gate valve. Shutoff the power immediately. When theliquid supply is below the pump (suction lift), close the suction gate valve,then the discharge gate valve. Immediately after, shut off the power or steamto the driver. This procedure keeps liquid in the pump, preventing damageshould the unit be started before being primed. If the pump has a recirculatingline, close its valve before the pump stops. Open the warm up valve if thepump is to be kept warm.

When a pump operates against a high discharge pressure, it should be stoppedin three steps:

1. Partially close the discharge valve.2. Shut off the power.3. As soon as the power is off.

49

Close the discharge valve rapidly. The discharge valve should be tightlyclosed by the time pump stops rotating. Then there is no possibility of backflow causing the pump to turn in a reverse direction. This procedure alsoprevents water hammer in high-pressure lines.

If the pump is not to be started, hang a suitable tag on it. Do not shut-off thecooling & sealing liquid supply until the pump shaft has stopped turning.Remember that an idle pump will partially drain through the glands while notoperating. So be sure to prime it before starting again. Never run anycentrifugal pump dry. Serious damage will almost always result.

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Chapter-6

MAINTENANCE OF CENTRIFUGAL PUMPS

6.1 GENERAL

Centrifugal pumps whether horizontal or vertical can be considered to havetwo basic types of parts rotating and stationary. Rotating parts include theimpeller, shaft, wearing rings, shaft sleeves, bearings and mechanical seals.Stationary parts include the casing, bearing housing, suction and dischargeflanges, packing, leak off tubing, and base plate. Most maintenance work oncentrifugal pumps is concerned with the rotating parts, but some work is alsoperformed on the stationary parts.

6.1. ROTATING PARTS

6.1.1. IMPELLER

Immediately after removing the impeller from a pump inspect its eye, vanes,shrouds, wearing rings, passages, hub and other parts. Wear may occur at theeye, vanes, shrouds and other impeller parts. Corrosion, cavitation and erosionare generally accompanied by a wasting away of the impeller or vane surfaces.Where the attack is severe, the thinned section impeller may be sprayed with arubber, plastic or metallic coating., which is done by an organizationspecializing in it. If cavitation is severe, it may be necessary to change thesuction conditions or install an impeller suitable for the existing suctionconditions.

If the impeller is dirty when it is removed from the pump, clean it carefullybefore making an inspection. Use a soft wire brush or a steam lance to removethick gummy residues. Scale, coke and other deposits can be removed bychemical cleaning or sand blasting. In either case, precautions must be takento see that the impeller material is not damaged by the cleaning methodchosen. Pitting of the impeller may be caused by cavitation, which can occurwithout audible noise.

While it is possible to recondition an impeller that is worn or corroded, it isoften better to replace it with a new one suitable protected to resist wear orcorrosion.

Often, before the impeller can be removed for inspection, scale and burrs mustbe removed from the soft with a file. To prevent damage of the packing andfittings, clean the shaft thoroughly.

If a bronze impeller is a shrink fit on the shaft, as many are, slip a metal sleeveover shaft and against the impeller hub, to protect the shaft while the impelleris heated. Start heating the impeller with a torch from the outside of the shroudworking toward the hub. Revolve the impeller while heating it so its

51

temperature will be equalized. When the impeller is loose, pry it off the shaft,being careful to press only against the shroud. Wear asbestos gloves whenlifting the hot impeller.

IMPELLER RUNOUT

With pumps having bearings at each end of the shaft, mount the impellers,wearings rings, spacer and shafts sleeves on the shaft and support theassembly between centres (Fig.6.1). Set a dial gage at zero and take readingsnear each end and at the centre of each shaft sleeve. Also take similar readingsat each impeller-wearing ring. For most pumps, if the run out is not more than0.0015 in, the assembly can be considered accurate and the shaft installed asis. If the reading is greater, check for a bent shaft, out-of-square, dirty orburned impeller end of a shaft or spacer sleeve.

Fig. 6.1BALANCING

Badly worn or corroded impellers may vibrate excessively. While the presenceof vibration is usually easy o detect, a special balancing machine (Fig.6.2) isneeded to detect how much unbalance exists. It is usually necessary to returnthe impeller and shaft to the manufacturer for a check of this type.

To balance an impeller by hand, press it on an arbor, the ends of which rest ontwo parallel and level knife-edges. If out of balance, the impeller will turn andcome to rest with its heavy side down. To balance the impeller, metal must beremoved from the heavy side. This must be done without impairing the pumpperformance or accelerating erosion. For this reason, drilling holes in theheavy side is undesirable.

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The best practice is to mount a shrouded impeller off centre in a lathe and takea cut from the shroud, deepest at the rim. This may be done on one or bothshrouds, depending on their thickness and the amount of metal to be removed.The semi open impellers remove metal from the shroud if the design permits,or from the underside of the vanes of open impellers.

Fig. 6.2

6.1.2. SHAFTS:

Check for a bent shaft by means of a dial gauge. Badly bent shaft should bereturned to the pump manufacturer for straightening because the average plantdoes not have the necessary facilities. A shaft may also be checked for true-ness by swinging between lathe or other centre and checking the run out with adial gauge.

Tap the impeller shaft key to see that it is tight. Twist of the shaft under load,expansion, or corrosion will progressively loosen the impeller.

RECONDITIONING A SHAFT:

Centrifugal pump shafts wear while in use. Typical wear points are at thepacking box and other places where the friction load is high. Keep the frictionwear low by using a good grade of packing and adjusting the glands evenly.Be sure that the gland follower does not ride on the shaft. As soon as thepacking becomes dry, replace it.

It a new shaft is costly, or wears is rapid, it may pay to add a tougher wearresisting surfaces to the shaft at the points of sliding or rotating contact. Thisprocess is known as hard surfacing and can increase the life of some partsfrom four to thirty times.

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Center the shaft in a lathe supporting it in a steady rest located near the pointof work, if necessary. Use calipers to measure any unevenness to find thedepth of cut needed to give the necessary thickness of metal overlay.

When the build up is done by welding keep the welding deposits low and laythe bead longitudinally along the shaft. For two inch and smaller shafts,alternate the beads at 900 intervals until the shaft is completely covered at itsworn section.

Before spraying a shaft with metal, take a rough thread cut of 24 threads perinch. Then break the apex of the thread with a knurling tool. Mount the gun onthe tool post with the nozzle about 6 in from the work. Hold the gun by handwhen spraying the undercut or dovetail section at each end of the machinedarea.

Take a rough finishing cut, leaving enough metal for filing and lapping of theshaft to its final size. File lightly moving the file back and forth as the shaftrevolves. When the shaft, is correctly filed, polished it with No. 0 and 00emery cloth. The shaft is now ready to be replaced in the pump. There areother methods of hard-surfacing shafts. For best results, consult themanufacture of the equipment for this work.

6.1.3. WEARING RINGS:

Wearing ring clearance is of extreme importance because, as the clearanceincreases in a given pump leakage of liquid past the rings becomes greater,reducing the efficiency of the pump. For wearing rings typical none allcombinations are:

1. Bronze with a dissimilar bronze.2. Cast iron with bronze.3. Steel with bronze4. Monel metal with bronze and5. Cast iron with cast iron.

There is less leakage after accidental contact between casing and impeller-wearing rings made of these materials. Figure 2.8 shows typical wearing ringclearance, recommended by one manufacturer. One method of measuringwearing-ring clearance is shown in Fig 6.3.

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.

Fig. 6.3

WEARING RING INSTALLATION:

Centrifugal pumps fitted with wearing rings come supplied with the rings. Soit is not necessary to install rings on a new pump. Once the rings wear, theymust be replaced. To do this, first secure suitable replacement for the rings inthe pump from the manufacturer. Remove worn impeller rings which arethreaded or shrunk in place, by heating the ring with a torch, be careful not toheat the impeller. Or insert a few pieces of dry ice in the impeller eye to shrinkthe impeller away from the ring. The casing wearing rings to be removedusing similar methods.

Since many impeller rings are shrink fits, heat the ring before slipping intoplace and pinning. For ring diameters of 2.5 to 6in., interference between thering and the impeller is 0.001 to 0.0015 in. between 6 and 12 in., interferenceis 0.002 and 0.0025 in. insert the pin after the ring is in place. Figure showsthe methods and tools used in taking measurements of the impeller and casingwearing rings.

To restore the clearance in a pump having a single wearing ring, obtain a newring bored undersize and true up the impeller hub by turning it down in a lathe.Sometimes it may be possible to build up either the casing ring or the impellerhub, machining both to give the correct clearance. This is difficult and feasibleonly if the pump is large and the equipment to do the work is available.

Pumps with double wearing rings can have their clearance renewed in threeways:

1. Obtain a new oversized impeller ring and use the old casing ring, boredtrue to the larger diameter.

2. Obtain a new casing ring bored undersize and use the old impeller ringturned to a smaller diameter.

3. Renew both rings, if necessary.

When a new wearing ring is put on an impeller, its surface is often off-centrewith the shaft. So, after mounting a new ring, check its wearing surface.Machine, if necessary. Do this whether the ring is pressed, bolted, or screwedon the impeller.

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Measure the clearance of flat wearing rings with a feeler gauge between thestationary and rotating parts. In multistage pumps, where the wearing ring maybe L-shaped, the lip of the L prevents using the gauge. A fairly close check ofthe clearance may be made by mounting a dial indicator on the impeller andsetting it to zero with the casing ring resting on the impeller wearing ring hub.Without moving the impeller or dial indicator, push up on the stationary ringfrom below and record the maximum dial reading. This is the diametralclearance. Divide by 2 to get radial clearance, tolerances are always plus forcasing wearing rings, minus for impeller wearing ring.

6.1.4. SHAFT SLEEVES:

These wear when packed too tightly. They may be reconditioned by weldingor metallizing in the same manner as described above for shafts. Where wearis extreme, replacement of the worn sleeve with a new one is oftenrecommended. Use a sleeve puller to remove old sleeves from the shaft. Whenthe sleeve is rusted to the shaft, use the impeller nut to help loosen the sleeve.In extreme cases a hammer and chisel may be needed to split the sleeve beforeremoval. Be careful not to damage the shaft with the hammer or chisel. Afterinstalling a new shaft sleeve, check its concentricity on the shaft.

6.1.5. BEARINGS:

Feed pump and the ring oiled sleeve bearing and pressure lubricatedKingsbury thrust bearing opened for inspection and maintenance. Check thejournal and thrust bearings as shown in fig. Ball bearings may be checked asshown in fig.

Clearance between the shaft and the babbit of sleeve bearings should notexceed 150% of the original value.

SLEEVE BEARINGS:

These can be re-babbitted in the field, if desired, but this practice is generallyconfined to larger plants. To replace the metal in babbit-lined bearings:

1. Machine or melt out the old metal and place it with new metal in a meltingpot.

2. Wash the shell in a week solution of lime and water. Rinse in cleanrunning water. Do not touch or wipe the area to be metalled. Dry withcompressed air.

3. Dip the shell in a bath of hydrochloric acid (1 part acid to 5 parts water)for 10 min, or paint with this solution, using an old but clean brush. Rinsein water.

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4. Plug oil holes with heat resistant plastic or other material (fire clay, etc)coat all the surfaces not to be tinned with a fire-clay wash (made up like acement wash).

5. Swab the surface to be tinned with a flux. This can be made by dissolvingzinc in nitric acid until no more will dissolve and the acid is neutralized orkilled.

6. Place the bearing shell in a molten solder bath. Keep it there until it is ashot as the solder 5750 to 625 0F remove and inspect the tinning. Any barepatches must be scraped bright, fluxed and retinned.

7. Place the bearing shell in a previously prepared cylindrical jig of suitablesize and centre an undersized length of shaft in the shell.

8. Pour the babbit at 8000 to 8500 F into the space between the jig and amandrel placed inside it. The diameter of the mandrel should be slightlyless than that of the pump shaft. Use a ladle enough to complete thepouring in one filling.

9. Puddle the molten babbit with an iron wire, working the wire up anddown. This prevents cavities forming during cooling.

10. If the metal shrinks from the edges of the bearing shell, correct by penningit outward. Clean all the hammer marks when fitting and scraping thebearings.

Set up large bearings in a lathe chuck and bore to the correct size after thebabbitt is cool. Bolt both halves of the bearing together before the boring isbegun. Make the final fit by scraping and checking. Small bearings can bebored in a drill press. Use an adjustable double-edge cutting tool. Take finecuts to prevent the tool from digging into the metal scrape to size.

ROLLING CONTACT BEARINGS:

Keep all rolling contact bearings clean at all times dirt means damages. Workon the bearings with clean tools in clean surroundings. Remove all outside dirtfrom the housing before exposing the bearing. Handle bearings only withclean dry hands. Treat a used bearing as carefully as a new one. Use cleansolvents and flushing oils. Lay the bearings on only clean paper. Projectdisassembled bearings from dirt; use clean lint free rags if the bearings arewiped. Keep the bearings in oil proof replacing a bearing. Install new bearingas removed from their packages, without washing. Keep bearing lubricantsclean when applying and cover containers when not in use.

An arbor press is usually the best tool for removing a bearing. But most fieldwork is done with some from of bearing putter. To remove a bearing, press orpull only on the ring which is tight. Press or pull straight and square to keep

57

the ring from cocking. Never press or pull against the bearing shields orseparators.

6.1.6. MECHANICAL SEALS:

Remove the seal cover containing the stationary face and pull the face fromthe cover by hand. Using two wire hooks, engage them in the holes providedand remove the rotating face. Although normally only the faces require repair,it is a good idea to remove the spring assembly for cleaning and inspection.Use the same wire hooks as before.

If the stationary face is scarred, it probably has dirt or scale embedded in it.Take a light clean-up cut to remove the metal to below any possible dirt. If theface is made of carbon, machine it from the outside, working inward, toprevent chipping. It is best to use a carbide-tipped tool with a 50 lead angle and70rake. Zero rake can be used for bronze faces.

If the rotating face is only slightly scored, lap it without machining. But if thescoring is deep, take a fine cut in a lathe. If the face is stellite or hardenedchrome steel of 500 brinell or softer, machine it with a carbide-tipped toolhaving zero rake and 70 lead angle. If possible, hold it in a four jaw chuck toavoid distortion.

Next, the faces must be lapped. If good lapping plates are not available in theplat, it may be possible to purchase them from the pump or seal manufacturer.The plate should be made of soft close-grained cast iron and properly chargedwith diamond powder in paste form. Loose lapping compound should not beused for carbon or bronze faces. Both are porous and the loose compound fillsthe pores, and in effect makes a lap of the sea face. Washing the face is notalways effective.

Use two lapping plates, one charged with 750-mesh powder for roughing andone with 1,600 mesh for finishing. Apply diamond paste from its tube, and rollit into the plate surface with a hard roller. The plate is ready for lapping whenthe diamond particles are embedded in the iron with the points protruding.

Keep the plate wet with solvent at all times to the wash the residue into theplate grooves. Use a figure-sight stroke, covering as much of the plate aspossible to obtain a flat face. When finished, a lapped carbon face shouldhave a high polish. Do not leave loose particles of carbon on the face, they areabrasive and will cause wear.

The rotating face of a seal can be lapped using the same technique, but with alight pressure, to avoid damaging the plate. Do not use loose compound ondiamond-charged laps, if an optical flat is available, check the faces forflatness. Remember, it is impossible to obtain a flat seal face using a lappingplate that is not flat. When reinstalling the seal, be sure that all parts arethoroughly clean, use new gaskets.

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Routine Maintenance may be classed as work done primarily to rectify theefforts of normal wear in the pump. Over haul and repair operations areperformed to rectify the results of excessive wear, over heating, damage fromsolids in the liquid, or injury or wear from any other cause.

6.2 STATIONARY PARTS

6.2.1 CASING

When the pump casing is open, examine the waterways for corrosion anderosion. If the casing is pitted or eroded, this can often be corrected bywelding brazing, or metal spraying, depending on the material from which thecasing is made and the facilities available for repair. If the signs of wear aresevered, consult the manufacturers on the possibility of using more resistantmaterials. Clean and paint the waterways before closing the pump casing.

Most split casing pumps have one or more gaskets between the upper andlower halves. These should be replaced whenever the casing is opened forinspection or repairs. To prevent delays during pump reassembly, be sure tohave a new gasket on hand before the casing is opened. New gaskets should bethe same thickness as the original. Trim the inner edges of the gasketaccurately along the inside of the pump water ways. Have the gasket mountingsurfaces absolutely clean before applying the gasket.

Since paper and asbestos gaskets may dry out while a pump is stored or idle,check and tighten, if necessary, the flange bolts before starting the pump.Since the gaskets may be further compressed by differential expansion duringpump operation, check

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