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Centrifugal Pumps:Overview of Design, Operation and

Malfunctions

By

D. Craig Sever

And

Charles T. Hatch

Bently Nevada Training Development GroupBently Nevada Corporation

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Copyright (C) 1999 Bently Nevada Corporation. All rights reserved.

The information contained in this document is subject to change without notice.

The following are trademarks of Bently Nevada Corporation in the United Statesand other countries:

Actionable Information, Actionable Information to the Right People at the RightTime, ADRE, Bently Align, Bently Balance, Bently Nevada, CableLoc,

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Move Information, Not Data, NSV, Preformalign, Process CenteredMaintenance, PROXPAC, Proximitor, REBAM, Seismoprobe,

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The Bently Nevada Corporation Orbit Design, Bently Balance and Design, System 1Enabled and Design, and M2 and Design are all trademarks or registered marks of

Bently Nevada Corporation in the United States and other countries.

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Table of Contents

1. INTRODUCTION TO CENTRIFUGAL PUMPS................................................................................. 1

THE ROLE OF PUMPS AND THE CONSEQUENCES OF PUMP MALFUNCTION ................................................... 1WHAT IS A PUMP? ....................................................................................................................................... 2TYPES OF PUMPS.......................................................................................................................................... 3

2. DESIGN AND OPERATION OF CENTRIFUGAL PUMPS............................................................... 6

TERMINOLOGY OF ENERGY IN LIQUIDS ....................................................................................................... 6THREE FUNDAMENTAL WAYS CENTRIFUGAL PUMPS ADD ENERGY............................................................ 7PUMP COMPONENTS AND THEIR PURPOSES ................................................................................................. 7

Impeller................................................................................................................................................... 8Casing ................................................................................................................................................... 13Multiple Stages ..................................................................................................................................... 16Inlet Geometry ...................................................................................................................................... 16Seals...................................................................................................................................................... 17Sealless Pump Designs ......................................................................................................................... 19Wear Rings ........................................................................................................................................... 20Shaft Sleeves......................................................................................................................................... 21Thrust Balancing................................................................................................................................... 22Bearings ................................................................................................................................................ 23Couplings .............................................................................................................................................. 24

PERFORMANCE, OPERATION, AND TERMINOLOGY .................................................................................... 25Pump Performance Curves: Important Pump Parameters..................................................................... 25System Curves ...................................................................................................................................... 26Pump Operation: How Pump and System Curves Relate ..................................................................... 27The Best Efficiency Point ..................................................................................................................... 27Specific Speed....................................................................................................................................... 28Net Positive Suction Head and Suction Specific Speed........................................................................ 29

3. MALFUNCTIONS OF CENTRIFUGAL PUMPS .............................................................................. 32

GENERAL CONCEPTS ................................................................................................................................. 32PUMP MALFUNCTIONS............................................................................................................................... 33

High 1X Vibration due to Unbalance.................................................................................................... 33Radial Loads (Misalignment and Sideload) .......................................................................................... 38Rub........................................................................................................................................................ 47Shaft Crack ........................................................................................................................................... 53Fluid-Induced Instability....................................................................................................................... 62Structural Resonances........................................................................................................................... 70Cavitation.............................................................................................................................................. 72Vane Pass Frequencies.......................................................................................................................... 77

4. REFERENCES........................................................................................................................................ 78

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1. INTRODUCTION TO CENTRIFUGAL PUMPS

THE ROLE OF PUMPS AND THE CONSEQUENCES OF PUMP MALFUNCTION

Vast numbers of processes require liquid to move from one location to another.These processes can be seen in nuclear and non-nuclear power generation, oil pipelines,petrochemical refineries, municipal wastewater and domestic water treatment facilities,both large and small buildings, on ships and offshore oil platforms, and manufacturingplants, and the list could go on. In virtually all of these processes, pumps play theessential role of providing the propulsion necessary to move the liquid.

Pumps are a generally robust and reliable class of rotating machinery. However,pumps are critical machines in many processes because their loss can create serious oreven catastrophic results. Power generation relies on boiler feedpumps, condensatepumps, and water circulation pumps to circulate water through the thermodynamicprocess that converts fuel into electrical power. Nuclear power generation would beimpossible without the variety of pumps to circulate water through the primary reactorcore loop, secondary power generating loop, and cooling water loop. Power generationpumps are typically large and custom, one-of-a-kind design. The failure of a powergenpump can result in significant financial loss due to pump damage, as well as damage toassociated equipment. For example, a large high-pressure boiler evaporating about amillion pounds of water per hour could suffer extensive damage within minutes ifallowed to run dry due to a failed boiler feedpump.

Process industries such as petrochemical refineries are also vulnerable to similarfinancial consequences. The processing of liquid product employs large numbers ofpumps. A failed pump can shut down an entire process resulting in revenue losses on theorder of tens or even hundreds of thousands of dollars a day. In order to avoid suchlosses, many process industries find it necessary to devote large portions of theirmaintenance budgets to pumps.

Safety is an even greater concern than the financial impact of pump malfunctions.Public and plant personnel can be seriously endangered by accidents stemming frompump failures in processes that handle radioactive or toxic liquids. For instance,operating conditions can affect the reliability of pump seals. If a malfunction causesvibration, temperature, or pressure to change radically or to move outside of normaloperating ranges, these seals may leak and expose plant workers and the surroundingpublic to the adverse effects of hazardous liquids.

Environmental damage due to pump failure can also be a very serious problem.Hazardous materials released into the environment through leaking pump seals can havesignificant environmental impacts. The consequences of such unintended releases are notlimited to environmental damage, there may be heavy financial costs as well.Environmental regulations governing hazardous materials have become very stringentand environmental regulatory agencies may require the filing of a report, impose largefines, shut a plant down, or all of these, depending on the nature and amount of liquidreleased.

All of these factors combine to make pumps a class of rotating machinery thatdeserve in-depth examination.

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WHAT IS A PUMP?Almost everyone is familiar with pumps and their basic function. We readily

recognize that the water pump in an automobile circulates engine coolant through theradiator and water jacket. However, it is helpful to establish the function of pumps inmore precise terms. Pumps can be compared to the engine of a car. It is well known thatan automobile engine accelerates the mass of the car against the effects of inertia,overcomes frictional resistance caused by air, tires, etc., and overcomes the gravitationalresistance of moving a car up a hill (i.e., elevation changes). Driving against a strongheadwind or up a steep grade gives one a special appreciation for these effects. Likewise,liquid in a pipe must be accelerated against the inertia of its mass and once accelerated toa desired velocity (or flow rate), energy must be added on a continual basis to keep theliquid flowing against frictional resistance and elevation changes. In actual practice, theinertia of flowing liquids are largely ignored because it is of less concern than the othertwo forms of resistance.

The idea of pumping against an elevation change is not hard to imagine. As with acar, it is simply the resistance encountered when moving a liquid uphill against earth’sgravitational pull. However, not all may be as familiar with the concept of frictionalresistance to liquid flow. Just as friction occurs between two blocks of wood that arerubbed together, friction also occurs between individual molecules of liquid that “rub”together while flowing down a pipe. The molecules of liquid rub because they are not allmoving at the same velocity. Liquid molecules immediately adjacent to a pipe surfacehave zero velocity while molecules in the center of the pipe have maximum velocity.This can be seen by observing flow in a river where the flow is slowest at the edges andbecomes swifter toward the center. It follows then that there must be a variation in speed,or gradient, between the molecules closest to the stationary pipe surface and those in thecenter of the flow. This gradient means that adjacent liquid molecules have slightlydifferent speeds causing them to rub against each other and produce friction. Thisfriction combined with gravity creates significant resistance that a pump must overcomeif a liquid is to flow.

A closer examination of the concept of friction in liquids allows us to recognize thatthe magnitude of pipe friction loss depends on several factors. Rougher pipe surfacescreate more drag on a liquid than smoother surfaces and hence more friction. Smallerdiameter pipes have less cross-sectional flow area than larger pipes which yields greaterresistance to flow. In addition, certain properties of the liquid itself are contributingfactors. Emptying a can of motor oil versus a glass of water illustrates how the higherviscosity oil molecules cling to each other more than the water molecules. This “cling” islargely due to the cohesion between molecules. The greater the cohesion, the greater theamount of energy required to make a liquid flow. This translates directly into greaterfrictional resistance for higher viscosity liquids. The actual flow of liquids in pipes isquite complex and these are just a few of the factors that affect frictional resistance.However, this simple explanation gives us sufficient understanding of the task that pumpsmust perform. Just as a car engine provides energy to keep the car moving againstfriction and gravity, so too a pump provides energy to keep a liquid moving againstelevation and frictional resistance.

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Energy addedby Pump toovercomeFriction andElevation

Energy lostto Friction

Energy lostto Friction

Graph of Energy Change in Liquid Flowingthrough Piping System and Pump

Pump supplies an abruptincrease in Energy

Energy Grade Line

Figure 1.1 Energy Grade Line (EGL) shows how the energy of a liquidchanges as it flows through a piping system, pump, and change in elevation.Energy losses are represented by a decreasing EGL while energy gains arerepresented by an increasing EGL. Energy is required to make a liquid flowagainst the effects of frictional and gravitional resistance, the purpose of apump is to provide the energy necessary to overcome these resistances.

The relationshipbetween energy lost toflow resistance andthe energy gainedfrom a pump can beshown graphically in adiagram called anenergy gradelinediagram (Figure 1.1).

Of course, pumpsdo not accomplish thetask of supplyingenergy by themselves.Pumps are actuallyenergy converters;they take rotative shaftenergy from a driverand convert it toincreased energy inthe pumped liquid.The goal is to pump asefficiently and costeffectively as possible,using the minimalamount of mechanicalenergy per unit ofenergy added.

TYPES OF PUMPS

Pumps fall into two broad categories depending on how they add energy to thepumped liquid. The first category is known as displacement pumps and these utilizeplungers, pistons, diaphragms, screws, gears, or other similar means, to exert a forcedirectly on the liquid. Except for screw and gear type pumps, displacement pumps use acyclical process that imparts the energy in pulses. The second category is referred to askinetic pumps because they add energy by passing the liquid through an impeller which“speeds up” the liquid thereby increasing its kinetic energy. In other words, kineticpumps do not push on the liquid quite as directly as displacement pumps. They addenergy using a different principle that will be explained in more detail in followingsections. In contrast to displacement pumps, kinetic pumps add energy in a smoother andmore continuous process.

Kinetic pumps are sometimes referred to as centrifugal pumps. Most of the pumpsused in power generation and process industries are centrifugal pumps. While allcentrifugal pumps use an impeller to add kinetic energy, there are many different designsdepending on the specific application. Centrifugal pumps may be single-stage (one

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impeller) or multi-stage (more than one impeller) and may rotate about a horizontal orvertical axis. Single-stage centrifugal pumps may have their impeller over-hung(supported at one end only) or have their impeller supported on both ends betweenbearings. In addition to the impeller, the design of the pump case also varies widelydepending on application. These are just a few of the design differences amongcentrifugal pumps.

The pumps shown in Figure 1.2 and 1.3 illustrate a typical single-stage end-suctionvolute pump. It is only one of many variations among centrifugal pump designs, albeit acommon one. Figure 1.2 illustrates most of the components that are common to all(radial and mixed flow) centrifugal pumps.

Casing

ImpellerCoupling

Suction

Discharge

Bearings (2)

ShaftMeanFlow Line

Wear Rings (4)

Seal LubricationPort

ShaftSleeve

Sealing Area

SuctionEye

Figure 1.2 Cross-section of a typical end-suction centrifugal pump with single-suction, over-hungimpeller. Section on left is taken through pump shaft, section on right is taken through impeller and volutealong mean flow line.

Figure 1.3 Centrifugal pump used to pump water. Suction enters from left in both photos, dischargesthrough top. Whereas the pump illustrated in Figure 2 is supported by its own mounting feet, the pumpshown above mounts directly to the driver (electric motor) housing with no additional support.

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This paper focuses on centrifugal pumps because they are the most widely used pumpdesign in the world. In addition, centrifugal pumps represent a significant portion of thecapital, operating and maintenance costs of the power generation and process industries.

This paper will discuss the root causes underlying some of the more commoncentrifugal pump malfunctions, how to recognize their characteristic symptoms and howto correct them. In order to understand pump malfunctions, it is first necessary tounderstand how they are designed and operated.

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2. DESIGN AND OPERATION OF CENTRIFUGAL PUMPS

TERMINOLOGY OF ENERGY IN LIQUIDS

We have established that the fundamental purpose of pumps is to add energy to liquidso that it can flow against the effects of frictional and gravitational resistance. Beforecontinuing with the explanation of how pumps accomplish this task, it is first necessaryto understand the terms used to commonly describe this type of energy. Those who workwith pumps refer to the energy added by a pump as head (H). Head is measured in unitsof feet or meters. Head can take on three forms with each form being measured by adifferent means. The first form is static pressure head, or simply pressure head (HP).Pressure head is the energy measured with a pressure gauge. The second form iselevation head (HE). Elevation head is the potential energy that a liquid has by virtue ofits relative vertical position in a system. Thus, the higher a liquid is, the greater itselevation head. The third form is velocity head (HV). Velocity head is the kinetic energyof a liquid due to its velocity. Velocity head is commonly measured with a pitot tube.The total energy in a liquid consists of the sum of these three forms of energy.

The total energy, or head, of a liquid can be distributed in any proportion among thethree forms. The total energy may exist completely in one form to the exclusion of theother two, or it may exist as 30% pressure head, 30% elevation head, and 40% velocityhead, or it may exist in any other combination as long as the sum of the three formsequals 100% of the total head. For example, the water at the bottom of a swimming poolwill have no elevation head (compared to water at the pool’s surface) and will have novelocity head (assuming there is no circulation in the pool). However, it will have energyin the form of pressure head and this is exactly the pressure felt on one’s ears whilediving to the bottom of a deep swimming pool. Conversely, water at the top of the poolwill have potential energy because of its elevation head but it will have no pressure head.The lack of pressure head is sensed by the absence of pressure on one’s ears immediatelybelow the water surface. Water situated at levels between the top and bottom of the poolwill have some combination of pressure and elevation head depending on depth.

We can convert the pressure head to velocity head by opening an imaginary valvelocated at the bottom of the pool. Water will flow through the valve and we canintuitively understand that higher pressure in the pool will correlate directly with a highervelocity through the valve.

If we calculate the total head in the high elevation water at the top of the pool, and thetotal head in the high pressure water at the bottom of the pool, and the total head in thehigh velocity water flowing out of the valve it will be the same in all three cases. Thefact that the total head converts among its three forms without increasing or decreasing(assuming no energy losses or gains by external means) is known as Bernoulli’s law.Bernoulli’s law is an expression of the fundamental principle of the conservation ofenergy.

We must understand that head can exist in one of three forms and that it canconvert between them because centrifugal pumps operate by first adding velocityhead and then converting some portion of it into pressure head.

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THREE FUNDAMENTAL WAYS CENTRIFUGAL PUMPS ADD ENERGY

All centrifugal pumps use an impeller to add velocity head to a liquid. However, notall impellers accomplish this in exactly the same manner. Depending on the methodused, impeller designs are grouped into three general types. The difference between themis the direction in which each type forces the high velocity liquid to flow.

1. Radial flow impellers increase liquid velocity in a direction perpendicular (orradial) to the pump axis (Figure 2.1a).

2. Mixed flow impellers increase liquid velocity in a direction that is a mixture ofperpendicular and parallel flow with respect to the pump axis (Figure 2.1b).

3. Axial flow impellers increase liquid velocity in a direction parallel (or axial) tothe pump axis (Figure 2.1c).

Each type of impeller provides a certain combination of performance features.Hence, each type is best suited to meet the needs of particular applications.

PUMP COMPONENTS AND THEIR PURPOSES

Centrifugal pump designs range from small and simple to large and intricate.However, no matter how complex or simple the overall machine, there are parts commonto all designs that provide the same function.

The following is a discussion of these common pump components. It is intended thatthis will provide a context for the latter discussion of pump malfunctions.

Impeller

High VelocityOut

LowVelocity

In

Figure 2.1a Radial Flow Impellerdirects flow radially outward frompump axis.

LowVelocity

In

Impeller

High VelocityOut

LowVelocity

In

Figure 2.1b Mixed Flow Impellerdirects flow both radially andaxially to pump axis.

High VelocityOut

LowVelocity

InImpeller

Figure 2.1c Axial Flow Impellerdirects flow axially along the pumpaxis.

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ImpellersThe impeller, or more specifically, the impeller vanes, are that part of the pump

where the rotative shaft energy from the driver is converted into kinetic energy in thepumped liquid. Radial and axial flow impellers perform this conversion using differentmechanisms while mixed flow impellers combine the two methods.

Radial Flow ImpellersA radial flow impeller is essentially a rotating disk with several evenly spaced radial

vanes protruding on one side (Figure 2.2). Liquid is guided into the “eye” or center ofthe impeller via the suction passage of the pump casing where it is then caught by theleading edges of the vanes. (The vanes are usually curved backward against the directionof rotation, this will be explained shortly.) Once caught by the vanes, centrifugal forcedrives the liquid toward the periphery of the impeller (hence the name centrifugal pump).

The liquid accelerates as it travels outward. One way to understand the change invelocity is to think of it as having 2 vector components. One component, U, is equal toradial distance (r) times angular velocity (ω). The other component, VT, is the velocity ofthe flow tangential to the vanes and is related to the velocity of liquid flowing through thevane passages (Figure 2.3).

Rotation

Figure 2.2 Flow of liquid through radial flow impeller. Pumped liquid enters center of impellerwhere it is caught by vanes and driven outward by centrifugal force. Total velocity of the liquidincreases as it moves further out toward the periphery of the impeller.

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The velocity at inlet and outlet is determined by summing the respective U and VTcomponents (vector components are summed graphically by placing the tail of one to thehead of the other). A visual comparison of the length of the total velocity vectors at inletand outlet shows that total velocity V is greatest at the periphery of the impeller. Theincrease in U from inlet to outlet accounts for the gain in total velocity V.

From this description, we see that the net increase in velocity head is the differencebetween V2 and V1. V1 cannot be counted as energy gained because it was alreadypresent in the liquid prior to entering the impeller [8].

While this description represents an overly simplified and highly idealized approach,it is a useful model for understanding how radial flow impellers boost velocity head andit provides a basis from which actual pump performance can be calculated. There areenergy losses (for example, fluid and mechanical friction) which cause the actualperformance to be less than ideal. Manufacturers include these losses when estimatingthe actual performance. Even with the best estimation techniques, the actual performancemust always be determined by testing.

V1

U2VT2V2

VT1

+

r2

r1ωωωω

= r2ωωωω

U1 = r1ωωωω

Figure 2.3 Vector Components (U and VT) of Total Velocity (V) at Impeller Inletand Outlet. The vector components sum to the make up the total velocity. Thegrowth in U from inlet to outlet explains the difference between V2 and V1. Thebackward curve of the vanes determines to what degree an increase in VT2 reduces V2.Since the direction of VT2 is tangent to the vane, the net effect of increasing backwardvane curve is an overall reduction in total velocity, V2, as VT2 increases. Note alsothat U2 depends on impeller rotative speed , ω , and impeller diameter, r2. Anincrease in either variable results in increasing total velocity, V2.

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Axial Flow ImpellersAxial flow impellers are usually

included in the general category ofcentrifugal pumps. However, they do notuse centrifugal force to increase velocityhead. Rather, axial flow impellers operateon the same principles as propellers (Figure2.4). Axial flow vanes are shaped to pushthe liquid in the direction of the pump axis,unlike radial flow pumps that move theliquid at right angles to the shaft axis.

Mixed Flow ImpellersThe third type of centrifugal pump impeller is really a combination of the two just

described. As Figure 2.1b shows, liquid is accelerated radially and axially. Mixed flowimpellers combine goemetric features of both radial and axial flow impellers.

The basic shape of each of the three types of impellers provides a differentcombination of head versus flow characteristics. Radial flow pumps are used in high(pressure) head, low flow applications while axial flow pumps are used where low(pressure) head, high flow is needed. Mixed flow pumps occupy a range in between.

These head versus flow characteristics are described by a parameter known as apump’s specific speed. This meaning of this term is explained in the section belowentitled “Specific Speed”.

Impeller Design Features:1. Vane CurvatureThe vanes of most radial and mixed flow impellers are curved backward (i.e.,

opposite the direction of impeller rotation) as shown in Figures 2.2 and 2.3. A few radialflow impellers are produced with straight vanes.

The backward vane curvature is partly responsible for the decrease in head as flowthrough the pump increases. (The main contributor to this head versus flow characteristicis frictional resistance in the pump flow passages, backward vane curvature serves toaccentuate this effect.) Decreasing head versus increasing flow is desirable from thestandpoint of pump operation as explained in more detail below in the section entitled“Pump Operation: How Pump and System Curves Relate”.

In order to understand how backward vane curvature produces this effect, we need torefer to the velocity triangles in Figure 2.3. There we see that the total velocity head, V,at inlet and outlet can each be thought of as the sum of two individual vectorcomponents, U and VT. The value of U depends on impeller radius and on angularvelocity (i.e., shaft rotative speed) and remains constant when these variables are heldconstant. The value of VT rises and falls as a function of the amount of flow through theimpeller (this is no different than, for example, the change in velocity of water through agarden hose as more or less water flows through it.) In addition, the direction of VT istangent to the vane surface. Since VT flows tangentially to the vanes which are directed

Figure 2.4 Axial Flow Impeller.

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mostly opposite to rotation at the point where they contact the impeller periphery,increases in VT act to cancel U. If U is held constant, then the net effect of an increasein VT is a reduction in the total velocity head, V.

One can see that the reduction in total velocity head with increasing flow dependsdirectly on the degree of backward vane curvature. More backward curvature results ingreater decrease in total velocity head with increase in flow.

2. Impeller DiameterThe outside diameter of the impeller is another important design parameter that

determines the amount of velocity head that an impeller can develop. Knowing that theone of the two vector components which sum into the total velocity head is U and that itis directly dependent on outlet radius and angular velocity (or shaft rotative speed), werealize that increasing or decreasing outlet radius will have a like effect on the totalvelocity head, V (Figure 2.3).

Increasing the head that a pump can produce by increasing impeller diameter is notwithout a cost. Any increase in diameter will also require an increase in the torquerequired to turn the impeller. Since power is a direct function of torque and shaft rotativespeed, a larger diameter impeller will require more power for a fixed speed.

Minor reductions in impeller diameter (called “trimming”) are sometimes made inorder to reduce the pump discharge head or to reduce the pump power consumption. Forexample, the discharge piping could burst if it does not have the strength to withstand thepressure head created by an oversize pump, or maybe the piping can handle the extrapressure but the excessive power consumption of an oversize pump is uneconomical.From the previous discussion of vane curvature and impeller diameter, we know that thisproblem could be dealt with by modifying any one of several variables that determine thehead a pump will produce. Flow could be increased (which raises VT thereby loweringV), shaft rotative speed (or angular velocity, ω) could be reduced (which lowers Uthereby lowering V), or vane curvature could be increased (which causes VT to furthercancel U thereby lowering V), or impeller diameter could be reduced (which lowers Uthereby lowering V). Except for reducing impeller diameter, these parameters are not soeasily changed in most applications. Generally, the process determines the flow and thedriver is often a single speed electric motor. Furthermore, the vanes are an integral partof the impeller and are not modifiable unless the impeller is changed. This leavesreducing the impeller diameter as the most cost-effective solution for situations of thiskind.

Pump casings are designed to accommodate a range of impeller diameters whichallows impellers to be removed, trimmed, and reinstalled in the same pump casing. Thepump affinity laws predict the effect that trimming will have on pump performance.While it is beyond the scope of this paper to explain these laws, they deserve mentioningbecause they are so widely used for predicting how impeller trim will affect head versusflow characteristics.

3. Open versus ClosedThis categorization indicates whether the impeller vanes are enclosed by shrouds on

the front, back, or front and back of the vanes. These shrouds are an integral part of theimpeller and thus rotate with it. The shroud of the impeller shown in Figure 2.2 isrepresented by the flat disk which supports the vanes.

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The shrouds serve to keep the liquid flowing through the vane passages in the properdirection. The pumped liquid is naturely driven back to suction by the high dischargepressure. Enclosing the vane passages with shrouds helps to keep the liquid flowing inthe intended direction instead of leaking back to suction through open areas between thevanes and the sidewall of the casing. This leakage represents wasted pumping energyand a reduction in efficiency.

The impeller shown in Figure 2.2 is actually a semi-open impeller because it isenclosed on one side by a shroud. Strictly speaking, an open impeller is simply a hubwith vanes attached to it [1]. Closed impellers have shrouds on both the front and backsides (Figure 2.14). Closed impellers provide the greatest reduction in leakage andtherefore are more efficient than open impellers. However, they are also more expensiveto manufacture. In addition, a tight clearance running joint must be provided between theimpeller shrouds and the pump casing. This joint is usually lined with replaceable wearrings which also add cost to the pump (see section entitled “Wear Rings” below).

4. Single versus Double SuctionImpellers can be designed with suction inlets, or “eyes”, on one or both sides. Single

and double suction impellers are shown in Figures 2.14 and 2.15 respectively. Doublesuction impellers have several advantages over single suction impellers. First, theirdesign provides a better balance of the axial forces that occur when pumps are operatedoff of design capacity (see section entitled “Thrust Balancing”). Second, double suctionimpellers have a larger suction area than single suction pumps for a given flow whichmeans that less energy is required to push flow into the suction. In other words, doublesuction pumps have a lower Net Positive Suction Head Required (NPSHR) than singlesuction pumps (NPSHR is explained below in “Net Positive Suction Head”).

Single suction impellers are usually preferred because they are less expensive. Singlesuction pumps are easier to manufacture and less likely to clog when handling suspendedmatter such as sewage [1].

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CasingFunctionIt is apparent that the liquid and impeller must be contained in some kind of vessel

which directs the flow toward the discharge. However, there is more to the case designthan simply catching and containing the high energy liquid as it comes off the impeller.The pump case has another equally important role – it must convert some of the velocityhead into pressure head.

As liquid leaves the impeller periphery its velocity head is very high – in fact it is toohigh and the pressure head is too low for many applications. Some of that velocity headmust be converted into pressure head in order to be useful.

The conversion of velocity head into pressure head occurs in the pump case. Figure2.5 shows the how velocity headdecreases while pressure headincreases as the flow movesthrough the discharge side of thecase. The conversion processfollows the principle ofconservation of energy as statedby Bernoulli’s law. Since totalamount of energy must remainconstant (assuming not losses orgains), pressure head mustincrease as velocity head isreduced.

The way to reduce velocity isby increasing the cross-sectionalarea of the flow through theprocess of diffusion. Simply put,diffusion occurs when flow areais expanded. The expansioncauses a reduction in velocityand an accompanying increase inpressure. There are two commoncase designs which accomplishthis in an efficient manner(“Efficient” in this instance means turbulent liquid flow is non-recover

Casing Design:1. VolutesThe most common type of cas

casing is so called because of its expanding flow passage necessary f

The smallest point of flow areacutwater divides the liquid coming through the volute and the other sid

FGhct

Pressure

Velocity

CasingSuction

Impeller CasingDischarge

Flow Path

Outlet Tipsof ImpellerVanesInlet Tips

of ImpellerVanes

igure 2.5 Velocity vs. Pressure Head of Flow Through Pump.raph shows the relationship between velocity head and pressureead of flow through the pump. Casing discharge is designed toonvert velocity head into pressure head while preserving theotal amount of head.

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no energy losses through turbulence. The energy inable).

ing design is the single volute casing. The volutespiral shape (Figure 2.6). The volute provides theor diffusion to occur. where the volute begins is called the cutwater. Theoff the impeller into two flows with one side flowinge flowing toward the discharge.

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Figure 2.7a Balanced radialforces in pump operated atdesign capacity. Balancedradial forces produce aminimal net force on theimpeller.

F

Figure 2.7b Unbalancedradial forces in pumpoperated below or abovedesign capacity. Unbalancedradial forces result in a netforce, F. Direction andmagnitude of F varies withflow.

The clearance between the cutwater andthe periphery (Figure 2.6) of the impeller is acritical design dimension because it must besized to strike a compromise betweenefficiency and pressure pulsations. Pumpefficiency increases as the clearance betweencutwater and impeller is reduced. However,if the clearance is too small, then largepressure pulsations resulting in pump failurecan occur. This is explained in more detailbelow in the section entitled “Vane PassFrequency”.

Volutes have an additional drawbackwhich has been the cause for many brokenpump shafts and failed seals and bearings.The pressure of the liquid in the volute actson the projected area of the impeller to produce a radial force. Since the pressure actsaround the full circumference of the impeller, it actually produces many radial forces thatact in all directions upon theimpeller (Figure 2.7a). Theseradial forces are generallybalanced when the pump isoperated at its Best EfficiencyPoint (BEP). The sum total ofthese nearly balanced forces is anet radial force that is minimal or,in some cases, practicallynonexistent. However, whenoperated above or below BEP, theforces become unbalanced whichcan result in a significant netradial force, F (Figure 2.7b). Thedirection and magnitude of thisnet radial force will varydepending on operating point relative to design capacity. The net radial force can be asmuch as fifteen times the force at design capacity (Figure 2.10).

(The term Best Efficiency Point (BEP) is described below in the section “BestEfficiency Point”. However, as the name implies, the pump is producing the maximumoutput per input when operated at the BEP.)

Despite this drawback, the single volute is still the most commonly used case design.Part of the reason for this is that it costs less to manufacture than other designs [2].However, the problems presented by excessive radial forces that result from operating offof design capacity have spawned the following case designs.

Single Volute

Direction ofRotation

Cutwater

Cutwater-to-ImpellerClearance

Figure 2.6 Single Volute Pump Casing. Thespiral shape of the volute can be seen in the figure.The cutwater divides the flow coming off theimpeller. The clearance between cutwater andimpeller must be optimized for best efficiency andlowest pressure pulsations.

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Double Volute

Direction ofRotation

Figure 2.8 Double Volute Pump Casings containan additional volute positioned 180° to volute ofcasing sidewall. The double volute balances thesideloads produced when the pump is operated offof design capacity.

Radial Load

DoubleVolute

SingleVolute

VanedDiffuser

FlowBest Efficiency Point (BEP)

+

+

Figure 2.10 Radial (sideload) force versus flow forthree types of pump casings. Radial force is minimumwhen flow is at the Best Efficiency Point (BEP).Vaned diffuser produces least amount of radial forcewhen the pump is operated off of BEP.

Vaned Diffuser

Direction ofRotation

Impeller

Casing(Concentricshown)

Figure 2.9 Vaned Diffuser. Note that the number ofimpeller and diffuser are not the same. The unequalnumber of vanes minimizes pressure pulsations whichwould be magnified if the same number of each wereused.

2. Double VolutesDouble volute casings have two opposing

volutes positioned 180° opposite of each other(Figure 2.8). This feature makes the doublevolute more effective than the single volute atminimizing the radial loads produced byoperation away from the BEP (Figure 2.10).

3. Vaned DiffusersAnother type of casing design which balances

hydraulically produced sideloads is the vaneddiffuser. The vaned diffuser contains severalvanes set around the periphery of the impeller(Figure 2.9). Essentially, each vane acts as aminiature volute. The distribution of manyevenly spaced “small volutes” makes the vaneddiffuser the most effective of the three casingdesigns at minimizing sideloads (Figure 2.10).The vaned diffuser may be contained in either aconcentric or volute shaped casing.

Pump designers pay careful attention to the number of impeller and diffuser vanes.Equal numbers of vanes and certain other combinations can lead to destructive highvibration which occurs at the “vane pass frequency” (see section “Vane PassFrequency”). Designers follow guidelines to avoid these combinations of vane numbers.

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Multiple StagesMultistage centrifugal pumps are, as the name implies, pumps which contain more

than one impeller (each stage represents an impeller and casing volute). Centrifugalpumps with as many as fifteen or so stages are not unusual (even twenty stage pumpshave been built). The stages are connected in series so that the discharge from one stageflows into the suction of the stage immediately downstream. Each stage increases thehead by a certain amount with the total head added by the pump being the summation ofall the stages.

The reason for designing multistage pumps lies in the fact that efficiency suffers iftoo big an increase in head is attempted with a single stage. Thus, in applications whichrequire high head, increasing head in smaller incremental steps using multiple stagespreserves efficiency.

Common applications that require high head multistage pumps are boiler feedwaterpumps, reactor feedpumps, and pipeline booster pumps. Vertical pumps can also bemultistage.

These are a few notable particulars about the design of multistage centrifugal pumps:• cases may be split axially or radially.• impellers are generally single suction. However, double suction impellers are

sometimes used in the first stage since they reduce the required NPSH of a pump.• the angular orientation of impeller vanes and volutes are offset, or staggered,

between stages. The staggered orientation results in a more balanced angulardistribution of the radial sideloads produced at the different stages. In addition,this helps to reduce pressure pulsations and vibration which occur at vane passfrequency.

Inlet GeometryThe inlet geometry refers to the flow passage in the pump casing from the point

where the inlet piping attaches at the suction flange to the point where the liquid contactsthe impeller (Figure 1.2). It is a general term that includes all casing geometry from thepump’s suction flange right up to the eye of the impeller. (The inlet of a pump is alsoreferred to as the suction.)

The inlet geometry of a centrifugal pump is worth discussing for the reason that is amajor determinant of a pump’s required Net Positive Suction Head (NPSH). NPSH is thepositive pressure required at the pump inlet (i.e., suction flange) to prevent the pumpedliquid from vaporizing into bubbles, or cavities, within the low pressure regions of theimpeller vane passages. NPSH is a very important parameter because centrifugal pumpswill cavitate if the available NPSH falls below the required NPSH. NPSH and cavitationare explained in much greater detail below in their respective sections, however, a briefdescription is as follows. Cavitation occurs when the pressure (i.e., the Available NPSH)of the incoming liquid falls below the vapor pressure of the liquid. Cavitation is anundesirable condition because the vapor bubbles can cause serious damage to theimpeller when they collapse. The minimum pressure required to prevent cavitation istermed the Net Positive Suction Head Required (NPSHR) because it denotes the pressureat the suction flange that is required to prevent the liquid downstream within the flowpassages of the impeller from flashing into vapor bubbles (i.e., cavitating). Since

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pressure in the impeller must stay above the liquid’s vapor pressure to avoid cavitating, itfollows that the pressure measured upstream at the suction flange must also be kept abovesome minimum level. This difference in pressure is caused by the frictional losses thatoccur as the liquid flows through the inlet passage between these two points.

Therefore, a well designed pump inlet will minimize these frictional losses so thatless NPSH is required. A pump that requires less NPSH is advantageous because thispressure must be provided by the inlet piping system and this can result in greater overallexpense of the installation. Good inlet design practices can include large suction sidediameter, smooth flow passages, and other features which reduce frictional losses. Thesuction specific speed is a number that provides a way to compare the effectiveness of apump at reducing the NPSH required to prevent cavitation (see “Suction SpecificSpeed”).

Two commonly used suction configurations are the end-suction (Figure 1.2) and in-line suction designs. It is beyond the scope of this paper to describe these designs,however, the reader should be aware of their existence and their importance. Thereferences listed at the end of this paper provide greater detail on these as well as othertypes of inlet designs [1,2].

SealsSeals are a particularly important pump component because they are probably the

most frequent cause of regular pump maintenance and thus the cause of a high percentageof overall pump maintenance cost [2]. They are also frequently one of the first parts to beaffected by a malfunction. Seals are critical items because the leakage of a hazardous ortoxic liquid can have severe safety and economic consequences.

Other than the sealless designs described in the next section, all centrifugal pumpsrequire a seal where the shaft penetrates the case. The seal must prevent the highpressure liquid contained in the case from leaking through the joint where the rotatingshaft (or shaft sleeve) and stationary components are in contact. There are two generalapproaches to centrifugal pump shaft sealing: packing and mechanical seals. Eachapproach will be explained separately with a brief comparison to follow.

PackingPacking is the oldest and one of the most common shaft sealing systems for

centrifugal pumps [2]. The main components are shown in Figure 2.11.The term stuffing box refers to the general part of the pump that houses the packing

assembly. It is the point where the shaft penetrates the casing. The stuffing box can beeither an integral part of the casing or it can be a bolt-on assembly. The stuffing boxcontains the stuffing box throat which is an annular space surrounding the shaft (or shaftsleeve if there is one). Square cross-section packing rings of a pliable material (usually afibrous or metallic substance) are “packed” into the stuffing box throat. This creates theseal between the shaft (or shaft sleeve) and casing. These rings must be held tightly inplace and this is done with the packing gland. The packing gland is a collar that iscompressed against the rings, typically by studs with nuts that can be tightened orloosened to provide compression as needed.

The packing gland is situated on the atmospheric side of the gland assembly so that itcan be accessed for tightening or loosening. On the opposite, or high pressure end, of thepacking rings is the stuffing box bushing. The stuffing box bushing provides a seat

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against which the ringscan be compressed. Theannular clearancebetween this bushing andshaft (sleeve) is tight toprevent the packing fromextruding into the pumpand loosing compression.This tight clearance alsolimits the flow of liquidthat can leak in case thepacking fails completely.

An effective stuffingbox seal involves a fairlyhard radial “squeeze”, orcompression, of the non-rotating packing ringsdown onto the rotatingshaft (or sleeve). Thiscreates a hard contactbetween the packingrings and shaft that mustbe kept lubricated.Without this lubrication,the packing ring materialwill burn and the shaft sleeve will wear excessively. Thus, a means of lubricating thepacking must be provided.

This lubrication is typically accomplished by routing a small amount of pumped fluidinto the packing rings via a lantern ring (Figure 2.11). The lantern ring has spaces thatallow the lubricating fluid to flow circumferentially and seep into the joint between thepacking rings and shaft (sleeve) in order to provide the necessary lubrication. Sometimesit is necessary to use other lubricating fluids if the pumped liquid has poor lubricity orcontains abrasives. The leakage of lubrication can be significant, on the order of severaldrops per minute.

Lubrication of the packing is not the only design feature necessitated by the hardradial compression. High wear of the rotating shaft surface is inevitable even withlubrication. To avoid periodic replacement of expensive pump shafts, almost all packedpumps use less expensive, replaceable shaft sleeves made of a hardened or hard-coatedmaterial (see “Shaft Sleeves” below).

Mechanical SealsMechanical seals have been developed to address the shortcomings of stuffing box

and packing gland assemblies. The main components are shown in Figure 2.12.Mechanical seal designs are quite varied but all based on the same general concept.

They do not attempt to seal directly against the rotating shaft as does a packing assembly.Rather, a mechanical seal moves the joint off the shaft and places it between a pair ofsealing faces, one which rotates with the shaft and the other which is stationary with the

PackingGland

Packing Rings (5)

Lantern RingStuffing Box

Shaft Sleeve

Stuffing BoxBushing

Shaft

Lubrication

Figure 2.11 Stuffing Box and Packing Assembly. Packing glandcompresses the packing rings against the stuffing box bushing and shaftsleeve to seal in the high pressure liquid on left side of assembly. Lanternring receives pressurized lubrication via the threaded passage above.Lubrication flows circumferentially through the lantern ring to lubricatethe packing rings-to-shaft sleeve contact area.

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case. The mechanical seal facesare oriented perpendicularly tothe shaft axis and held in contactby one or more springs. Thus,mechanical seal designs havebeen able to eliminate the radialcompression required by stuffingbox and packing assemblies.

While the amount of leakagethrough a mechanical seal isgenerally less than throughpacking, some is still requiredfor lubrication. The need forlubrication exists because therotating-to-stationary seal faceswould quickly be destroyed ifallowed to run dry.

Mechanical seal designsemploy various means oflubrication. The lubricatingfluid can be gas or liquid.

There are other variations inmechanical seal designs. Theseare described in more depth inthe references listed at the end of this paper [1,2].

Mechanical Seals versus PackingMechanical seals generally leak less than packing, boost mechanical efficiency due to

lower friction losses, require less maintenance than packing (when properly selected,installed, and operated), and can handle higher pressures. Thus, mechanical seals arewell suited for applications where leakage of pumped liquid through packing wouldcreate safety, environmental, or production problems, such as with toxic or radioactiveliquids.

On the other hand, mechanical seals have a few disadvantages when compared topacking. When they fail, they usually do so much more quickly and catastrophically.Also, their initial cost is generally higher and they are less tolerant of axial shaftmovement.

Sealless Pump DesignsThere are some applications where even the low leakage of mechanical seals is

unacceptable. The number of such applications has grown as environmental and safetyregulations have become increasingly stringent. The demand for zero-leakage pumps hasgiven rise to sealless designs. Two types of sealless pump designs will be mentionedhere [2].

Magnetic drive pumps and canned motor pumps eliminate the shaft-through-casepenetration and its associated seal. This is accomplished by enclosing the rotating partsin a cylindrical containment shell that tightly wraps the rotor. In magnetic drive pumps,

EndPlate

Spring

Fixed Collar(rotates withshaft)

Seal Face(rotating)

Seal Face(fixed)

Figure 2.12 Simplified illustration which shows the generalelements comprising mechanical seals. Dynamic seal occursbetween the rotating and fixed seal faces. Spring maintainscontact pressure between the faces. Variations in design due totype of lubricating fluid and contacting/non-contacting are notshown.

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an outer magnet external to the shell is rotated by a separate driver motor. The flux fromthe outer magnet then passes through the containment shell and turns an inner magnetattached to the impeller. In canned motor pumps, the motor rotor and pump impellershare a common cavity with the rotor and impeller inside the containment shell and thestator outside of it. Similar to the magnetic drive pump, there is no need for mechanicalseals because only the flux of the motor penetrates the containment shell.

One notable aspect of both types of designs is that they may contain theconfiguration by which the malfunction of fluid-induced instability (a.k.a., whirl andwhip) can occur. Briefly, any design configuration that rotates an inner cylinder at adifferent speed inside an outer cylinder with a fluid trapped between them in a closeclearance is susceptible to fluid-induced instability (see “Fluid-Induced Instability”).

Sealless pumps may have this design configuration in possibly two locations. First,both types of pump have a rotor turning in close proximity to a stationary outercontainment shell with the process fluid in between them. Secondly, both designstypically use radial fluid-film (sleeve) bearings that are lubricated by the process fluid.As explained below in “Fluid-Induced Instability”, this configuration is one of theconditions necessary for whirl or whip to occur. Whirl and whip can cause highamplitude vibration that is very destructive.

In general, sealless pumps are not known for vibration stemming from fluid-inducedinstability. However, the potential exists and it has been known to occur.

Wear RingsIf we step back and consider that the basic function of a pump is to raise pressure, we

realize that it is much higher onthe discharge side of the impellerthan it is on the suction side.The pressure difference betweensuction and discharge acts todrive the liquid back towardsuction, i.e., in the wrongdirection. The liquid can notflow backwards through theimpeller vane passages becausecentrifugal force drives it in theproper direction (that is, unlessrecirculation is occurring –recirculation is explained belowin “Cavitation”). However, thesuction-to-discharge pressuredifference will cause the liquidto leak back through any otheravailable paths. If the impelleris closed, the liquid can leakback to suction in the spacebetween the impeller shroud andthe pump casing. If the impeller

Impeller

CaseWear Ring(Case Mounted)

Wear Ring(Impeller Mounted)

Retaining Screws

Leading Edgeof Impeller Vane

Double Flat RingsFigure 2.13 Wear Rings. Replaceable wear rings are mounted tocase and impeller at the close-clearance gap called the leakagejoint. Clearance should large enough to prevent contact betweenrings and small enough to minimize leakage from discharge backto suction. The “Double Flat Ring” style is shown. Several otherwear ring configurations are also available.

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is open, liquid will leak back to suction over the vane tops. Such leakage meansdecreased pump efficiency because the work previously done to move the leakage fromsuction to discharge has to repeated. This work represents wasted energy.

In order to prevent this leakage, the clearances between certain points of the impellerand the case are made as tight as possible. For open impellers, this means that the vanetops should run close to the case without touching. With closed impellers, clearancesbetween the casing-mounted and impeller-mounted wear rings are kept to a minimum(Figure 2.13). (Less frequently, the tight impeller-to-case clearance is located at theperiphery of the impeller on the discharge side.) For closed impellers, these clearancesvary from about 0.30 mm (0.012 inches) to about 0.76 mm (0.030 inches) depending onoverall impeller diameter [1].

The tight clearance between these rotating and stationary parts can present adrawback. Though not intended to rub, these surfaces may contact and wear away whichwill open up the clearance resulting in lowered efficiency. In addition to wear fromsurface contact, a corrosive or abrasive liquid can also wear away these surfaces with thesame detrimental effect. To overcome this problem, one or both of the wear surfaces isusually fitted with a renewable ring – called a wear ring. The name is a bit misleadingbecause under good pump operating practices these rings should never contact.However, their design allows them to be replaced should wear occur.

Shown in Figure 2.13 is a double flat-ring, one of just several wear ring designs.Other wear ring types are single flat-rings, L-type rings, and labyrinth type rings.

Shaft SleevesCentrifugal pump shafts are usually fitted with a sleeve which protects the shaft from

wear at stuffing boxes or where it is in contact with corrosive and abrasive liquids (Figure1.2). These sleeves are renewable parts meant to be replaced during pump overhaul.They deserve mention because they are a common part of many centrifugal pumpdesigns.

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DischargePressure

DischargePressure

SuctionPressure

Single-Suction (Closed) Impeller

AxialThrustForce

WearRing

Figure 2.14 Axial Thrust Force of Single-Suction Closed Impeller. The difference betweensuction and discharge pressures produces a net force that acts in the axial direction. The wear ringis located at the leakage joint between the impeller and case. This narrow joint separates thesuction and discharge pressures.

Thrust BalancingCentrifugal pumps experience axial thrust because of the difference between the

suction and discharge pressures acting on the cross-sectional area of the impeller. Figure2.14 shows the pressure distribution surrounding a single-suction, closed impeller. Onlydischarge pressure exerts a force on the discharge side of the impeller. However, thesuction side of the impeller has combination of high discharge pressure and low suctionpressure acting on it. The combination of pressures acting on the suction side of theimpeller are lower than the discharge pressure which acts over the entire discharge side ofthe impeller. The unequal pressure on the two sides results in a net axial thrust force.

The axial thrust force is not a constant. Since it is a by-product of the differencebetween suction and discharge pressure, it will vary as this difference varies. Differentoperating points (see “Pump Operation” below) will produce changes in this difference.Also, axial thrust will vary with impeller diameter. A single-suction impeller that istrimmed (trimming is a common practice of machining down the outside diameter of animpeller so that it will produce less head and consume less power) will have differentsuction-to-discharge pressure than an untrimmed impeller. Thus, trimming can alsoaffect axial thrust. Consequently, centrifugal pumps must be designed to handle an axialthrust force that varies with operating conditions.

Double suction pumps contain opposing impeller vane passages that theoreticallyshould cause axial forces to cancel. Figure 2.15 shows how the impeller is designed toprovide a balanced distribution of suction and discharge pressures. However, othercircumstances often disrupt this balance resulting in axial thrust [1], such as:

• Unequal flow into the two suction eyes. Can be caused by elbows (elbows arebends in the piping) located too close to pump suction.

• Unequal leakage through the two leakage joints. Can be due to uneven wear atthe wear rings.

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• Unequal discharge pressure between the two sides of the discharge. Can becaused by asymmetrical waterways or an impeller located off-center.

Since axial thrust is a certainty in most centrifugal pump designs and a possibility inothers, all centrifugal pumps incorporate thrust bearings. In addition to thrust bearings,other balancing devices may also be incorporated. Single-suction pumps may havebalance holes through the impeller that allows suction pressure to leak to the dischargeside. An additional wear ring on the back side of the impeller prevents excessive suction-to-discharge leakage. The wear ring on the back side of the impeller is situated so thatsuction pressure on that side can balance suction pressure on the front side. Multistagepumps may use balancing drums (a.k.a., balancing pistons) and/or balancing disks.These additional balancing devices serve to relieve the thrust bearings of much of theaxial thrust present in the pump. The reader who desires greater detail on the design andoperation of these balancing devices is referred to the references.

BearingsAll rotating machines, including centrifugal pumps, require bearings to support and

position the rotor axially and radially. These bearings must maintain relatively constantrotor position under loads that fluctuate.

The most common types of bearings used in centrifugal pumps are either fluid-filmbearings or rolling element bearings.

Fluid-film bearings.Most fluid-film bearings are oil lubricated. The simplicity and load-carrying capacity

of plain, cylindrical bearings (also referred to as sleeve bearings) make it the mostcommonly used type of fluid-film bearing. However, since sleeve bearings sometimesexperience the malfunction of fluid-induced instability (whirl and whip), other designsare also used. Figure 3.33 in the section “Fluid-Induced Instability” shows some of thesedesigns and describes how they reduce the potential for fluid-instability.

DischargePressure

SuctionPressure

Double-Suction Impeller

SuctionPressure

DischargePressure

WearRings

Figure 2.15 Zero Net Axial Thrust Force of Double-Suction Impeller. Suction anddischarge pressures are balanced. Net axial thrust force is minimized.

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Some pump designs rely on the pumped liquid for lubrication. The magnetic driveand canned motor pumps described above in “Sealless Pump Designs” are examples ofpumps that fall into this latter category. Their bearings are completely contained withinthe shell that separates the rotating from stationary parts and so they take advantage ofthe pumped liquid in which their rotating parts are immersed. The applications thatrequire sealless pumps usually do so because of the harmful nature of the liquid to bepumped. The liquid in these applications is often corrosive or abrasive and thus quitehard on bearings. Filtering screens and hardened bearings are a few of the methods thatmanufacturers have used to counteract the detrimental effects of using pumped liquid forlubrication. [2]

Rolling element bearings.Rolling element bearings are very commonly used in smaller centrifugal pumps.

They include ball, roller, and tapered rolling element bearings. Generally, ball bearingshave the greatest application in centrifugal pumps because they are capable of carryingboth radial and axial loads.

Rolling element bearings are sometimes referred to as “antifriction” bearings. Theterm is true in the general sense, however, some friction is still present under normaloperating conditions. For this reason, all rolling element bearings incorporate some typeof lubrication. Lubrication may be grease, oil, and in certain designs, water.

CouplingsCentrifugal pumps require torque from a driver in order to move the pumped liquid

against system resistance. This torque is transmitted from driver shaft to pump shaftthrough a coupling. Since it is impossible to perfectly align driver and pump shafts,couplings must transmit torque while allowing for variation in alignment within aspecified tolerance.

Many pump applications use flexible couplings since they are designed to toleratesmall variations in alignment (excessive misalignment can be a serious pump malfunctionas explained below in “Radial Loads”).

Solid couplings are normally used only where the pump has no bearings and themotor must support the shaft. Vertical pumps are the primary example of this type ofapplication. Solid couplings require extremely precise alignment. Not only must theinitial alignment be very precise but it must also remain very stable under operation.

A type of commonly used centrifugal pump that eliminates couplings is the close-coupled pump. Close-coupled pumps have the pump housing mounted directly onto themotor housing via close-tolerance fits. This allows the pump and motor to share the sameshaft. Because the shaft is one solid piece, no coupling is required.

An additional advantage of close-coupled pumps is that no bearings are required inthe pump. The pump and motor are positioned close enough so that the motor bearingsalone are sufficient to carry the loads generated by the impeller.

The close-coupled configuration imposes special requirements on the shaft (or shaftsleeve) material. Because the shafting penetrates into the casing and contacts the pumpedliquid, it must be able to resist any corrosive effects.

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η

H P

Flow (Q)

Best EfficiencyPoint

Head (H)

Power (P)

Efficiency (η)

“Drooping” Headat Low Flow(Radial Flow Pumps) +

+

Pump Performance CurvesFigure 2.16 Pump Performance Curves. Flow (Q) is the amount of liquid flowing the thepump. Head (H) is energy added to the liquid by the pump. Power (P) is energy supplied tothe pump by the driver. Efficiency (η) is a measure of how well the pump converts the energysupplied to it by the driver into energy added into the liquid. Efficiency peaks at the BestEfficiency Point (BEP). The “drooping” head at low flow is characteristic of some radial flowpump curves. The “drooping” chararcteristic is notable because a pump operated in thisregion can experience unstable operation (see “How Pump and System Curves Relate”).

PERFORMANCE, OPERATION, AND TERMINOLOGY

Pump Performance Curves: Important Pump ParametersThe sole purpose of a centrifugal pump is to use rotative shaft energy from a driver to

raise the head, or energy level, of the liquid flowing through it. The relationship betweenthe rotative shaft energy input by the driver, the head output by the pump, and theefficiency of this energy conversion process is expressed in the pump performancecurves.

The three parameters in Figure 2.16 are plotted against flow, Q, for a constant speed.Most pumps in use at this time are not variable speed, thus the curves provided bymanufacturers’ will show pump performance at a fixed speed only.

The SI units of flow are meter3 per hour (m3/hr) and the U.S. customary units of floware gallons per minute (gpm). The SI and U.S. customary units of the plotted parametersare:

•Rotative shaft energy from driver, P – kilowatts (kw) or brake horsepower (bhp)•Head added to output by pump, H – meters (m) or feet (ft)•Efficiency, η - % of energy output (pump head) to energy input (power from driver)

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Figure 2.16 shows that head decreases as flow increases. This increasinghead/decreasing flow curve shape is often referred to as “rising to shutoff”. The “risingto shutoff” shape is true for all centrifugal pumps with the exception of some radial flowpumps which “droop” at low flow – that is, have a decrease in head as flow decreases.Stable pump operation requires that pumps rise to shutoff. Manufacturers generallyrecommend that radial flow pumps with drooping curves not be operated in the droopingregion.

The steepness of the curve varies depending on the type of impeller. Radial pumpstend to have the flattest curves. Mixed flow pumps have steeper curves while axial flowcurves are the steepest of the three types.

Efficiency (η) is a measure of how well the pump converts the energy supplied to itby the driver into energy added to the liquid. The pump efficiency curve rises, peaks, andthen falls off. A pump operating at the peak (Best Efficiency Point) is producing themaximum head for the least amount of power input. Pump efficiency is affected by thefollowing losses:

• hydraulic losses – frictional resistance to liquid flow through the impeller andcase passages

• volumetric losses – leakage from discharge back to suction past wear rings(closed impellers), or the front of vanes (open impellers)

• mechanical losses – friction between mechanical parts such as seals, packingrings, shaft, and bearings

• disk friction losses – frictional resistance of the liquid trapped between rotatingimpeller (which can be thought of as a disk) and the stationary case

These losses increase the amount of power required by the pump to output a desired flow.The combination of these losses make up the total pump efficiency.

The practice of trimming, described above in the section entitled “Impeller”, involvesmachining down the outside diameter of an impeller in order to reduce the head outputand the power required. Trimming effectively shifts the entire head curve downwardwithout changing its shape. Manufacturers show how trimming affects performance byplotting several curves for a single pump where each curve corresponds to differentimpeller diameter.

System CurvesA pump is not an isolated piece of machinery, it operates within a larger system

(Figure 1.1). When using the term “system” in the context of pumps, we are referring thepipes, fittings, and valves that deliver liquid to the pump suction and carry it away fromthe pump discharge.

A pump application engineer tasked with specifying a pump needs to know howmuch head is required to overcome the resistance of a given flow through the system. Asystem curve displays this information by plotting head on the vertical axis versus flowon the horizontal axis.

Note that the head plotted in a system curve is the energy lost in the liquid due to thefrictional resistance and elevation change. This is not to be confused with the headplotted on a pump curve, that is the energy added to the liquid by the pump. Thus,increasing head on a system curve means more head is being consumed whereasincreasing head on a pump curve means that more head is being produced.

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System curves generally have a parabolic shape (Figure 2.17). Frictional losses areresponsible for this shape because fluid friction increases with the square of the flowvelocity. Changes in elevation of the flow (i.e., flow uphill or downhill) shift the curveup or down without changing its shape.

Pump Operation: How Pump and System Curves RelateThe pump H-Q curve defines the head a pump will produce at various flows and the

system H-Q curve defines the head that a system will consume, also at various flows.When the two curves are put together in a single graph (Figure 2.17), the pump operatingpoint is defined. This operating point occurs at the intersection of the two curves. Thepump will supply exactly the amount of head needed to overcome system resistance atthe given flow.

Pump curves that “rise to shutoff” allow the pump to function in a stable operatingmode. The “drooping at low flow” regions of some radial pump curves are unstablebecause there are two flows for a given head. Pumps operating in the unstable regiontend to “hunt” or fluctuate between the two points as they search for their operating point.These pressure and flow fluctuations result in surging.

The Best Efficiency Point (BEP)The efficiency curve in Figure 2.16 shows that there is one particular flow where

every pump will operate with maximum efficiency - this operating point is known as theBest Efficiency Point (BEP). A pump operating at its BEP is producing head with theminimum amount of losses (pump losses are described in “Pump Performance Curves”above).

The BEP is important for two reasons:

Flow (Q)

SystemCurve

Head (H)

+

+

Pump versus System Curve

PumpCurve

Pump OperatingPoint

Figure 2.17 Pump operating point occurs where the pump and system curves intersect.

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1. Economics: Operating costs are minimum when a pump operates with maximumefficiency.

2. Radial Loads: Radial (side) loads are minimum when a pump operates at its BEP(the sections entitled “Casing” and “Radial Loads” explain the source and effectsof operation off of design capacity in more detail).

Specific SpeedSpecific speed is a very important parameter because it provides a way to characterize

different pump impeller designs with respect to their head (H) versus flow (Q)characteristics.

The differences between radial flow, mixed flow, and axial flow impeller geometriesexist because each design is best suited for providing the different H-Q combinationsrequired by different applications. Radial flow impellers can deliver high head/low flowperformance but not low head/high flow performance. Conversely, axial flow impellerscan only deliver low head/high flow performance. Mixed flow pumps fall in a largegeneral class somewhere in between axial and radial flow. One impeller type does notsatisfy all applications. Even if one impeller type could function at all the differentcombinations of head and flow, its operating efficiency would be poor. Good economicpractice demands that pumps be optimized for the intended application.

Consequently, there needs to be a way to compare pump impellers with respect totheir H-Q optimization. The specific speed, NS, is an index that makes this comparisonpossible. The specific speed is a number that can be calculated for every pump using thefollowing equation:

4/3S HQN

N = (1)

where N = pump rotative speed (rpm),Q = flow at BEP and full impeller diameter (gpm)H = pump head at BEP and full impeller diameter (ft).

The SI unit version of specific speed is NSM where flow, Q, is given in meters3/hourand head, H, is given in meters. The conversion factor between the two is: NS =51.65NSM [2].

As noted, specific speed is calculated at the BEP and full impeller diameter.However, once calculated, the specific speed for a particular pump is constant fordifferent values of rotative speed, N, and impeller diameter. The pump affinity lawsprovide the basis for this fact. The pump affinity laws allow manufacturers and users topredict the effects of impeller trim or speed changes. Further explanation of pumpaffinity laws is beyond the scope of this paper. However, they are worth mentioningbecause they are such important tools to pump manufacturers and users. The referenceslisted provide more detail on this subject [1,2].

In short, specific speed is a function of impeller geometry. If you know a pump’sspecific speed, you can guess its impeller type. Figure 2.18 shows this relationship. Therelationship between impeller geometry and specific speed holds true regardless ofimpeller size. Radial flow pumps provide high head/low flow and thus have a lowspecific speed. In comparison, axial flow pumps provide low head/high flow and have a

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high specific speed. Figure 2.18 shows that between the two extremes is a continuum ofimpeller profiles that will deliver different combinations of head versus flow.

Specific speed is useful when comparing the H-Q performance of different pumps.This is helpful when selecting the best pump design for a particular application. In mostpump applications, flow (Q), head (H), and speed (N) are predetermined. For example,flow will be determined by the process requirements, head by the system friction andgravity losses, and speed by the driver to be used. (In practice, driver speed is somewhatof a rough variable. Electric motors of different speed settings are available.) Combiningthese terms into specific speed, NS, lets the pump application engineer select the rightimpeller geometry for the application.

Net Positive Suction Head (NPSH) and Suction Specific Speed (S)Net Positive Suction Head (NPSH) and Suction Specific Speed (S) are two parameters

that describe how much suction head a pump requires in order to prevent cavitation(cavitation is explained in detail below in the section entitled “Cavitation”).

Net Positive Suction Head (NPSH)Centrifugal pumps cannot “pull” or suck liquid into themselves. (This is true whether

a pump is operating at full capacity or just starting up.) Instead, liquid must be pushedinto them by a driving or “positive” pressure. If the driving pressure is insufficient, theliquid will turn to vapor (or cavitate) at the point in the pump where pressure drops to itslowest level. Friction losses cause the driving pressure to decrease as the liquid flowsfrom the suction flange to the point where the impeller vanes begin to raise the pressure.Since the point of lowest pressure happens to be in the impeller vane passages, this wherecavitation will occur. In order to avoid cavitation, the pressure measured upstream at thesuction flange must be high enough so that the pressure downstream in the impellerremains above the liquid’s vapor pressure at all times.

The driving pressure is properly referred to as the Net Positive Suction Head. As theterm indicates, NPSH is measured at the pump suction flange simply because it cannot be

500

3000

1000

Axis ofRotation

Impeller Profile versus Specific Speed (NS)

Values of Specific Speed (NS)

1000

0

1500

4000

6000

8000

1500

0

2000

0

Radial Vane Francis Vane Mixed-flow Vane Axial Vane(Propeller)

200075

0

Figure 2.18 Relationship between Impeller Profile and Specific Speed. The specific speed numberindicates the H-Q performance of a pump impeller. Radial flow pumps deliver high head/low flowpeformance and thus have a low specific speed. Axial flow pumps have a high specific speed because theydeliver low head/high flow performance. Other impeller types fill the gap in between the two extremes.(This figure taken after Figure 2.15 in [2]).

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measured within the rotating impeller vanes where cavitation can occur. The leastamount of NPSH that will prevent cavitation is termed the NPSH Required (NPSHR).The NPSHR differs from pump to pump because it is dependent on how well the pumpinlet design prevents friction losses. Not only does the NPSHR differ between pumps,but for any given pump it changes with flow. This is explained by the fact that friction isvelocity dependent and velocity is a function of flow.

In addition to pressure, the temperature of the pumped liquid also determines whetheror not cavitation will occur (see “Cavitation”). Consequently, the NPSH Required toavoid cavitation is based on a certain temperature. The temperature of the pumped liquidmust remain below the temperature upon which the NPSH Required values are based inorder to avoid pump cavitation.

Theoretically, cavitation will not occur as long as the NPSHR is less than NPSHavailable in the pumped liquid (the emphasis on “theoretically” will be explainedshortly). The Available Net Positive Suction Head (NPSHA) is really a sum (or net) ofseveral positive and negative pressures acting of the liquid. They are:

• Atmospheric head: the static pressure acting on the liquid, usually atmosphericpressure measured at a known reference point.

• Suction head or suction lift: the elevation head from the reference point to thesuction flange. Positive if reference point is higher than suction flange, negative ifreference point is lower than suction flange.

• Friction head: the friction losses in the piping between suction and referencepoint, therefore a negative pressure.

• Vapor pressure head: the vapor pressure of the pumped liquid at the operatingtemperature, a negative pressure.

Most pump and fluid dynamics reference books show how to calculate NPSHA. Thereader who desires greater detail on this subject may wish to refer to those sources [2].

In actual practice, cavitation can still occur even if the NPSHA exceeds the NPSHR.The reason for this lies with the test method that manufacturers use to determine theNPSHR figures. Manufacturers test a pump by operating it at a steady flow withexcessive NPSHA. The NPSHA is then gradually reduced until the onset of cavitation isdetected. The NPSHA at which cavitation begins is figured to be the NPSHR for thegiven flow.

The problem with this procedure is with the method that manufacturers use to detectthe onset of cavitation. Cavitation is determined to be present when the head producedby the pump has dropped 3% in response to the reduction in NPSHA. This method ismisleading because cavitation is actually present before the 3% breakaway (so termedbecause cavitation causes the curve to “breakaway” from the normal H-Q curve). Thus,merely maintaining NPSHA in excess of the NPSHR may not be enough to preventcavitation damage.

Consequently, most pump operators use a margin or ratio between Available andRequired NPSH to avoid cavitation. Some operators use a fixed margin, for instance 5feet of head minimum difference. Other guidelines on NPSH calculate ratios that varywith certain pump parameters and, sometimes, also the liquid being pumped.

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Suction Specific SpeedAnother important pump parameter is the Suction Specific Speed (S). The suction

specific speed is calculated in a manner similar to the pump specific speed (NS):

4/3NPSHRQN

S = (2)

where N = pump rotative speed (rpm),Q = flow at BEP and full impeller diameter (gpm)NPSHR = the Required NPSH (ft).

The suction specific speed is similar to the pump specific speed in that it also is anindex. However, while specific speed compares impeller geometry, the suction specificspeed compares the pump inlet geometry. Within certain limits, a high suction specificspeed is desirable because it indicates that the pump produces fewer friction lossesthrough the inlet (i.e., has a low Required NPSH) making cavitation less likely.

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3. MALFUNCTIONS OF CENTRIFUGAL PUMPS

GENERAL CONCEPTS

Centrifugal pumps, like other rotating machines, experience the malfunctionscommon to rotating equipment. In fact, those readers familiar with compressors willrecognize their similarity with centrifugal pumps. Both add kinetic energy using thesame principles described earlier and both share some general similarities in constructionand design. However, the fact that pumps handle liquids while compressors handle gasesmust not be overlooked. Liquids are much more dense and viscous than gases andessentially non-compressible. This causes the symptoms of pump malfunctions tomanifest themselves somewhat differently than compressor malfunctions. The heavydamping of liquids acts to suppress shaft relative vibration amplitudes and in particular,subsynchronous vibrations. In addition, the high density of liquids creates fluid forcesthat are not found in compressors and turbines. These fluid forces have a couple ofeffects. First, they are responsible for a few malfunctions that exist only in centrifugalpumps (e.g., Hydraulic Unbalance). Secondly, they often modify the balanceresonances. A pump operated “dry” (no process liquid surrounding the impeller) mayexperience a different (often lower) balance resonance than when run “wet” (impellerfully immersed) [4]. (Caution: some pumps may suffer damage from dry running. Dryruns should only be performed under certain conditions by those fully knowledgeablewith the pump and its requirements for safe operation.)

Pump malfunctions are often accompanied by some general signs. These are “high”vibration, excessive noise, reduced bearing and/or seal life, high bearing temperatures,and poor performance (higher than normal power consumption or lower than normaloutput). The manner in which these signs relate to the various malfunctions will beoutlined in more detail in each of the following sections.

Some of the following malfunctions are discussed in depth in the Machine LibraryMalfunction Diagnosis articles. Those papers that have been written will be noted andreferenced in each section. The reader who desires more detail can refer to them.Otherwise, each section will provide a brief explanation of the general mechanismunderlying the malfunction. Then, each malfunction will be related to those things incentrifugal pumps that may act as root causes.

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PUMP MALFUNCTIONS

High 1X Vibration due to Unbalance

Centrifugal pumps are susceptible to two types of unbalance: mass unbalance andhydraulic unbalance. Mass unbalance in centrifugal pumps is the same malfunction thatwe experience with other types of rotating machines. However, hydraulic unbalance isunique to centrifugal pumps. Whereas both types of unbalance produce high 1Xvibration, they are caused by different phenomena and are corrected by different means.Therefore, the causes and corrective actions of each type of unbalance will be discussedseparately. However, since they produce similar behavior, their symptoms will bediscussed together.

High 1X due to Mass Unbalance

Definition of Mass UnbalanceRotors inherently contain some mass unbalance that causes a 1X vibration. (The rotor

includes shaft, shaft sleeves, impeller, and couplings – anything that is locked to andmoving in unison with the shaft.) This type of unbalance is known as mass unbalancebecause it originates from the mass of the rotor. The resulting vibration varies directlywith the amount of unbalance. If the vibration exceeds a prescribed level, then damage tothe pump, its driver, or attached structures can occur.

Mass unbalance is caused by the fact that the mass center and the geometric center ofthe rotor do not lie at the same point. If we consider a single cross section anywherealong the rotor axis, the mass center is the point about which all the mass is equallydistributed. The mass center can be thought of as the balance point. If you balance a flat,circular object (like a dinner plate) on the tip of your finger, you are supporting it directlyunder the mass center. The geometric center is different, it is the point within the rotorabout which the geometry (or shape) isequally distributed or symmetric. We aregenerally accustomed to thinking of therotor mass and geometric centers as one inthe same, however, this is never true in thereal world. Typically, a rotor is slightlyheavier on one side because ofmanufacturing tolerances, or deposits,pitting, etc. The heavier side causes themass center to be offset in the samedirection. The offset, Rr, between the twocenters is directly responsible for thecondition in a rotor system referred to asmass unbalance (Figure 3.1).

The offset is a problem because, atrotative speeds below balance resonance(or critical speed), rotors turn about their

++Mass

Center

GeometricCenter

UnbalanceForce

Rr (Offset between Mass andGeometric Centers of Rotor)

Figure 3.1 Unbalance Force. A rotor is supported bybearings and, to a much lesser extent, by seals.These supports cause a rotor to turn about itsgeometric center. However, the offset between themass and geometric centers creates an outwardpointing centrifugal force, or unbalance force.

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geometric center and not their mass center. The offset between the two centers creates anoutward pointing centrifugal force, or unbalance force. This force is identical to that feltwhen twirling a string with a rock tied to the end of it. The unbalance force is defined bythe equation for a centrifugal force:

2Ω= rr RMForceUnbalance (Eq. 1)

From this equation we see that the unbalance force depends upon Mr (the mass of rotor),Rr (the rotor Smass to geometric center offset), and Ω 2 (the square of the rotative speed).

A few items about the unbalance force are worth noting. First, the outward directionof the unbalance force explains why it causes vibration to increase. Though this is fairlyobvious, it helps us to visualize the effect of the unbalance force upon vibration when weconsider the fact that it is literally pushing the rotor outward. Secondly, the unbalanceforce points outward from the geometric center in the same angular direction as the masscenter. Thus, the angular direction of the unbalance force will shift only if the masscenter shifts. Since the location of the mass center is generally stable, so is the directionof the unbalance force. The mass center can change only if there are changes in the rotorassembly such as changing shaft bow, deposits, pitting, or a loose rotating part. Sincethese changes are the exception rather than the norm, any apparent shift in the directionof the mass center should prompt us to look for the reason why the shift occurred.Thirdly, the unbalance force rotates at the same speed (or synchronously) as the rotor.This last point is notable because it explains the 1X nature of the vibration response tomass unbalance.

Mass unbalance is discussed in greater depth in the corresponding Machine LibraryMalfunction Diagnosis article [11]. The reader who desires more information on thistopic may wish to refer to that source.

Causes of Mass UnbalancePumps usually come from the OEM in a well balanced state. However, the balance

state can worsen if changes occur in the pump over time or due to maintenance. Whilesome examples of such changes are given below, anything that adds or removes mass orshifts the center of mass may have an adverse affect on a pump’s balance state. The firstthree examples illustrate increase in unbalance without the addition or removal of mass;the mass center shifts simply because parts on the rotor shift relative to each other. Thesecond three examples illustrate an increase in unbalance due to the addition or removalof mass.

1. Replacing OEM parts with more loosely toleranced non-OEM parts.Non-OEM parts may be manufactured to looser tolerances. Parts with looser

tolerances may shift the mass center because they do not sit on the shaft as concentricallyas an OEM part. For example, large, custom pumps such as boiler feedpumps and theattached OEM couplings are typically well balanced by the factory. However,replacement of the coupling with a non-OEM part during maintenance may increase theunbalance. [5]

2. Loose fit impellers.Loose fit impellers are those which mount to the pump shaft with a clearance fit.

They are typically locked against rotation to the shaft with a key. Unbalance can occur

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when a loose fit impeller is removed from the pump shaft and balanced on an expandingmandrel (which is tight fitted), then removed and remounted back on the pump shaft.The effect on unbalance due to the slight shift in mass center between the expandingmandrel and the looser pump shaft becomes increasingly pronounced with largerimpellers. [1]

3. Shrink fit impellers.Shrink fit impellers are locked to the shaft by an interference fit. The impeller and

shaft experience residual stress in the region of the fit. The residual stress can relax overtime due to temperature cycling, shaft vibration and flexing. As residual stress relaxes,impellers can cock or bow the shaft causing the mass center to shift. [1]

4. Clogs.Many pumps handle liquids that contain solid objects in the process stream (e.g.,

sewage pumps). While such pumps are usually designed to freely pass objects of acertain size, clogs can still occur. This is especially true for closed impellers. [5]

5. Deposits.Many pumps process liquids containing substances that can deposit on the impeller.

Deposits will add mass as they collect, however, mass will also be removed if they breakoff. In either case, deposits tend to collect or break off in irregular patterns that upset thebalance state of the rotor.

6. Pitting (impeller erosion).Impellers can lose material through pitting. Pitting may occur for a couple of reasons.

First, the pumped liquid may be so corrosive or abrasive that it erodes the impeller and,especially, the highly exposed vanes. Secondly, pitting can occur even in harmlessliquids for another reason: cavitation. Cavitation is discussed in greater detail below in“Cavitation”. However, suffice it to say here that cavitation can cause very severedamage, even eating holes clear through impeller vanes.

Corrective Actions for Mass UnbalanceIt should be evident from these examples that mass unbalance is often a symptom of

another underlying root cause. While the corrective action will probably includerebalancing the pump, it should also include diagnosis and correction of any underlyingroot cause.

High 1X due to Hydraulic Unbalance

Definition and Cause of Hydraulic UnbalanceCentrifugal pumps are subject to another source of high 1X vibration known as

hydraulic unbalance. While hydraulic unbalance is similar to mass unbalance in itsvibration signature, it has a different underlying cause and, hence, a different correctiveprocedure.

Hydraulic unbalance originates in the fluid forces acting on the impeller. Just as theimpeller acts on the liquid to increase its angular momentum (see “Impellers” above), sotoo the liquid acts against the impeller in an equal and opposite reaction (recall Newton’sthird law). If the liquid does not flow evenly through all of the impeller vane passages,then these reaction forces will be unbalanced.

The references sited in this paper do not describe the exact unbalance mechanism.However, there is one possible explanation consistent with pump theory that does stand

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out. Recall from the earlier discussion of velocity triangles (“Impellers”) that the totalvelocity head added to the liquid depends on two parameters: U and VT where Urepresents the flow velocity due to angular speed at a given radius and VT represents thetangential flow velocity due to flow along the vane surface. Both parameters dependupon the impeller geometry and its tolerances. If the impeller radius varies excessivelyaround the periphery, then U will vary excessively. Likewise, if vane angles are notsymmetrical within a specified tolerance, then VT will also vary excessively. Either oneof these or the combination of the two may cause the fluid forces surrounding theimpeller to be asymmetrical. This mechanism points to unacceptably high geometrictolerances as being the underlying root cause of hydraulic unbalance.

Corrective Action for Hydraulic UnbalanceThe corrective actions stated in the references are also in agreement with the

mechanism just described. The only way hydraulic unbalance can be remedied is byswitching to a more precisely manufactured impeller. In fact, it seems to be wellestablished that for a particular manufacturing method, there is fixed amount ofunbalance that will have to be tolerated. The amount of hydraulic unbalance associatedwith a manufacturing method is described by one reference as KH, where KH is thenormalized hydraulic unbalance force (lb) [4]. Values of KH for a sand-cast impeller areas high as .10 while precision-cast impellers (i.e., investment cast) may have values of.005 to .025 and machined impellers can values as low as .0025.

A recommendation to switch to a more precisely manufactured impeller was thecorrective action suggested in one Bently Nevada Machinery Diagnostic Services casehistory involving a vertical slurry pump. The MDS engineer diagnosed a hydraulicunbalance due to errors in the impeller geometry as the root cause of the high 1Xamplitudes observed. The hydraulic unbalance was confirmed when the high 1Xamplitudes disappeared during a dry run of the impeller.

Vibration Characteristics of Unbalance (Mass and Hydraulic)As noted above, high 1X (shaft relative or casing vibration) is the predominant

vibration component that accompanies mass unbalance or hydraulic unbalance. The high1X will be especially noticeable when its frequency is close to a balance or structuralresonance.

Effects of High 1X VibrationThere are several tight clearances in pumps that are vulnerable to high vibration.

Seals, packing, bearings, wear rings, couplings all contain these clearances and can bedamaged when forced to operate beyond design limits. Also, high 1X can exciteresonances in adjacent structures and cause their stress levels to exceed design limits (see“Structural Resonances” below).

Other Sources of High 1X VibrationMany other malfunctions also exhibit increasing 1X vibration. From our

understanding of rotor dynamics (see Machine Library Dynamic Stiffness and RotorResponse) we know that vibration (or rotor response) is the ratio of unbalance force todynamic stiffness. Therefore, vibration may increase due to reduced system stiffnesswhile the unbalance force remains constant. Some examples of this are: “softfoot” (i.e.,

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loose bolts, degraded foundation), bearing or seal wear, shaft crack, alignment changes,quadrature stiffness changes due to different pumped fluid or lube oil characteristics.

These examples help to illustrate why it is so important to verify the source of the 1Xbehavior before balancing. Balancing a machine will not address a problem whose rootcause is due to some other malfunction, such as a shaft crack.

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Figure 3.2 Perfect internal alignment exists in amachine when the centers of all of the internalparts are collinear.

Radial Loads (Misalignment and Sideload)The term radial load refers to any load that acts on the rotor in a unidirectional radial

direction. When discussing radial loads, we only include those loads whose direction andmagnitude are constant or vary gradually over time with process changes. We are notincluding radial loads, such as unbalance, that rotate with the rotor. Rotating radial loadshave different symptoms and produce different effects from (relatively) constant radialloads.

Radial loads act to push the rotor to one side of the bearing. This effect is not entirelyunwanted. For instance, sleeve bearings are less prone to fluid-induced instability whenthe rotor does not ride in the center of the bearing. Therefore, sleeve bearings are usuallydesigned to take advantage of normally existing radial loads, like gravity, to keep therotor from operating in the bearing center. In contrast, when discussing radial load as amalfunction, we are referring to loads that exceed the design of the machine. Radialloads in excess of design limits can lead to serious pump damage if not detected andrectified in their early stages.

Two common sources of excessive radial load in pumps will be discussed here: 1)misalignment and 2) the radial load on an impeller (or sideload) that occurs whenoperating a pump too far away from its Best Efficiency Point (the Best Efficiency Point,or BEP, is explained above in “The Best Efficiency Point”). Since these two types ofradial loads have different root causes, their underlying mechanisms and their correctiveactions will be explained separately. However, since misalignment and sideload manifestsimilar behavior, their symptoms and effects will be discussed together.

The reader who desires more discussion on the topic of misalignment can refer to thecorresponding Machine Library Malfunction Diagnosis article [12].

Radial Load due to Misalignment

Definition of MisalignmentMisalignment is a very important source of radial load because it is responsible for so

many pump failures. One reference [1] even stated that “Outside of serious unbalance ofpump components, there is no singlecontributor of poor mechanicalperformance more significant than pooralignment.”

In order to understand howmisalignment creates radial loads, wemust first understand the broader conceptof alignment. Perfect internal alignmentexists when the centers of all of amachine’s bearings, interstagediaphragms, and seals are located on thesame line and that line is the centerline ofthe machine (Figure 3.2). Two machineswould be in perfect external alignment ifthe centerlines of their shafts were on the

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same line (Figure 3.3, top). In practice,some degree of internal and externalmisalignment always exists. Flexiblecouplings are designed to accommodate acertain amount of misalignment, and thatamount will depend on the type of couplingbeing used. When the misalignmentexceeds the allowable tolerances for thecoupling in use, the machines are said to bemisaligned.

There are two basic types of externalmisalignment. Parallel misalignment occurswhen the centerlines of two machines havethe same angular orientation, but areseparated from each other (Figure 3.3,middle). Angular misalignment occurswhen the centerlines of two machines havedifferent angular orientations (Figure 3.3,bottom).

An additional type of “misalignment”involves the axial position of two machinescoupled together. Coupled machines canhave correct parallel and angular alignmentbut still suffer incorrect axial alignment.The tolerance for axial position for twomachines will depend on the type ofcoupling that is used. Rigid couplings havea very low tolerance for axial positionerrors, while disk pack and diaphragmcouplings have somewhat more, but stillsmall tolerance for error. Gear couplingshave a higher tolerance for axial positionerrors.

Misalignment is a three-dimensionalproblem. Each machine has a centerlinethat exists at some orientation and positionin space, and the centerline of an adjacentmachine will have a different orientationand position. To make external alignmentproblems easier to solve, the three-dimensional centerlines of the machines areprojected on two perpendicular planes(Figure 3.4). Then, the alignment problemdimensional problems.

Parallel Misalignment

Angular Misalignment

Aligned

Figure 3.3 Two machines are in perfectexternal alignment (top) when their shaftcenterlines are collinear within an allowabletolerance zone (red). With parallel misalignment(middle) the shaft centerlines are offset, butparallel. With angular misalignment (bottom), theshafts are oriented at different angularorientations in space. Misalignment usuallyinvolves a combination of parallel and angularmisalignment. In the figure, the shafts are showncentered in the bearings. In reality, gravity loadedshafts would rest in the bottom of the bearings

ith th hi ff

39

Top View

Side View

Figure 3.4 The 3-dimensional misalignmentproblem is usually broken down into two 2-dimensional problems.

can be treated as two, separate, two-

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Causes of MisalignmentThe condition of misalignment can result from any one or a combination of several

causes. When discussing these causes, it is important to distinguish them from thecondition of misalignment. While the condition of misalignment may cause undesirableeffects in a pump, we must not forget that the misalignment itself usually results from aneven lower level of root causes. This is important to keep in mind because, if we re-alignthe pump without diagnosing and correctingthe underlying cause, then misalignment mayreappear as the situation worsens.

1. Thermal growth.As the temperature changes during a

startup, the linear dimensions of a machinecan change in complicated ways, with the hotparts growing more than the cooler parts.Dimensional changes in the machine supportsand casing can cause the machine to rise orfall and/or change angular orientation as itheats up. Any adjacent machine will alsochange, and that change will most likelyfollow a different pattern. Thus, if the twomachines were aligned in a cold condition,they would become misaligned in a hot condeliberately misaligned in the cold condition, calculated to produce correct alignment in the ho

Because the temperature of a machine can vawith load, and it may be difficult to set a cold alignment for all anticipated operating load condi

2. Foundation problems.Foundation problems can cause a shift in m

problems can include cracked grouting, a loosesoaked concrete can lead to deterioration of the f

3. Soft Foot.Soft foot is a condition where one or mo

tightening hold down bolts. When one foot is tightening down the soft foot will warp the mainadequate shimming or by an excessive numbesupport. There should be no more than 3 to 4 scaused by a warped or bowed soleplate, an imachining of feet or the soleplate, or a foot nobowed machine casing can also cause soft foot.

4. Piping strain.Piping strain can warp a machine casing a

bearing supports. Pipe strain can result from lomissing piping supports. Poor piping fit can put tPiping should never be forced to mate with the m

Cold Misaligned

Hot Aligned

Figure 3.5 Machines are deliberatelymisaligned cold so that, when they reach hotoperating temperature, thermal growth will alignthe machines.

40

dition. For this reason, machines areand the cold misalignment is carefullyt condition (Figure 3.5).ry with load, alignment can also changealignment that produces acceptable hottions.

achine position over time. Foundation soleplate, and loose anchor bolts. Oil

oundation and a loss of support strength.

re machine feet are not coplanar afternot properly supported (the soft foot),

chine casing. Softfoot can be caused byr of shims, which can produce a springyhims under a foot. Soft foot can also bemproperly installed soleplate, impropert parallel to the soleplate. A warped or

nd cause misalignment by moving theose piping hangers or bent, broken, orremendous loads on the machine casing.achine through the use of force.

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5. Improper alignment.While not often the case, it is always possible that the current alignment is not

sufficient and therefore needs to be redone. If other root causes have been carefullyconsidered and ruled out, then it is possible that the pump and its driver simply need to bere-aligned.

Radial Load due to Pump Sideload

Definition of SideloadingThe hydraulic forces that act radially on the impeller cause the second common

source of radial load in centrifugal pumps. These forces combine to create a resultantforce known as a sideload. Sideloads are explained in greater detail above in “Casing”.However, in short, high sideloads occur when centrifugal pumps are operated off of theirBest Efficiency Point (BEP). Single volute pumps are especially vulnerable to sideload.While the magnitude and direction of sideloads vary with flow, they meet the basiccriteria for radial load in that they are relatively constant and only change gradually whencompared to rotative speed.

Sideloads can reach very high magnitudes - high enough to break pump shafts, causerubs or do other serious damage. The magnitude and direction of sideload can undergoextreme variations. In fact, the sideload at shutoff (no flow) can be as high as 10 to 15times greater than at BEP and that direction can change by almost as much as 180degrees [1].

Vibration and Temperature Characteristics of Radial LoadExcessive radial load can be suggested by one or more of the following

measurements:1. High Bearing Temperature.High fluid-film bearing temperature is often the first warning of a possible high radial

load condition. The high radial load can cause high shearing stresses in the lubricatingfluid of an overloaded bearing. The extra work done in overcoming these higher thannormal fluid stresses produces extra heating of the fluid. The fluid, usually oil, transfersthis excess heat to the bearing babbitt.

Oil drain temperature is not a very useful indicator of the temperature in the bearing.It is limited because, at that point, the oil exiting the bearing is a mixture of oil that haspassed through the load zone of the bearing and oil that has bypassed the load zone. It isbest used for plant heat load calculations or oil temperature regulation, but it onlyprovides a vague picture of the machine condition.

Resistive Temperature Devices (RTDs) or thermocouples that are imbedded in thebearing babbitt can provide better warning. Ideally, the RTDs should be installed atseveral different circumferential positions in the bearing. The are two reasons for this.First, in some machines, the direction of the radial load on a bearing can vary withoperating conditions, and it can be difficult to predict where the maximum load occurs ina bearing. Second, if a machine becomes misaligned, load shifting can produce radialloads that act in unpredictable and unexpected directions.

While high bearing temperatures indicate overloading, an abnormally low bearingtemperature indicates that the load in that bearing may be below normal. Given the loadshifting that takes place with radial loading, one bearing may show an unusually high

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temperature, while another, adjacentbearing may show an unusually lowtemperature. For this reason, bearingtemperatures should be monitored andcompared along the machine train.

2. Average Shaft Centerline Position.For a horizontal, fluid-film bearing

machine train which is perfectly aligned,and in which gravity is the primary radialload, the average shaft centerline positionwill change during a startup or shutdown ina typical way (Figure 3.6, green).Normally, the shaft position angle will besomewhere between 0° and 45° from thedirection of the applied load. If themachine train experiences abnormal radialload, then load shifting will cause changesin the behavior of shaft centerline plots(Figure 3.6, red). For example, the radialload due to misalignment can be in adifferent direction, and the direction andamount of the misalignment load canchange as the machines heat up. Heavilyloaded bearings will have operatingeccentricity ratios that are higher thannormal, while lightly loaded bearings willhave operating eccentricity ratios that arelower than normal. If the misalignmentbecomes severe enough, shaft operatingpositions may move to unusual locations,such as near the top of a bearing (Figure3.6).

Differences in operating position canbe most apparent across a couplingbetween two machines, where the rotormay operate in opposite quadrants of thebearings (Figure 3.7).

Average shaft centerline plots shouldbe examined at every axial position andcompared for signs of abnormality.Average shaft centerline plots are mostuseful when clearance circles are knownand included on the plot. That way,operation in an abnormal quadrant can bemore easily detected. Shaft centerline plotsshould be compared to previously archived

SlowRoll

NormalRunningPosition

AssumedRadial Load

AbnormalRunningPosition

Misalignment Load

Figure 3.6 Comparison of normal and abnormalshaft centerline behavior during a startup of atypical, horizontal, gravity loaded machine withplain, cylindrical, fluid-film bearings. Here, theradial load is assumed to be vertically downward.The machine is rotating X to Y (CCW). The dashedcircle defines the bearing clearance. As speedincreases, the hydrodynamic oil wedge becomesstronger, and the normal rotor moves up andslightly away from the bearing wall (green). Whenmisalignment forces are present, the behavior canbe quite different (red), and the rotor can end upoperating in an unusual quadrant in the bearing.(Note that rotors operating in tilt-pad bearingsnormally tend to move straight up toward the radialload with increasing speed.)

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1 23 4

Figure 3.7 The misaligned machine train isshown with the operating average shaft centerlinepositions for each bearing. Note that, for thisexample, the rotor positions in bearings 1 and 4 areapproximately normal, while the rotor positions inbearings 2 and 3 are in opposite quadrants,indicating a possible misalignment condition.

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data and examined for changes.3. Orbits.Orbits can be very helpful for diagnosis of radial load. Normally-loaded, healthy

rotors that operating in plain, cylindrical fluid-film bearings tend to produce direct, orunfiltered, orbits that are elliptical in shape and where the frequency is predominately 1X(Figure 3.8). The ellipticity of such orbits can fall into a wide range and still beconsidered normal.

Elliptical and lemon-bore bearings tend to normally produce orbits that are moreelliptical than those produced by plain cylindrical or tilt-pad bearings. Also, the majoraxis of the ellipse in elliptical and lemon-bore bearings tends to be aligned with thebearing geometry.

Because radial loads can change magnitude and direction with load, orbits can vary insize and shape with load. Also, any resonances will affect the size and appearance of theorbit.

Because of the many possibilities, a database of normal operating orbits for aparticular machine should be archived for later reference.

The shape of a direct orbit is sensitive to the amount of the radial load that acts on therotor. As the radial load is increased, the orbit will become more flattened, and part of theorbit path may partially follow the curvature of the bearing. (Note that elliptical andlemon bore bearings tend to normally produce more elliptical orbits than would occurwith plain cylindrical bearings. For these bearings, the orientation of these elliptical orbitstends to be more aligned with the bearing geometry.) The orbit may also become bananashaped, containing a 2X vibration component that is visible on spectrum plots (Figure3.8C). 2X components can be amplified if the rotor operates at half of a resonance speed.In extreme cases of radial load, the rotor may become so constrained that the orbitfollows a line that matches the curvature of the bearing (Figure 3.8D) or, if unbalance issmall, may shrink to nearly a point. Assuming that unbalance is the primary source ofrotor vibration, the details of the orbit behavior will depend on the degree of radial load,the amount of unbalance forcing in the rotor, and the attitude angle and eccentricity ratio

A B C D

Figure 3.8 Unfiltered orbits. Each orbit shows eight shaft revolutions. Orbit A is a normal orbitfrom a generator bearing on a small steam turbine generator set. The orbit is mildly elliptical andpredominately 1X. Orbit B is from a Frame 6 gas turbine bearing. The orbit shows evidence ofconstraint along the lower edge, suggesting a misalignment problem. Orbit C is from the exciterbearing on a 500 MW steam turbine generator set. Note the highly elliptical, banana shape. Thebanana shape will produce a 2X vibration component which would be strongest in the horizontaldirection. Note the curvature of the right side of the orbit, which suggests that the shaft may befollowing the geometry of the bearing boundary. Orbit D is from a HP steam turbine bearing. Theorbit is highly flattened, suggesting a high, misalignment-induced radial load. (Note that line orbitscan occur for other reasons.)

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1 23 4

Figure 3.9 The misaligned rotor of Figure 5 isshown with possible orbits. All orbitscorrespond to the same operating speed. The orbitsize (vibration amplitude) is partially controlledby the bearing stiffness, which is a function ofeccentricity ratio. Thus, the heavily loadedbearing 2 orbit is small, while the lightly loadedbearing 3 orbit is relatively large. The bearing 2orbit partially follows the contour of the bearing.Bearings 1 and 4 are approximately normallyloaded.

in the bearing.A rotor that is unloaded in a bearing

because of misalignment or sideload mayoperate at a low eccentricity ratio and have anorbit that is nearly circular. Because ellipticalorbits are the norm, a circular orbit suggestsan unusually low radial load that could be dueto misalignment or sideload.

Multiple orbits should be displayed forevery axial position in the machine train andcompared with each other (Figure 3.9). Ifstartup or shutdown data is available, thesemultiple orbits should be examined over theentire speed range of the machine for evidenceof high radial loads. The orbits (which containdynamic position information) should becorrelated with average shaft centerline plots(which contain average position information)over the length of the machine train.

4. Vibration.Assuming that the source of vibration

originates in the rotor (for example, due to unbalance), the amount of casing vibrationwill depend on the transmissibility of rotor vibration through the bearings and into thecasing. (Casing vibration will also depend upon how well the machine is mounted to thefoundation.) The very high fluid-film bearing stiffness that exists at high eccentricityratios acts to more effectively couple the rotor to the casing. Thus, in a radial loadedmachine, the rotor may transmit more vibration to the casing, and the machine mayexperience higher than normal casing vibration. Rotor shaft relative vibration, because ofthe increased constraint on the rotor (increased Dynamic Stiffness), may decrease asmore of the vibration energy is transmitted to the casing.

If, because of radial load, a particular bearing is unloaded, the rotor may becomemore decoupled from the casing (transmissibility will decrease) at that location, and thecasing vibration there may decrease. Under this circumstance, shaft relative rotorvibration may increase as the rotor support Dynamic Stiffness decreases.

Thus, either increases or decreases in casing vibration could be an indication of amisalignment or sideload condition. An increase in casing vibration coupled with adecrease in rotor shaft relative vibration (and vice versa) suggests either condition.

Remember that casing vibration can increase if the machine support structureweakens or loosens, or if the machine develops a soft foot. The reduced stiffness of themachine support allows vibration to increase. Sometimes, tightening loose foundationbolts will reduce casing vibration back to normal levels.

Parallel misalignment at the coupling can produce “cranking” of the rotor shafts. Thiswill usually produce a 1X and 2X shaft relative vibration component that exists over theentire speed range of the machine. The vibration may transmit to the casing, but onlyshaft relative measurements will reveal the cranking action at slow roll speeds. The 2X

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component occurs because of opposed high spots reacting in different bearings. Theeffect is similar to the 2X generated in a bicycle crank.

Misaligned disk and diaphragm couplings can produce an axial “pumping” action thatresults in axial vibration. This axial forcing is available to excite any rotor system axialresonant frequencies. It is also possible for the axial vibration to couple into lateralvibration, showing up in radial vibration measurements. Properly functioning gearcouplings are much more axially compliant and less likely produce axial vibration.

Effects of Radial LoadAs mentioned above, radial loads for which the machine is designed (preloads) can be

beneficial because they suppress fluid-induced instability in sleeve bearings bypositioning the rotor at higher eccentricity. (This effect is discussed in greater detail inthe section entitled “Fluid-Induced Instability” and in the corresponding Machine LibraryMalfunction Diagnosis article [15].)

Aside from the positive effect of suppressing fluid-induced instability, radial load canwreak havoc with many critical parts of a centrifugal pump and be the primary cause ofpump failure. The adverse effects of radial load most commonly include:

1. Rub.Extreme radial load can cause the rotor to wipe bearings and seals or to rub at wear

rings (see “Rub” below). A rub at fluid-film bearing can result in metal to metal contactand wiping of the bearing babbit. A rub at seals or wear rings can open up clearances,resulting in higher leakage flows and a loss of efficiency.

2. Shortened Bearing Life.Bearings can be damaged by high radial load even in the absence of a rub. Normally

loaded fluid-film bearings have a babbitt temperature of 160°F to 180°F (70°C to 80°C).Overloading of a fluid-film bearing will produce higher shear forces in the oil, resultingin higher oil and babbitt temperatures. Bearing babbitt will start to creep at 240°F(115°C) and melt at 260°F (125°C), leading to bearing failure.

Rolling element bearings are also highly sensitive to radial load. Rolling elementbearings have finite lifetimes that are a strong function of radial load. The L10 life (thetime that 90% of similar bearings will survive) for a point contact ball bearing goes downas the third power of the applied load. Thus, load shifting due to misalignment orsideloading can, by increasing the load, drastically reduce the useful life of a rollingelement bearing.

3. Damaged Seals or Packing.Mechanical seals and packing are designed to operate within certain limits of shaft

deflection and position. Radial load can deflect shafts and push seals and packing outsideof their design limits. Re-occurring failure of mechanical seals or packing may indicateexcessive shaft deflection due to radial load.

4. Cracked Shafts.Radial load is cited by more than one reference as a common reason for broken pump

shafts [1, 2]. Radial loads that deflect shafts beyond their design limits can create highreversal stresses. These stresses can fatigue the shaft and cause it to break.

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5. Shortened Coupling Life.Incorrect alignment can shorten coupling life by producing extreme heat in elastomer

couplings. Also, gear couplings can experience extreme wear and dry element couplingscan experience high fatigue.

f=

RubCentrifugal pumps are susceptible to rubs in much the same way as other types of

rotating machines. Beyond the similarities, though, there is a key difference. Thesubsynchronous vibration that can accompany rub in other types of rotating machines isseldom, if ever, seen in pumps.

Aside from this key difference the discussion of pump rubs basically mirrors thegeneral topic. The reader who desires more discussion on the topic of rub can refer to thecorresponding Machine Library Malfunction Diagnosis article [13]. The main pointsconcerning rub will be summarized below and particulars about pumps will be noted.Also noted will be the pump components that are especially vulnerable to damage by rub.

Definition of RubRub is an undesired contact between a rotating and stationary part. Normally,

bearings serve the purpose of separatingthe rotating part of a machine from thestationary part. When machine parts moveto a position where contact can occur atplaces other than the bearings, the parts“rub” on each other, hence its name.Machines with rub can suffer seriousdamage because the rubs create stresses forwhich the machine was not designed.

Rub contact can occur in the radialdirection or in the axial direction, or in acombination of both (Figure 3.10). Thissection will be concerned primarily withradial rub.

Rubs can force contact between therotor and stator that lasts for just a fractionof the total time required to complete a full

Axial Rub

Radial Rub

Combination Rub

Figure 3.10 Rub can occur in the radial direction,the axial direction, or a combination of both.

CircularOrbit

Figure 3.12 Full Annular Rub (1X, forward).The rotor maintains contact with the clearancesurface of the stator throughout the entirevibration cycle.

vibration cycle or it can force contact that

ClearanceCenter

AveragePosition

ClearanceBoundary

DynamicPosition(Orbit)

Rub

Figure 3.11 Partial Rub. The dynamic motion ofthe shaft centerline (orbit) is added to the averageshaft centerline position. When the totaldisplacement exceeds the allowable clearance,rub occurs.

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lasts for the entire duration of the vibration cycle. Thus the duration, or dwell time, of therub contact can vary significantly. Rubs that force contact between rotor and stator foronly part of the vibration cycle are referred to as partial rubs (Figure 3.11). Rubs thatforce contact between rotor and stator throughout the entire vibration cycle are referred toas full annular rubs (Figure 3.12). Partial rubs can be further separated into Normal-Tight and Normal-Loose while full annular rubs can be separated into forward andreverse. These subgroups are explained in reference [13].

Rubs produce significant forces that act on the rotor. The combination of these forceswith the different type of rubs tend to produce unique patterns of vibration characteristics.By observing these patterns, we can identify rub as the malfunction at hand and hopefullyidentify its source as well.

These effects and their symptoms will be described below. However, we must firstlay the groundwork by discussing the causes of rub.

Causes of RubMachines are designed from the outset to prevent unwanted contact from taking

place. Thus, for a rub to occur, something else in the machine must have moved out ofthe design position (or allowable position range) to some position that results in contact.For this reason, rub is almost always a secondary malfunction. There is usually anothermalfunction that is the root cause of the rub.

1. Radial LoadRub can be caused by radial loads such misalignment or sideload (see “Radial Loads”

above). Misalignment can be either internal or external. Internal misalignment affects theposition of the internal parts relative to the rotating shaft. For example, an out of positionseal or diaphragm could cause this. External misalignment affects the position ofmachines relative to each other. External misalignment can produce unwanted loads onthe rotor system and cause the rotor to move out of normal operating positions insideeither or both machines. Sideloads can have effects similar to misalignment. A heavysideload can deflect a shaft sufficiently to force a rub where clearances are tight (seals,packing, wear rings) or to create high fatigue stresses which lead to a shaft crack.

2. High VibrationHigh vibration produces a large amount of dynamic motion of the rotor inside the

machine. This dynamic motion, which is described by an orbit, is in addition to theoperating average shaft centerline position (Figure 3.11). Rub will occur when theinstantaneous position of the rotor exceeds the allowable clearance.

Recall from the preceding “Unbalance” section that, among other sources, highvibration can be caused by hydraulic or mass unbalance. In one case history, hydraulicunbalance was the root cause of a rub between the impeller vane tips and casing of avertical slurry pump [6].

3. Axial ThrustThe normal flow of liquid through a centrifugal pump creates high axial thrust forces.

Centrifugal pump designs include a wide variety of features and devices to accommodatethese forces [1]. However, the axial thrust balancing within a pump can fail for anynumber of reasons or external factors may create axial thrust for which the machine wasnot designed. For example, a 90° piping elbow situated too close to the inlet of a doublesuction pump can upset flow to the suction and cause one suction eye to receive more

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flow than the other. The uneven flow between the two suction eyes will create unevenpressure on the two sides of the impeller with a resulting axial thrust force.

4. Locked SealsSometimes, a floating seal may lock up. If the seal locks up, the rotor may rub on the

seal.

Effects of Rub on the Rotor SystemThe general concept that rub is an unintended contact between rotating and stationary

parts is quite straightforward. However, the effects of rub upon the rotor system and themechanisms by which these effects produce the vibrations characteristic of rub are fairlycomplicated. These mechanisms and their effect on vibration are discussed at length inreference [13]. The reader who wishes greater depth of discussion can refer to thatarticle.

Vibration Characteristics of Rub1. Changes in 1X VibrationSteady State: At steady state, rub will

produce changes in 1X vibration becauseof rotational energy transfer to lateralvibration energy and because of changesin stiffness. A light rub is more likely toincrease 1X vibration amplitude, whileheavy rub can severely constrain the rotorand reduce 1X vibration. Heavy rub canalso result in more energy transfer to themachine casing, causing an increase of1X vibration on the casing.

If the rotor system is operating near abalance resonance, the 1X vibrationamplitude can increase or decreasedepending on which side of the resonancethe machine is operating at. Theresonance is moved to a different speedbecause of the rub-induced stiffnesschange.

Thermal bow effects due to rub canproduce changes in the amplitude orphase of the 1X response vector.Occasionally, these changes can becontinuous over time (Figure 3.13).

Startup and Shutdown: Duringstartup and shutdown, rub-inducedchanges of the balance resonance speedcan produce changes in observedbehavior through the resonance (Figure3.14). For this reason, it is always good tohave reference startup and shutdown

0° 270°

180°90°

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**

* ** * *

* **

*

******

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**

**********

* ***

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***

* * ** *

*

****

****

*****

*** *

**** ** * *

* * **

*

*****

* * * ** * *** ****

*

******

03:30:0803:34:02

3.0 mil pp Full Scale

* **

Figure 3.13 Steady synchronous rub at an oil seal ina thrust bearing box. The 1X response changescontinuously, completing one revolution of the polarplot in about four minutes.

Speed

Phas

eAm

p

RubStartup

NoRub

RubShutdown

Figure 3.14 Startup and shutdown partial(Normal-Tight) rub behavior. 1X Bode plotcomparisons of experimental startup andshutdown data for no rub, light radial rub duringstartup, and light radial rub during shutdown.

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Rub

Figure 3.15 Full spectrum cascade plot showing a radial rub during startup. The first balanceresonance occurs at about 1800 rpm (horizontal red line). The rub initiates at about 4400 rpm duringthe higher vibration associated with the second balance resonance (blue line). Note that the rubvibration tracks the 1/2X order line and has significant reverse components. Also, at the onset of1/2X, the rotor speed is twice the rub-modified first balance resonance frequency at about 2200 cpm.The inset displays eight shaft revolutions of the direct orbit at 4400 rpm. Note the locked Keyphasordots that indicate that the vibration frequency is a pure integer ratio (1/2X in this case).

Bode and polar plots available for reference.2. Subsynchronous VibrationSubsynchronous vibration, if present at all, will usually be small due to the heavy

damping of the pumped liquid. Subsynchronous vibration may not be present becausethe requirements for ½X are that running speed be more than 2 times the rub-modifiednatural frequency, more than 3 times the same rub-modified natural frequency for 1/3Xand so forth for each subharmonic. Thus the rotative speed must be high relative to thecritical speed (balance resonance) and this is not often the case. If this requirement is metand subsynchronous vibration is present, then amplitudes are usually low because of theheavy damping provided by the pumped liquid.

If a rub does produce subsynchronous vibration, it will follow the rules laid out inreference [13]. These subsynchronous frequency components will be pure integralfractions of running speed and they will follow the rules for Normal-Tight and Normal-Loose partial rubs. The full spectrum in Figure 3.15 shows subsynchronous behavior

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typical of a partial rub. The reverse frequency components are strong indicators of apartial rub.

Note that it is not possible to absolutely verify that a vibration frequency is a pureinteger ratio by using spectrum. There is always some uncertainty in the displayedfrequency on a spectrum plot because of the limited resolution of the spectrum. A direct(unfiltered) orbit with Keyphasor dot display should be used to verify the integerrelationship.

3. Supersynchronous and Reverse Precession VibrationRubs, particularly partial rubs, can

produce sharp changes in the rotortrajectory as the rotor rebounds from thecontact surface. Sharp changes in directionwill produce harmonic frequencies onspectrum plots. For example, if thevibration frequency is predominately 1X,then it is possible to see 2X, 3X, etc.harmonics in the spectrum (Figure 3.16).In the unlikely event that the predominantrub-induced vibration frequency is 1/2X, itis possible to see 1X (as a mixture ofnormal 1X rotor response and the harmonicof 1/2X), 3/2X, 2X, 5/2X, etc.

Reverse components are also oftenpresent in full spectrum plots. Because therub usually involves tangential frictionforces in the reverse direction, fullspectrum plots with supersynchronousvibration will show significant reverse precefrequency.

4. Average Shaft Centerline Position ChRub can produce a dramatic change in the

the change, the average shaft centerline ppronounced effect for light rub, less so for hcenterline position during startup, shutdosymptomatic of a rub.

5. OrbitsDirect (unfiltered) orbits should be exam

activity that may be taking place in the machinoted. For example, the sharp changes in trajthe orbit shape.

Very importantly, only direct orbits with subsynchronous vibration is a pure integer rKeyphasor dots yields the denominator of a frdots could indicate 1/2X, 3/2X, 5/2X and so integer ratio, then the Keyphasor dots will r

1/3X

2/3X

Figure 3.16 Full spectrum of rub showingharmonics. Rub is producing 1/3X, with aharmonic at 2/3X. The 1X line probably containsboth 1X rotor and a harmonic of the 1/3X. 4/3X,5/3X, etc., supersynchronous harmonics are alsovisible. Note that significant reverse componentsexist for many of the harmonics.

51

ssion components at the supersynchronous

anges trajectory, or orbit of the rotor. Because ofosition can change. This can be a veryeavy rub. Sudden changes in average shaftwn, or steady state operation can be

ined and correlated with any other unusualne. Changes in direct orbit shape should beectory produced by rubs will be apparent in

Keyphasor dots can be used to verify thatatio. On an orbit, the number of displayedequency ratio. For example, two Keyphasoron. If the frequency really is locked to anemain locked in place through subsequent

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vibration cycles. If they move in position steadily, then the frequency is not a pure integerratio.

Vibration that consists of a mixture of 1X and rub-induced 1/2X can produce orbitswith complicated shapes (Figure 3.15). This orbit shows the path of the shaft centerlinefor eight shaft revolutions. The two stationary sets of Keyphasor dots show the vibrationto be pure 1/2X.

6. Loss of EfficiencyRub can cause extreme wear of contacting parts. Seals and wear rings can be

especially vulnerable, and, because machine efficiency often depends on tight clearances,wear at these interfaces will usually result in degraded operating efficiency.

Machines with a significant loss of efficiency should be carefully inspected forevidence of wipes at seals, bearings and wear rings. Look for discoloration of parts due tohigh temperature, scratched or smeared bearing babbitt, and damaged turbine blades,compressor blades, or pump impellers.

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Shaft CrackCentrifugal pumps are vulnerable to cracked shafts for a variety of reasons. The

references cited below contained several examples of pumps that have failed due tocracked shafts [3, 5].

The problems created by cracked or totally broken pump shafts are not difficult toimagine. The consequences of a broken shaft can range from inconvenient and costly tocatastrophic and dangerous depending on the circumstances under which the failureoccurs.

Definition of Shaft CrackA shaft crack can be thought of as a slowly growing fracture of the rotor. If

undetected in an operating machine, a crack (also called a fatigue crack) will grow overtime until the remaining, reduced cross section of the rotor is unable to withstand thestatic or dynamic loads that are applied to it. When this happens, the remaining rotorsection will fail in a fast brittle fracture mode. The sudden failure will release the largeamount of energy that is stored in the rotating system, and the rotor will fly apart. Shaftfractures have caused machine parts to penetrate the machine casing and even penetratebuilding walls. Damage due to this kind of failure is catastrophic and can cause seriousinjury or death to anyone unfortunate enough to be standing near the machine at themoment of failure. Obviously, shaft crack detection is a very serious matter, andmachines that are suspected of having a crack must be treated with the utmost respect.

Shaft cracks begin in regions of high local stress. Shafts are subjected to large-scalestresses due to static or dynamic bending and torsional twisting, static radial loads, orresidual stresses from heat treatment, welding, or machining operations. These larger-scale stresses can be concentrated by geometric factors such as step changes in shaftdiameter, shrink fits, keyways, drilled holes, or other discontinuities.

Further stress concentration can occur at the microstructure level where surfacemachining imperfections, chemical surface damage, or material discontinuities (such asproduced by slag inclusions or chemical impurities) can produce high, local stressconcentrations. All of these stresses combine to produce a local stress field that changeswith time (i.e., with shaft rotation). The end result can produce a small local region wherestresses exceed the maximum that the material can withstand, and a microcrack will formin the material. Because shaft bending tends to produce the highest stresses at the outersurface, shaft cracks usually, but not always, start at or near the outer surface. Sometimes,because of chemical or other processing problems in the rotor billet, a microcrack mayexist inside the shaft before it is put into service.

Shafts, because of their rotation, are subject to periodically changing, or cyclical,stresses and can fail even though the actual maximum stresses remain well below theyield strength of the shaft material. Failures that occur via cyclical, or reversing, stressesare referred to as fatigue failures. Shafts can encounter reversing stresses for a number ofreasons, Figure 3.17 illustrates just one of these. If a rotor orbits about the center of therotor system in pure 1X precession, the stress at any particular outer fiber will see nochange in stress. However, if the rotor is offset from the axis of the rotor system(typically the case because of a radial process load or gravity), then rotor outer fibers willsee a 1X variation in stress. In addition, a 1X elliptical orbit (which is also typical)

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produces 2X stress cycling. Thus, evenunder normal, 1X operation, real rotors livein a complicated stress environment thatcontains a mixture of 1X and 2X stresscycling. Any sub- or supersynchronousvibration that may be present will producean additional complicated pattern of cyclicstresses in the shaft.

Once initiated, and if sufficiently highcyclic, or alternating, stresses are present,the crack tip will slowly propagate in adirection perpendicular to the orientation ofthe local maximum tensile stress at thecrack tip. The orientation of this stress fieldis affected by the type of stress (bending ortorsional) and by any geometric factors. Ifa rotor is subjected only to simple bendingstresses, then the stress field will beoriented along the long axis of the rotor,and the crack will propagate directly intoand circumferentially across the rotor sectiproduce a crack that is oriented at 45° relativeinto the rotor, but the crack will tend to fosystems, the stress field usually contains aBending stress is usually the dominantcomponent; thus, the crack will usuallypropagate into the rotor more or less as atransverse crack. However, other crackgeometries are possible.

As the crack propagates, less and lessmaterial is available to transmit loads in therotor shaft, and the local stress across theremaining shaft material becomes higherand higher. At some point, the section willbecome so small that, during the next loadapplication, the local stress intensity willexceed the fracture toughness of thematerial. The fracture toughness is ameasure of the material’s resistance to fastfracture and is a function of the alloy, heattreatment, the material temperature, and therate of loading of the shaft. When thefracture toughness is exceeded, theremaining section will undergo a fastbrittle fracture, and the rotor will break inhalf.

Rotor Centered

Rotor Offset WithElliptical Orbit

1 20

0

Tens

ion

0

Shaft Revolutions

Tens

ion

1 20

Com

p

Figure 3.17 An example of outer fiber stressvariation for a rotor in simple bending. When therotor moves about the center of the system (top)in a 1X circular orbit, the stress is constant. Whenthe rotor is displaced from the system center in a1X elliptical orbit (bottom), the rotor seesvariable stress with a mixture of 1X (from thedisplacement) and 2X (from the ellipticity).

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TransverseCrack

TorsionCrack

Figure 3.18 Transverse and Torsional Cracks.A transverse crack results from pure bendingstress in the shaft and propagates directly into theshaft. A crack resulting from pure torsional stressforms a spiral at 45° to the long axis of the shaft.Most shafts contain a mixture of bending andtorsion stress. The local stress field at the cracktip, which can be influenced by local geometry,determines crack propagation direction. Thecrack tip propagates perpendicular to thedirection of the maximum local tensile stress.

on (Figure 2). Pure torsional stress will to the long axis. The crack will propagaterm a spiral on the shaft surface. In rotor mixture of bending and torsional stress.

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a b c

Figure 3.20 2x snapping action of crackedshaft. The effect of asymmetric rotor stiffnesscan be demonstrated with a simply supportedruler. In the figure, identical downward loads (redarrows) are applied to a ruler in three positions.The responses are shown as blue arrows. Theruler has minimum stiffness and largestdeflection at position (a). It has maximumstiffness and smallest deflection at position (b).At intermediate position (c), the ruler has aperpendicular (quadrature) component ofdeflection. If the ruler is rotating with a similarunidirectional load, a snapping motion will beseen two times per revolution (2X).

Rotor shafts are usually manufactured out of materials with high fracture toughness.Rotor shaft cracks have exceeded 90% of the shaft cross sectional area before finalfracture, although certainly one should not depend on this. It is not an easy matter todetermine crack size in a rotating machine, and any machine suspected of having a shaftcrack should be shut down as soon as possible.

Effects of Shaft Crack upon Rotor System1. Reduction of Shaft StiffnessShaft cracks, like other malfunctions, have their own unique effects upon a rotor

system. The first of these is the overall reduction of shaft stiffness. This occurs in arelatively straightforward manner and can be visualized by comparing the stiffness of athin shaft with that of a thicker shaft. Given that all other things are equal, the thinnershaft will bend more easily than the thicker one because it has less stiffness, i.e., lesscross sectional area moment of inertia. A crack has the same effect because it reduces thecross sectional area moment of inertia of the shaft. Less cross sectional area moment ofinertia, or stiffness, means that the shaft will now show greater deflection in response tothe forces which act on the rotor. This leads to a change in rotor bow and 1X behavior.

2. Asymmetric Shaft StiffnessAnother effect of cracks upon pump shafts is asymmetric stiffness. Asymmetric

stiffness means that the shaft is stiffer in one direction than in another. Considering thefact that cracks generally start at the surface and work inward, stiffness will be reducedmore in the direction of crack growth than in the direction perpendicular to crack growth(Figure 3.19).

The significance of crack-induced asymmetric stiffness is that it can produce a

LowerStiffness

HigherStiffness

CrackRemaining

Shaft Section

Figure 3.19 Asymmetric stiffness of a crackedshaft. As a crack grows, the remaining shaftsection shape becomes asymmetric. Thisproduces an asymmetric rotor shaft stiffness thatrotates with the rotor.

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Fpmaswae

Fsmaf

snapping action of the rotor whenever a strong sideload is present. Since sideloads areusually present in centrifugal pumps to one degree on another (refer to “Casing”), so isthe snapping effect associated with shaft crack. The snapping action produced bysideload acting on a cracked shaft occurs twice per shaft revolution and, hence, shows upas a 2X frequency component (Figure 3.20).

Both of these effects are explained in more detail in the corresponding MachineLibrary Release 2 Malfunction Diagnosis article and the reader who desires fullerexplanation of these effects can refer there. These effects help explain the vibrationcharacteristics that follow in the next section.

Vibration Characteristics of Shaft Crack1. The First Rule of Crack Detection (1X)The first rule has to do with 1X filtered

vibration: If a rotor is cracked, it is verylikely to be bowed. And that bow is likely tochange over time. A change in rotor bowwill change the effective location andmagnitude of the heavy spot, which willproduce a change in 1X rotor response.Thus, continuous changes in 1X amplitudeand/or phase are the best primary indicatorof a shaft crack. As the crack grows and theassociated bow develops, 1X amplitude andphase will change in such a way as toproduce a non-repeating pattern on a Bodeor polar plot over time (Figure 3.21). Thetime scale of this change can range frommonths to weeks in the early stages of crackgrowth, to weeks to days as the rotor beginsto seriously weaken, and to hours as therotor nears catastrophic failure.

As failure nears, 1X vibration amplitudewill usually increase rapidly. At this point,1X vibration is likely to be the dominantsource of vibration in the system, so overalldirect vibration will also increase rapidly.Thus, steady increases in unfiltered peak-to-peak vibration over time should be takenvery seriously and investigated.

This increase in 1X vibration is not anabsolute - it can also decrease (Figure 3.22).As a crack propagates across the shaft,reduced stiffness may shift balanceresonances downward to a lower speed.Depending on the mode shape and thelocation of the crack, some resonances (ormodes) of a rotor could be affected more

1X Uncomp

1 Mar

1800 RPMGenerator Outboard X

Time

Am

plit

ude

Phas

e La

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1 Apr 1 May 1 Jun 1 Jul 1 Aug

igure 3.21 Example 1X APHT plot showingossible 1X vibration changes due to a crack. Theachine runs at a constant speed, and amplitude

nd phase change slowly over time. Near the end,haft stiffness will drop rapidly as the crackeakens the section (red arrow), and vibration

mplitude will increase rapidly. The details ofach machine’s behavior are different.

56

Speed

Phas

eAm

p

1 2 3 4

HighStiffnessLow

Stiffness

igure 3.22 As a crack propagates, rotor shafttiffness will decrease, and resonant speeds mayove downward. Here, the changes in 1X

mplitude and phase that could occur are shownor four different operating speeds.

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VerticalFrequency

Relationship

Rot

or S

peed

(rpm

)

Precession Frequency (kcpm)

Figure 3.23 Full spectrum cascade startup data from a rotor with ashaft cross-section asymmetry and a unidirectional radial load. Asignificant 2X component appears when the running speed (left axis)reaches ½ of the balance resonance speed near 3600 cpm. Note that areverse 2X component is also visible but is smaller than the forwardcomponent. Thus, the 2X component is forward and elliptical.

than others. Also, rotor mode shapes may change, depending on the location of the crackand how it effects the stiffness distribution of the shaft.

Because of the changing bow of the rotor, the amplitude and/or phase of the 1Xfiltered slow roll vector is also likely to change as the crack propagates. Slow roll vectorsshould be compared to historical slow roll data.

Occasionally, a diagnostician may encounter a machine with a “balance problem.”Perhaps the machine had no history of such a problem before. Changing rotor bow due togrowth of a shaft crack will change the location and magnitude of the effective heavyspot of the rotor. If this happens, a previous balance correction may soon be renderedineffective, and the 1X vibration will increase again. If the root problem is a shaft crack,repeated attempts to balance the machine will not solve the problem.

Changes in 1X rotor behavior in resonances are an indication that something haschanged in the rotor system. A significant downward shift in a balance resonance speed isa clear indication that the stiffness of the rotor system has decreased. One then has to askwhy this has happened. A weakening shaft due to a crack is a possibility.

1X vibration is usually very sensitive to the presence of a shaft crack because of therelationship between crack growth and rotor bow, and it can provide significant earlywarning of a crack.

2. The Second Rule of Crack Detection (2X)The second rule has to do with 2X filtered vibration: If a rotor with a crack has a

steady, unidirectional radial load, then a strong 2X response may appear when the rotoris turning at half of abalance resonance speed.In addition, this 2Xcomponent is likely to bepredominately forward(although it may beelliptical). The 2Xsnapping action of therotor produces lateral andtorsional impulses in thesystem. Because theseimpulses occur twice perrevolution, the rotor willrespond at the 2Xfrequency. If a resonanceexists at twice runningspeed, then the 2Xvibration will beamplified. This forms avertical relationship on ahalf or full spectrumcascade plot (Figure 3.23).

Note that a rotor couldpass through such a 2Xspeed relationship during

f=

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0°

**

***

****

***

**

*

270°

180°

90°

501119084840

38353231

3031

293426674359

5541

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****

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180°

90°

270°

0°59616947

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171517561853

8752170

48685008

425525

3236776

ROTN 19 June 28 June

4 mils pp Full Scale

2X Filtered Data Note Phase Change in Leading Direction

Figure 3.24 2X filtered polar plots showing data from two startups of a machine with a rotor crack.The startup on 28 June shows drastically different behavior than the earlier startup on 19 June. On thelater startup 2X amplitude is larger, and some unusual, leading phase behavior is visible.

0 90180270360

10.0 8.0 6.0 4.0 2.0 0.0

1SEP 25SEP 19OCT 12NOV 6DEC

SPEED: 1187 rpm2X Filtered

Phas

e La

gA

mpl

itude

Figure 3.25 2X filtered APHT plot of a ReactorCoolant Pump with a crack. The pump operatedat a constant speed of 1187 rpm. As the shaftweakened, the reduced shaft spring stiffnesscaused a resonance that was originally abovetwice running speed to move down and passthrough twice running speed. Note that, when themachine was shut down, the 2X amplitude wasdecreasing.

startup or shutdown, or a rotor could normally operate at half of a resonance speed.Obviously, it is less likely that such a relationship would exist at normal operating speed.For that reason, and because of the additional requirement that a unidirectional radial loadbe present, a crack may or may not produce significant 2X vibration at running speed.

Experience with other machines types has shown that 2X vibration does not appearwhen operating at design speed in about 75% of shaft cracks. However, this is one of theinstances where a common machine malfunction can manifest itself differently incentrifugal pumps than in other types of machines. While 2X is not always present for

the reasons given in the precedingparagraph, 2X vibration is about as commonin centrifugal pumps as are changes in 1Xvibration. This is because centrifugal pumpsusually experience sideloading which, alongwith asymmetric shaft stiffness, creates thetwice per revolution snapping action that isresponsible for 2X vibration.

Like 1X vibration, 2X vibrationamplitude and/or phase can change as thecrack propagates through the rotor shaft.Startup and shutdown 2X Bode and polarplots should be examined for any evidenceof change (Figure 3.24). Also, 2X amplitudeand phase should be trended during steadystate operation. In one case, a reactor coolantpump developed a crack while the pump wasoperating at a constant speed. As the crackpropagated, the rotor shaft stiffness droppedso much that a resonance that originally

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existed above twice the operating speed moved down in frequency and passed completelythrough the 2X frequency before the pump was finally shut down (Figure 3.25).

Other Malfunctions With Similar Symptoms1. 1X BehaviorMany other malfunctions can produce a change in 1X rotor response.A loose bearing support or soft foot can cause a change in 1X vibration. Usually, but

not always, this is manifested as an increase in 1X vibration amplitude. This is the kind ofmalfunction that could develop over time with a slow increase in 1X vibration. Because itmimics the behavior of a shaft crack, it can be very difficult to determine the root cause.If casing measurements are available, an increase in casing vibration with little or noincrease in shaft relative might suggest a soft foot problem, while an increase in shaftrelative with little increase in casing might suggest a crack. But, there are no firm ruleshere.

Thermal growth and subsequent changes in alignment can affect the rotor bearingstiffness and produce changes in 1X vibration. Thermal bow of a rotor such as agenerator can also produce a similar change, as could an alignment change. Thesechanges in vibration should stabilize once the machine reaches thermal equilibrium atsteady speed and load.

Rub can cause changes in both 1X and 2X vibration. These changes can be sudden,occurring at operating speed, or the changes can show up as changes in transient behaviorduring startup or shutdown. Rub can disappear if the part in contact wear away (this canhappen in seals). Or, if the rub is severe, rub contact may be maintained for aconsiderable time. However, rub is not as likely to produce a steadily increasing 1Xvibration level over a long period of time.

A loose rotating part can produce changes in 1X response. If a part moves to adifferent angular or axial position on the rotor, the resulting total unbalance of the rotor islikely to change, and the 1X amplitude and/or phase will change accordingly. Loose partscan shift occasionally, producing stepwise changes in 1X response, or they can shiftcontinuously, producing a continuously changing response. Continuously moving partswill tend to produce a cyclic, repeating behavior on a polar or APHT plot. A loose part isnot likely to produce a steady, long-term increase in 1X vibration amplitude.

Clogged debris in an impeller can produce significant differences in heavy spotlocation in a machine. This will produce corresponding changes in 1X vibration responseand cause a machine to go out of balance.

A locked gear coupling can also produce a sudden step change in 1X vibration.The key to crack identification is to realize that a developing crack is likely to

produce a steady and accelerating increase in 1X vibration amplitude over time as theshaft stiffness decreases. While some malfunctions will produce periodic changes in 1Xvibration amplitude and/or phase, shaft cracks will tend to produce non-repeating patternson polar and APHT trend plots, with the 1X amplitude trending to ever higher levels.

2. 2X BehaviorNonlinearities in rotor system stiffness can cause harmonics (2X, 3X, etc.) of running

speed to appear in spectra. Nonlinear stiffness can be caused by high eccentricity ratios influid-film bearings or by rub impacting. Also, coupling problems can produce 2Xvibration.

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Pt. B Pt. C Impeller

Bearing Journals

Pump Shaft

Sideload

C

B

Fig. 3.26 Typical Points of Shaft Crack on Overhung Pumps. Shaft deflects under theheavy hydraulic sideload. High deflection causes the shaft to undergo one cycle of alternatingtensile and compressive stress per each shaft rotation. These fatigue stresses are maximum atthe shaft surface and are further concentrated (or multiplied) at the step changes in diameter atshoulders located at points B and C. Cracks can initiate in these regions of high fatigue stress.(After figure from reference [7].)

If any source of 2X vibration exists in a machine, it will be available to excite aresonance at half of a balance resonance speed. Thus, the presence of 2X at half aresonance, while suspicious, is not in and of itself confirmation of a crack.

Causes of Shaft Crack in PumpsRecall that cracks occur in regions of high localized stress. Pump shafts are subject to

all the same factors that create and concentrate stress in other machine types. In addition,pump shafts are equally, if not more, subject to high fatigue stresses that result in fatiguefailure. Some of the common sources of stress are listed below.

1. Stress ConcentratorsStep changes in diameter such as shoulders will concentrate stress. Overhung pump

shafts often have shoulders situated in the region where the bending stress fromsideloading is high (Figure 3.26, Pts. B and C). Pump designers will calculate themaximum stress expected in operation and try to minimize it with fillets and other gooddesign practices. However, actual operating conditions may present higher stresses thosefor which the pump shaft was designed. For example, one reference [7] saw “…severalpumps break shafts at point C or B because the pump was designed to run with packingfor support on long overhangs (large C & B dimension) and the pump was later changedto run with mechanical seals, i.e., no support. The seal performance was poor alsobecause of excessive deflection.” (Figure 3.26)

2. Residual StressResidual stresses may be leftover from the manufacturing process or may be

unwittingly created by well-intentioned maintenance practices. For example, a high-pressure boiler feedpump shaft that bent (for reasons not stated) was straightened bypeening with a blunt nose chisel in order to return it to service. The shaft developed acrack over the course of about a year which became evident when the stuffing boxesbegan to leak excessively (as rotor bow increased – recall the first rule of shaft crack).While not a foregone conclusion that the peening alone caused the shaft crack, it is quite

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possible a main contributor since the resulting compressive stresses may have exceededthe maximum allowable [3].

3. Radial LoadsSideloads: As noted previously, centrifugal pumps are usually subjected to high

sideloads which increase as the pump is operated further away from its BEP. Pumps areoften operated further off of design capacity than designers originally intended. This cansubject pump shafts to very high levels of fatigue stress.

Misalignment: Misalignment is also a common of source of excessive radial loadresulting in high fatigue stress [1].

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SpringForce

RadialLoad

PressureWedge

TangentialForce

Figure 3.27. The circulating fluid in a bearingor seal forms a pressure wedge when the rotoris displaced from the center (left). This reactionforce can be separated into a tangential force (red)and a spring force (green). The spring force triesto move the rotor back toward the center of thebearing, but the tangential force tries to move therotor tangentially. The tangential force isultimately responsible for destabilizing the rotor.

Fluid-Induced InstabilityDefinitionFluid-induced instability is a large amplitude, usually subsynchronous vibration of a

rotor that is caused by rotor interaction with a surrounding fluid. The term “instability” issomewhat of a misnomer. When a rotor operates in fluid-induced instability, it is actuallyoperating in a stable limit cycle of high vibration. But the rotor is unstable in the sensethat it is operating outside desired operational limits.

The large amplitude, subsynchronous vibration can cause rotor-to-stator rubs on seals,bearings, impellers, or other rotor parts. The vibration can also produce large-amplitudealternating stresses in the rotor, creating a fatigue environment that could result in a shaftcrack. In addition, the bearing surface is subject to alternating stresses that can lead tofatigue failure of the babbitt.

Fluid-induced instability is a potentially damaging operating condition that should beavoided.

Cause of Fluid-Induced InstabilityWhen a fluid, either liquid or gas, is trapped in a gap between two, concentric

cylinders, and one is rotating relative to theother, the fluid is set into motion aroundthe gap (Figure 3.27). This situation existsin fully lubricated (360° lubricated) fluid-film bearings, in seals, around impellers inpumps, or when any part of a rotor iscompletely surrounded by fluid trappedbetween the rotor and the stator. In thissection, we will talk primarily about fluid-film bearings of basic cylindrical shape.However, it should be understood thateverything written here about bearings alsoapplies to seals, pump impellers, and anyother region in a machine where a liquid orgas is trapped in a small clearance betweena rotor and a stator.

When a rotor moves away from thecenter of a bearing, the converging fluidforms a pressure wedge (Figure 3.27, left).The pressure profile creates a force that canbe separated into two components. A directcomponent or spring force, Fs, exists that acts like a spring and points back toward thecenter of the bearing:

rKFs = (1)

where K is the effective spring constant of the bearing at that eccentricity ratio, and r isthe distance from the center of the bearing.

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At the same time, a quadrature component acts in a tangential direction in the samesense as rotor rotation. It turns out that this tangential force, Ft, is a function of bearingdamping, fluid circulation, rotor speed, and distance from the equilibrium position:

rjDFt Ωλ= (2)

where D is the bearing (or seal, impeller, etc.) damping, λ (lambda) is the FluidCircumferential Average Velocity Ratio, Ω (capital omega) is the angular velocity of therotor (the speed of the rotor in radians/sec), and r is the distance from the center of thebearing.

The j is 1− . Practically, all that means is that the action occurs at 90° relative to thespring force, Fs, in the direction of rotor rotation (Figure 3.27, right). (See reference [16]for much more detail.)

What is λ? Put most simply, λ is a measure of the amount of fluid circulation in thebearing. It is defined as the ratio of the average angular velocity of the fluid to the angularvelocity of the rotor. For a plain cylindrical, fully lubricated (360° lubricated) bearing λ istypically a little under ½, around 0.49 or so. But the value of λ can be influenced by thegeometry of the bearing, the rate of end leakage out of the bearing, the eccentricity ratioof the rotor in the bearing, and the presence of any pre- or antiswirling that may exist inthe fluid.

Note that the strength of the tangential force depends not only on the rotation speed,Ω, but also on the strength of fluid circulation around the rotor (λ). It is much stronger (λis much higher) when the rotor is surrounded with fluid (the fluid-film bearing is fully, or360° lubricated). Properly loaded fluid-film bearings are normally only partiallylubricated, and λ is usually small. Thus, properly loaded bearings are unlikely to be asource of very large tangential forces unless the bearing becomes flooded with an excessof lubricant. Note that fluid-film bearings can become unloaded, for example because ofmisalignment, transition to fully lubricated operation, and generate high tangential forces.

The spring force, Fs, acts to stabilize the rotor because it pushes the rotor back towardthe center of the bearing. However, the tangential force, Ft, acts to destabilize the rotorby pushing it at a right angle (i.e., tangentially) to the bearing center. If conditions areright, the tangential force will drive the rotor in a large amplitude, forward, circular orbitconstrained only by the stiffness of the surrounding stationary cylinder (e.g., bearing,seal, pump casing). When this occurs, the rotor is undergoing a fluid-induced instability(whirl or whip).

Reference [15] explains these conditions in detail and describes the differencebetween whirl and whip. Additional detail on this subject is beyond the scope of thispaper. The reader who desires more detail is referred to that article.

At this point, the primary concept to note is that the “ingredients” for fluid-inducedinstability can exist in several regions of a centrifugal pump. This includes impellers, therotors in magnetic drive and canned motor pumps, as well as in bearings and seals.

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Vibration Characteristics of Fluid-Induced Instability1. Subsynchronous VibrationThe primary symptom of fluid-induced instability is forward, subsynchronous

vibration. The frequency of the subsynchronous vibration due to oil whirl is usually lessthan 0.5X (Figures 3.28 and 3.29). For pumping whirl (whirl originating in the pumpedliquid surrounding the impeller), this frequency can occur in the range of 0.7X to 0.9X

Frequency (kcpm)

Rot

or S

peed

(rpm

)

Whi

rl

Whip

High EccentricityNatural Frequency

Low EccentricityNatural Frequency

Thre

shol

dof

Sta

bilit

y

Figure 3.28 Full spectrum cascade plot of a rotor system startup. The rotor system starts into fluid-induced instability (in whirl) at about 2400 rpm, the Threshold of Stability. At this time, subsynchronous,forward precession begins at a frequency near 0.475X. The initial whirl frequency is about 1300 cpm, whichis the low eccentricity natural frequency of the rotor system. As speed increases, the whirl orbit becomeslarger, the bearing becomes stiffer, and the rotor system natural frequency shifts to a higher frequency.Thus, the whirl tracks at a sub multiple of running speed. At about 2900 rpm, the high 1X rotor vibrationassociated with a balance resonance causes the rotor to operate at a high dynamic eccentricity ratio. Theresulting higher bearing stiffness pushes the Threshold of Stability temporarily above running speed, andthe fluid-induced instability disappears. After the resonance, 1X vibration declines, the orbit diameterdecreases, the bearings stiffness decreases, and the Threshold of Stability once again falls below runningspeed; thus the fluid-induced instability reappears. When the rotor dynamic motion reaches higheccentricity, the rotor shaft becomes the weakest spring in the system, and the instability frequency locks into the high eccentricity natural frequency in whip. The orbit inset shows the orbit of the rotor inside thebearing in whirl, and the magenta circle shows the approximate bearing boundary. At this dynamiceccentricity ratio (about 0.6), the bearing controls the spring stiffness of the rotor system (see Figure 3). Thepair of Keyphasor dots are shifting slowly in a direction opposite to rotation. This indicates that thefrequency of vibration is a little less than 1/2X.

64

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due to the increasedfluid circumferentialaverage velocitygenerated by theimpeller [4], [9]. Thefrequency of thesubsynchronous whirlvibration is related tothe fluid swirling rate,lambda (λ), of the fluidcausing the instability.In whip, however, thefrequency of vibrationwill lock to a rotorsystem bending mode(Figure 3.29). Thesubsynchronous whipfrequency can rangefrom 0.3X to 0.8X orhigher depending uponthe fluid circumferentialaverage velocity ratio(λ) of the fluid causingthe problem.

Unlike rub, fluid-induced instabilityalmost never produces apure integer ratiovibration frequencysuch as 1/2X, 2/3X, 1/4X, fraction frequencies. Howdown or if the large amplinduced instability can lock

The subsynchronous vforward (Figures 3.28 andand fluid-induced instabilicomponents at the subsync

During a startup or shuspeed at some sub multifrequency (Figure 3.29). Aappear without any whirl.

Fluid-induced instabilisystem (usually the lowest will appear during startup the rotor is supercritically (

Frequency (kcpm)

Rot

or S

peed

(rpm

)

Whip

FirstBalance

Resonance

Figure 3.29 A rotor system can enter fluid-induced instability whipdirectly without encountering whirl first. In this case, the rotor operatesat a high eccentricity ratio within the bearing, and the bearing stiffness ismuch higher than the shaft stiffness. The rotor enters whip in a bendingmode that corresponds to the high eccentricity natural frequency. Thefirst balance resonance for this mode can be seen at approximately 2200cpm. Harmonics of the whip frequency are also visible. The whip orbit isalso shown inside the magenta bearing clearance. Note the jumble ofKeyphasor dots and the very high dynamic eccentricity ratio of about0.9. Shaft stiffness is the weak spring (Figure 3); thus, the naturalfrequency cannot be modified and the subsynchronous frequencyremains constant.

65

1/3X, etc. Instead, fluid-induced instability produces irrationalever, if the lubricating film between rotor and stator breaksitude instability vibration causes a rub elsewhere, then fluid- to an integer ratio.ibration caused by fluid-induced instability is almost purely 3.29). This is a very useful way to discriminate between rubty as a root cause. Rub tends to produce significant reversehronous frequency.tdown, whirl due to fluid-induced instability will track runningple (Figure 3.28), while whip tends to lock to a constants can be seen in the figure, it is possible for whip to suddenly

ty is always associated with a natural frequency of the rotormode). Often the balance resonance associated with that modeas 1X vibration (Figure 3.29). However, if the lowest mode ofover) damped (as can happen with rigid body modes), then the

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rotor will not have a resonance on that mode, and the 1X vibration associated with themode will not be visible during startup. This is the case for the machine in Figure 3.28.

2. OrbitsIf the vibration at the measurement plane is dominated by fluid-induced instability,

then the direct, unfiltered orbit will be predominately forward and circular (Figure 3.28).Orbits that are filtered to the instability frequency will always be approximately circularand forward.

The behavior of the Keyphasor dots will depend on the relationship of thesubsynchronous frequency to running speed (the Keyphasor trigger frequency). Ingeneral, the number of Keyphasor dots visible is related to the denominator of the nearestsubsynchronous integer ratio. For subsynchronous frequencies near 1/2X, two Keyphasordots will be visible. If the subsynchronous frequency is slightly below 1/2X, then theKeyphasor dots will slowly drift in a direction opposite to rotation. If the subsynchronousfrequency is slightly above 1/2X, then the Keyphasor dots will slowly drift in the samedirection as rotation. Vibration near 1/3X will produce a set of three Keyphasor dots inthe orbit that behave in a similar way. Vibration near 2/5X (0.4X) will produce an orbitwith 5 Keyphasor dots.

When the subsynchronous vibration is not near an integer ratio, the Keyphasor dotswill tend to form a chaotic pattern consisting of great many dots (Figure 3.29).

Note that, under the right circumstances, rub will produce subsynchronous vibrationat a pure integer ratio with locked Keyphasor dots. These dots will not drift around theorbit with time and will tend to stay in thesame location. Because rub producesinteger ratio subsynchronous vibrationfrequencies (such as 1/2X), Keyphasordots from a subsynchronous rub orbit willform a locked integer set. This is a verypowerful tool for discriminating betweenfluid-induced instability and rub. LockedKeyphasor dots imply rub, while movingKeyphasor dots imply fluid-inducedinstability.

Whip orbits, because of the lowersubsynchronous frequencies at which itusually occurs, are more likely to showchaotic Keyphasor dot behavior thanwhirl orbits.

If the vibration at the measurementplane contains a mixture of 1X andsubsynchronous vibration, then the orbitwill be more complex in shape. Thesubsynchronous vibration will cause theorbit to continually change shape, but themotion of the Keyphasor dots (for frequencies close to an integer multiple) will still tendto migrate in a small circle (Figure 3.30).

Fluid Instability Rub

Figure 3.30 Direct orbits showing a mixture of 1Xand subsynchronous vibration for eight shaftrevolutions. The fluid-induced instability frequencyis slightly less than 1/2X in whirl (the orbit is from alocation some distance from the source), while therub frequency is exactly 1/2X. In the instability orbitthe Keyphasor dots slowly migrate against rotation(black arrows) in a circular path (red), while the ruborbit dots are locked in place

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Whip

LowSpeed

HighSpeed

RadialLoad

Direction

Figure 3.31 Average shaft centerline plot showing thetransition from stable behavior (black) to fluid-inducedinstability whip (blue) inside the bearing for the datashown in Figure 6. The dashed circle shows the bearingboundary. As the instability develops, the averageeccentricity ratio in the bearing approaches zero.

500

3600

RadialLoad

Direction

Figure 3.32 Normal shaft centerlineplot from a shutdown of a steamturbine generator. During startup, theproperly loaded shaft centerlinewould start at the 500 rpm point andmove up to the right for X to Y(CCW) rotation. Compare to Fig. 7.

3. Average Shaft Centerline PositionIn classic fluid-induced instability, the journal will move about the center of the

bearing at a subsynchronous frequency in a forward, circular orbit. As the rotor orbitgrows larger in whirl or whip and begins to move around the bearing clearance, theaverage eccentricity ratio will begin to approach zero. That is, the average shaftcenterline position will approach the bearing center (Figure 3.31). Thus, it can be veryuseful to correlate the onset of subsynchronous vibration with movement of the shaftcenterline toward the center of the bearing.

A related issue concerns a potential cause of fluid-induced instability. Machines withfluid-film bearings are usually designed to operate in a partially lubricated condition at ahigh eccentricity ratio position. The shaft centerline plot of a normal machine has atypical behavior (Figure 3.32). If a machine becomes misaligned, then one or morebearings in the machine may become partially unloaded. When this happens, the shaftcenterline operating position will move to an abnormal position closer to the center of thebearing. Operation near the bearing center is more likely to result in full lubrication of therotor journal, causing fluid-induced instability. Thus, the shaft centerline plot can providea clue as to the root cause of the fluid-induced instability that is taking place in themachine.

Corrective Actions for Fluid-Induced Instability1. Reduction of Fluid CirculationThe fluid circulation is what creates the destabilizing tangential force. λ, the Fluid

Circumferential Average Velocity Ratio, is a measure of the strength of the fluid

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Figure 3.33 Bearing geometries that break upfluid circulation in the bearing reduce the valueof λ and promote stability.

circulation. Anything that acts to disrupt fluidstability.

Control of λ can be difficult for an end uthe OEM level. This is commonly done by utsimple cylindrical shapes (Figure 3.33). Tilt pthe pads are not continuous, fluid flow is enhanced.

Antiswirl injection involves injecting woseal in a direction opposite to rotation (Figurethe overall average fluid angular velocity successfully applied in both bearings and seals

2. Proper Loading of Hydrodynamic BeFluid-induced instability often origina

insufficiently loaded. Misalignment can shiftother bearings in the machine. The lightly locloser to the center of the bearing.

If a machine that once ran acceptably nothe shaft centerline plot to see where the rotorrotor is found to be operating in a particulaadjacent bearings are highly loaded, then tmachine should be checked. Correct alignmen

At the design level, fluid-film bearings adequate load. Over designed bearings could r

3. Adjustment of Supply PressureHydrostatic bearings normally operate

condition. In these types of bearings, the sprinfluenced by the lubricant delivery pressure supply pressure will increase the rotor sysinstability.

PressurizedFluid

PressurizedFluid

Figure 3.34 Antiswirl injection involvesinjection of pressurized fluid tangentially into abearing or seal in a direction opposite rotation.The injected fluid disrupts circulation and greatlyreduces λ

68

flow around the clearance will help rotor

ser and can be most easily accomplished atilizing bearing geometries that depart fromad bearings are an example of this. Becausedisrupted in the bearing and stability is

rking fluid tangentially into the bearing or 3.34). The injected fluid acts to slow downand reduce λ. This technique has been and has proven to be very effective.

aringstes in hydrodynamic bearings that are the load from one bearing to one or moreaded bearing will tend to position the rotor

w exhibits fluid-induced instability, check is operating in the bearing clearance. If ther bearing at a low eccentricity ratio whilehe external and internal alignment of thet should result in properly loaded bearings.in a machine should be designed with anesult in a fluid-induced instability problem.

in a fully lubricated (360° lubricated)ing stiffness of the bearing, KB, is stronglyin the bearing. Thus, increasing the bearingtem stiffness, K, and may eliminate the

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Seals act like hydrostatic bearings. The rotor in the seal area is normally completelysurrounded by the working fluid of the seal. Thus, increasing the seal fluid supplypressure may increase the stiffness of the seal and, if the seal is the source of theinstability, eliminate the instability.

Hydrodynamic bearings, on the other hand, normally operate in a partially lubricatedcondition at a relatively high eccentricity ratio. Increasing the lubricant supply pressuremay actually flood the bearing, causing it to operate in a fully lubricated condition. Thisis likely to destabilize the rotor system. If a hydrodynamic bearing is suspected of beingthe source of the fluid-induced instability, then reducing lubricant supply pressure mayeliminate the flooded condition and stop the instability. Obviously, care must be taken toavoid reducing the supply pressure to such a low level that causes damage to the bearing.

4. Adjustment of Lube Oil TemperatureFluid viscosity affects both the bearing stiffness, KB, and the bearing damping, D.

Thus, changing the fluid viscosity may have a significant effect on the fluid-inducedinstability.

It is difficult to predict ahead of time how changes in oil temperature will affect thespeed at which fluid-induced instability occurs (called the Threshold of Stability). Insome cases, a change in oil supply temperature of only a few degrees has produceddramatic changes in the fluid-induced instability behavior of the machine. Furtherexplanation of the effect of lube oil temperature on fluid-induced instability can be foundin Reference [15].

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Structural ResonancesRotors are not the only parts of a pump and its driver that can resonate, other machine

parts such as casings and brackets can resonate as well. In fact, it is not just machineparts that can resonate, any object with properties of elasticity (i.e., spring stiffness) andmass will resonate when excited at its natural frequency. This includes virtually everyobject found in the industrial environments that normally surround centrifugal pumps.The piping, piping support brackets, the pump pedestal and other support structures,roofs, walls, floors, etc. – in other words, any surrounding structure will resonate whenexcited at its natural frequency, hence the term structural resonance.

The vibration of a structural resonance can transmit into a pump and damagebearings, seals, couplings, and other vulnerable pump components. Conversely, vibrationcan originate from within the pump and damage an adjoining structure, although thisprobably less likely since the pump will tend to be more sensitive to vibration thanadjoining structures. In either case, it is important that structural resonances, if present,be identified and corrected if believed to be problematic.

Definition of Structural ResonanceResonance is the peaking of the amplitude of vibration that occurs when a periodic

force excites an object at its natural frequency. Anyone who has heard a rattle in their carhas experienced a structural resonance. The offending rattle occurs because some part isloose enough (that is, its stiffness is lowered) so that it vibrates when the frequency of theexciting forces (for example: engine, vibration of tires on road surface, etc.) match thenatural frequency of the part. That is why the rattle may appear and then disappear withchanges in speed. The rattle is most pronounced when the frequency of the excitingforces are closest to the natural frequency of the loose part. Recall that the naturalfrequency, denoted by ωn, of an object is defined by the equation:

MK

n =ω (1)

where K and M are object’s spring stiffness and mass respectively. This simplerelationship between stiffness and mass explains why stiffening the part by eithertightening or reinforcing it silences the rattle. The stiffened part now has a naturalfrequency that is beyond the range of the frequency of the exciting forces and is thusunable to resonate.

Machines “live” in a complex vibration environment. Vibrations of a wide range offrequencies originate from both inside and outside the pump. The frequencies ofvibrations originating from within the pump typically include 1X, but can also includesubsynchronous and supersynchronous frequencies. Vibration external to the pump canoriginate within adjacent machines and also from adjoining processes (process liquidflowing through a pipe can excite vibrations as well). All of these vibrations of variousfrequency and amplitude combine to make up a complex source of excitation.

Even though common sense tells us that vibrations do not stop at some artificialboundary surrounding the pump but rather will transmit in and out, the interactionbetween a pump and its adjoining structures is sometimes overlooked. The interactivity

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StructuralResonance

18024030036060

120180

10

8

6

4

2

00 1000 2000 3000 4000 5000 6000 7000

Phas

e Lag

(deg

rees

)

Bode Plot, 1X Compensated

Ampl

itude

(mils

pp)

Figure 3.35. Structural resonance on Startup and ShutdownPlots. The structural resonance appears as deviations from thenormally expected ampitude and phase lag curve shapes. Thecircled portions of the curve show increasing amplitude and phaselag. However, the deviations could just as easily have beendecreases.

of a pump with its environment means that there are additional sources of vibration thatneed to be considered when attempting to solve problems of high vibration.

Vibration Characteristics of Structural ResonanceStartup and Shutdown data: Structural resonance can cause vibration amplitude and

phase lag curves on startup andshutdown plots to deviate fromthe normally expected shape.The deviation usually occursover a limited frequency rangerelating to the structuralresonance. The deviation canbe either an increase or adecrease in vibration and phaselag depending on the phaserelationship between the rotorand structural vibrations(Figure 3.35).

Steady State data: As withstartup and shutdown plots,steady state plots can also showan increase or decrease invibration amplitude and phaselag depending on how thevibrations combine. A trendplot may show a change invibration if somethingadjoining the pump undergoesa change. Such changes couldinclude a broken pipe support bracket, change in an adjoining process, etc.

One must make careful note of the fact that these changes are not unique to structuralresonances. Other malfunctions can also cause change in vibration. For example, shaftcrack can also produce changes in 1X vibration amplitude and phase over time.

Effects of Structural ResonancePump seals, bearings, and couplings are typically the parts most affected by high

vibration, including those produced by structural resonance. In one case history [5], twopumps that had operated for several years with only minor problems and annualmaintenance began to experience multiple seal, bearing, and coupling failures and torequire monthly overhauls. The cause of the failures was traced to the pumps’ dischargepiping which were found to resonate at 1X the pump running speed. The dischargepiping had been rerouted about the same time that the pumps began to fail. The new piperun had left them insufficiently supported (i.e., lower stiffness) and thus able to resonateat a frequency that coincided with the pump speed. Pump reliability was restored tooriginal levels once vibration in the piping was reduced.

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CavitationAny discussion of centrifugal pump malfunction must include the important subject

of cavitation. Cavitation is one the most common centrifugal pump malfunctions. It iscapable of causing severe performance loss and pump damage resulting in significantfinancial impacts to pump owners. Consequently, cavitation is also the focus of muchinvestigation on the part of researchers.

Definition of CavitationThe term cavitation refers to the formation of tiny vapor bubbles, or “cavities”, within

the pumped liquid that subsequently collapse with tremendous force. There may be gasbubbles of some other dissolved substance in the pumped liquid, such as air, that areexpanding and collapsing along with the vapor bubbles. However, true cavitation refersto the vaporization and subsequent collapse of the pumped liquid itself.

The vapor bubbles are capable of causing severe damage when they collapse againstthe metal surfaces inside the pump. One reference contained a photograph showing animpeller vane that had been eroded completely through by cavitation [2].

Only liquid handling machines experience cavitation because liquids by nature willboil into vapor and then condense back into liquid given the right conditions.Compressors do not experience cavitation because the gas they handle already exists in avapor state and remains so throughout the entire compression process.

Mechanism of CavitationThe mechanism of cavitation is actually the process of liquid evaporation and

condensation. Thus, if we understand the conditions for evaporation/condensation, thenwe will have defined the general physical conditions that cause cavitation.

There are two ways to evaporate, or boil, a liquid: 1) increase the temperature of theliquid to its “boiling” temperature, or 2) decrease the pressure acting upon the liquid toless than or equal to its vapor pressure. This natural phenomenon is straight out ofstandard thermodynamic principles that show that evaporation is dependent on bothtemperature and pressure. Anyone who has cooked boiled foods at high elevation hasdiscovered that more time is required because boiling temperature lowers as theatmospheric pressure lowers. For example, water at sea level will boil at 100 °C (212 °F)versus 94 °C (202 °F) at 1524 meters (5000 feet) above sea level. Since condensation ismerely the opposite of evaporation, vapor will condense when 1) its temperature islowered below the boiling point, or 2) its pressure is raised above the vapor pressure.

Applying this to centrifugal pumps, we see that cavitation will occur when either thepressure inside the pump drops below the liquid vapor pressure or the temperature of theliquid inside is raised above its boiling point. While cavitation most often occurs becauseof the former, it is also quite possible for increased temperature to cause cavitation.Because the pressure inside a pump is a function of the Net Positive Suction Head(NPSH), the lack of available NPSH is the primary cause of pump cavitation. (Seesection “Net Positive Suction Head” for greater detail about NPSH.)

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Figure 3.37 Collapse of Vsurfaces inside the pump w

Fl

The location inthe pump wherecavitation will occuris the point wherepressure is lowest.Figure 3.36 showsthat this point isadjacent the trailing(low pressure) sideof the impellervanes.

All liquids havethe potential tocavitate since allliquids follow theprinciples ofthermodynamics.However, liquidsdiffer in the severityof cavitation damagethey may cause.Denser liquids, likewater, cause more damsuch as hydrocarbons. specific volumes will cr

Effects of CavitatioAs it was alluded t

“pop”. Rather, they forcollapse, they implodatmospheres (Figure 3.3

The effects of the lofrom inconsequential tannoyance due to the noise). The presence causing damage to the

Rotation

Formation ofVapor Bubbles

Collapsing VaporBubbles

igure 3.36 Location of vapor bubbles. Cavitation occurs where pressure isowest. This is along the trailing side of the impeller vanes.

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apor Bubbles. Cavitation produces vapor bubbles which can erode metalhen they collapse against those surfaces.

age when their vapor bubbles implode than do less dense liquids, Also, liquids with larger differences between liquid and vaporeate larger implosion forces when the vapor cavities collapse.no earlier, cavitation vapor bubbles do not form and then gentlym and collapse in a few thousandths of a second. As the bubblese with tremendous pressures estimated on the order of 104

7).

cal shock wave produced by these collapsing bubbles can rangeo extremely damaging. Cavitation may be little more than ansevere noise produced (although cavitation can occur withoutof noise due to cavitation does not necessarily mean that it ispump. Some pumps may noisily operate in cavitation for years

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without failure. The tendency for cavitation to damage a pump depends on impellermaterial and design and operating conditions.

Of greater concern than noise is the loss of hydraulic efficiency. A cavitating pumphas less liquid flowing through it because the lower density vapor cavities block flow.Reduction of hydraulic efficiency always accompanies cavitation, whether or not thelosses are significant depends on the amount of cavitation present.

At worst, the local shock wave from the implosion of the vapor bubbles can erodeimpeller vanes through the removal of material from metal surfaces. The severity oferosion can vary from surface pitting to holes clear through the vanes. This can occur ina matter of a few weeks. The loss of material on the impeller upsets mass and hydraulicbalance resulting in high vibration that can damage seals and bearings.

Characteristics of Cavitation1. Reduction in Pump HeadThe loss of efficiency described above will be recognizable as a drop in the head

produced. A three percent drop in head has traditionally been used as an indicator ofcavitation. However, the onset of cavitation starts before drop in head reaches threepercent. Thus, a pump might be cavitating even without a significant drop in head.

If the pump head has dropped because of cavitation, that does not necessarily mean itis damaging the pump. Whether or not cavitation will harm a pump depends on severalfactors including the impeller material and the nature of the pumped liquid.

Drop in head will also occur if pump rotative speed is reduced. Changes in speedshould be verified before assuming cavitation is present.

2. NoiseAs described above, cavitation may or may not be accompanied by noise. If it is, the

noise tends to be a steady “crackling” noise [1]. This is in contrast to the noise ofrecirculation (another type of cavitation that is described below) which has beendescribed as being a random crackling with high-intensity knocks [1].

3. VibrationCavitation increases vibration amplitude over a broad frequency range. The vibration

increases can be high enough to damage seals and bearings.4. Visual IndicatorsErosion on the low pressure side of the impeller vanes is a sign that cavitation is

caused by insufficient NPSH. Cavitation caused by recirculation will erode other areas ofthe vanes (this is described in more detail below in “Cavitation caused byRecirculation”).

Corrective Actions1. Increase the Available NPSHIncreasing the NPSH provided by the system will raise the pressure in the pump

above the liquid vapor pressure. The suction side piping should be evaluated for thepresence of bends, elbows or other fittings that might be reducing the pressure at thepump suction to unacceptably low levels.

2. Decrease the Required NPSHAnother way to prevent pressure in the pump from falling below the liquid vapor

pressure is to reduce the Required NPSH. Required NPSH is a function of the frictionloss experienced by the liquid as it flows from the suction flange to the point of lowest

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Discharge RecirculationSuction Recirculation

Figure 3.38 Suction and Discharge Recirculation. Suction recirculation occurs in the impellereye. Discharge recirculation occurs at the discharge tips of impeller vanes.

pressure in the impeller vane passages. Examples of changes that will reduce inletfriction losses are:

Use a pump design with a lower Required NPSH. For example, double suctionpumps generally require less NPSH than a comparable single suction pump because thedouble suction eyes provide a larger inlet passage with lower frictional losses.

Use a pump that runs at lower rotative speed. A lower speed pump will have to belarger in order to deliver the same head versus flow of a smaller pump with comparableperformance.

3. Cooling the suctionCavitation can also be prevented by lowering the liquid temperature on the suction

side of the pump. Temperature at the suction must be reduced to the point where theliquid stays below its boiling temperature when flowing through the impeller vanes.(“Boiling temperature” is pressure dependent just like vapor pressure is temperaturedependent, the boiling temperature of the liquid inside pump is not the same as theboiling temperature of the liquid at atmospheric pressure.)

Cavitation caused by RecirculationDefinitionThe type of cavitation discussed in the preceding section results from insufficient

NPSH. However, a malfunction known as recirculation can also cause cavitation. Thedistinction between cavitation caused by insufficient NPSH and that caused byrecirculation is important because they have different corrective actions.

The term recirculation refers to a reversal of flow within the pump. The normaldirection of flow through the pump is from suction to discharge. However, under certain

conditions liquid will reverse direction and flow toward the suction instead of continuingto discharge as intended.

Flow reversals create vortices and cavitation occurs at the center of the vortices. Theflow reversal and its associated cavitation occur in two main areas (Figure 3.38).

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All impellers will recirculate if flow drops below a specific level. The flow at whichrecirculation occurs is impeller dependent and cannot be changed without modifying thedesign [1].

Symptoms of Recirculation1. NoiseLike the cavitation discussed in the preceding section, recirculation can also produce

noise. However, recirculation noise tends to be of greater intensity than the noise fromlow-NPSH. It has been described as a random knocking sound, as if a loose bolt or nutwhere being rattled around inside the rotating impeller.

It may be possible to distinguish whether the recirculation is occurring in the suctionor discharge depending where the noise is of highest intensity. Suction recirculationnoise will be most noticeable near the pump suction while discharge recirculation noisewill be louder at the pump discharge.

2. VibrationRecirculation cavitation can result in increased radial and axial vibration.3. Pressure PulsationsRecirculation causes a sudden increase in the magnitude of pressure pulsations.

These pulsations are detectable using pressure transducers [1].4. Visual IndicatorsRecirculation cavitation erodes the pressure side of vanes unlike low-NPSH

cavitation which attacks the low pressure side of vanes.The location of erosion indicates whether the recirculation is suction or discharge.

Pitting near impeller eye indicates suction recirculation while pitting near discharge endof vanes indicates discharge recirculation.

Corrective ActionsSince insufficient flow through the impeller causes recirculation, this flow must be

increased. Two ways to accomplish this are:1. Increasing pump output.2. Rerouting some of the pumped liquid from discharge back through the suction.

One caution with rerouting liquid back through the pump is that temperature can rise tounacceptable levels. The mechanical work of the impeller upon the liquid flowingthrough it results in a slight temperature rise. The rerouted liquid can cause heat toaccumulate in the pump thus raising the temperature unacceptably.

If neither of these is acceptable, an additional possibility exists. The susceptibility ofan impeller to cavitation damage depends on several factors, one of which is the type ofmaterial used in its construction. If cavitation cannot be eliminated, then switching to animpeller of harder material presents an additional option.

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Vane Pass Frequencies

Definition of Vane Pass FrequencyVibrations which occur at a frequency related to the number of impeller vanes, the

number of casing vanes, and pump rotative speed are known as vane pass frequencies.Because single volute pumps have only one vane (i.e., the cutwater - Figure 2.6), vanepass frequencies are usually an integer multiple of rotative speed where the integermultiple is the same as the number of impeller vanes.

The source of excitation for vane pass frequencies is the interaction between thecutwater of the pump casing and the nonuniform velocity and pressure distribution of theliquid exiting the impeller vane passages. The finite vane thickness and slower movingliquid adjacent to the vane surfaces (called a boundary layer) create variations in thevelocity and pressure of the flow exiting the impeller periphery [4]. In addition, flowfrom each impeller vane is forced to make an abrupt change in direction as it passes thecutwater [1]. These variations in velocity and pressure at each vane exit are repeatedaround the circumference of the impeller in a pattern that is evenly spaced with the vanes.As these variations in flow pass the cutwater (or casing vane tips in the case of doublevolutes and vaned diffusers), a hydraulic reaction force excites the rotor at the vane passfrequency.

Corrective ActionsVibrations at vane pass frequency represent another source of excitation of structural

resonances and also additional stress to the pump and driver. The most effective methodfor minimizing these vibrations is to maintain sufficient radial clearance betweenimpeller and cutwater (or casing vane tips). In truth, this clearance must be correctlydesigned into the pump during its initial design. One reference recommended a clearanceof not less than 5% of the impeller diameter [1]. Other than trimming, pumps do notcome equipped with means to move the impeller further away from the cutwater. Anadditional means suggested for reducing the vane passing forces is the sharpening of thetrailing edges of the vane tips [4].

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4. REFERENCES[1] Karassik, I.J., Krutzsch, W.C., Fraser, W.H., Messina, J.P. "Pump Handbook,"

Second Edition, McGraw-Hill Book Co., New York, NY.

[2] Volk, M.W., "Pump Characteristics and Applications," Marcel Dekker, Inc., NewYork, NY.

[3] Karassik, I.J., "Centrifugal Pump Clinic," Marcel Dekker, Inc., New York, NY.

[4] Corbo, M.A. and Malanoski, S.B., "Pump Rotordynamics Made Simple,"Proceedings of the 15th International Pump Users Symposium, TurbomachineryLaboratory, Texas A&M University, College Station, Texas.

[5] “Boiler Feed Pumps”, “Vertical Pumps”, “Horizontal Pumps”, Applied DiagnosticsWorkshop - Book 1, Bently Nevada Corp., Minden, NV.

[6] “Vertical Slurry Pump”, Machine Diagnostics Case Histories, MachineLibrary,Bently Nevada Corp., Minden, NV.

[7] Jackson, Charles, "Centrifugal Pumps – Maintenance & Design, Shafts, Bearings, &Sleeves" Issue Number 2, CJ On Pumping, 7/8/70.

[8] Fox, R.W., MacDonald, A.T., "Introduction to Fluid Mechanics," Fourth Edition,John Wiley & Sons, Inc., New York, NY.

[9] Ibrahim, A. and Sace, E., "ADRE for Windows – instrumental in solving a complexvibration problem on a boiler feedwater pump," Orbit, Bently Nevada Corp., v.19,No. 1, March 1998.

[10] Eisenmann, Robert C., Sr., and Eisenmann, Robert C., Jr. "Machinery MalfunctionDiagnosis and Correction," Hewlett-Packard Professional Books, Prentice-Hall, Inc.,Upper Saddle River, New Jersey.

[11] Hatch, Charles T., "Malfunction Diagnosis: Unbalance and 1X Vibration,"MachineLibrary, Bently Nevada Corp., Minden, NV.

[12] Hatch, Charles T. and Fahy, Dave "Malfunction Diagnosis: Misalignment,"MachineLibrary, Bently Nevada Corp., Minden, NV.

[13] Hatch, Charles T., "Malfunction Diagnosis: Rub," MachineLibrary, BentlyNevada Corp., Minden, NV.

[14] Hatch, Charles T., "Malfunction Diagnosis: Shaft Crack," MachineLibrary,Bently Nevada Corp., Minden, NV.

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[15] Hatch, Charles T., "Malfunction Diagnosis: Fluid-Induced Instability,"MachineLibrary, Bently Nevada Corp., Minden, NV.

[16] Muszynska, A., "One Lateral Mode Isotropic Rotor Response to NonsynchronousExcitation," BRDRC Report No. 4, 1991, pp. 1-31, MachineLibrary, BentlyNevada Corp., Minden, NV.