tolerances pump systems article

7/27/2019 Tolerances Pump Systems article

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By Heinz P. Bloch, P.E. Consulting Engineer

If you are like most industrial facil-ities in the industrialized world,your worker and technicianresources are probably stretched tothe limit. Understandably, youmight be looking for ways to sim-plify some of your traditional workprocesses and procedures. You mayeven have had an experience thatreinforces the contention thathigh-tech tools are not always theanswer, and that back-to-basics

thinking has considerable merit.

While no reasonable and experienced reliabili-ty professional will take issue with the above find-

ings, industry must be cautioned against drawing

the wrong conclusions. A recent example of wrong

conclusions involves claims that rotating equip-

ment alignment is sufficiently accurate as long as

the shaft centerlines in their standstill, or cold, con-

dition are within 0.002 inches (0.05 mm) of each

other. If you blindly follow this questionable advice,

you may soon find yourself among the repair-focused dinosaurs that are struggling to survive.But, if you update your knowledge of shaft

alignment and alignment tolerances, you might beon the way to becoming reliability-focused.Indications are that only reliability-focused facilitieswill be around a few years from now!

How Shaft Alignment TolerancesMust Be Expressed

The only correct way to express shaft alignmenttolerances is in terms of alignment conditions at thecoupling, and we will describe several ways to dothis. It is incorrect to describe them in terms of cor-rection values at the machine feet, and we willexamine why. Before we get into too much detail,however, lets first examine our alignment objectives.

When two machines are directly coupled via aflexible coupling, any misalignment between their

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centerlines of rotation can result in vibration,which, depending on its severity, can produce pre-mature wear, or even catastrophic failure of bear-ings, seals, the coupling itself and other rotatingcomponents. Indeed, misalignment of the center-

lines of rotation has long been recognized as one ofthe leading causes of machinery damage.

Decades of well-documented observationattest to the fact that misalignment has beenresponsible for huge economic losses. The moremisalignment, the greater the rate of wear, likeli-hood of premature failure and loss of machine effi-ciency. Moreover, misaligned machines absorbmore energythey consume more power. Yet, evenexcellent alignment of the shaft centers of rotationdoes not in itself guarantee absence of vibration.

This is because there is still the possibility of imbal-ance of rotating components. Structural resonance,fluid flow turbulence and cavitation, or even vibra-tion from nearby running machines that is trans-mitted to adjacent machines through either foun-dation or piping also could be present.

Why Reliability-FocusedCompanies Will Not Use SimpleFoot Alignment

Absolute perfection in the alignment of shafts

is not realistically attainable, nor is it needed. Agood analogy is the polishing metal: No matter howlong it is polished and how fine the different pol-ishing media, a powerful microscope will detect asurface composed of peaks and valleys. The ratherobvious issue is quantification of alignment qualityand allowable deviationthe alignment tolerance.

We define misalignment by visualizing theshaft centerlines of rotation as two straight lines inspace. The trick is to get them to coincide to formonestraight line. If they dont, then there must exist

either offsetmisalignment (Figure 1) or angularmis-alignment (Figure 2), or a combination of both.

Furthermore, since the shafts exist in three-dimensional space, these misalignments can exist inany direction. It is most convenient, therefore, forthe purpose of description to break-up this three-dimensional space into two planes, the vertical and

the horizontal, and to describe the specific amountof offset and angularity that exists in each of theseplanes simultaneously, at the location of the coupling.Thus, we end up with four specific conditions ofmisalignment, traditionally called Vertical Offset(VO), Vertical Angularity (VA), Horizontal Offset(HO), and Horizontal Angularity (HA). These con-ditions are described at the location of the coupling,because that is where harmful machinery vibrationis created whenever misalignment exists.

The magnitude of an alignment tolerance (i.e.

the description of desired alignment quality), musttherefore be expressed in terms of these offsets andangularities, or the sliding velocities resulting fromthem. Attempts to describe misalignment in termsof foot corrections alone do not take into accountthe size, geometry or operating temperature of agiven machine. It can be shown that accepting thesimple foot corrections approach can seriouslycompromise equipment lifeand has no place in areliability-focused facility. An illustration of thefallacy of the foot correction approach will be

given later.How much vibration and efficiency loss will

result from the misalignment shaft centers dependson shaft speed and coupling type. Acceptable align-ment tolerances are thus functions of shaft speedand coupling geometry. It should be noted thathigh-quality flexible couplings are designed to tol-erate more misalignment than what is good for themachines involved. Bearing load increases with mis-alignment, and bearing life decreases as the cube ofthe load increase (i.e. doubling the load will short-

en bearing life by a factor of eight).Why, then, would high-quality flexible couplings

PUMPS & SYSTEMS www.pump-zone.com DECEMBER 2003 17

Fig. 1 Offset Misalignment Fig. 2 Angular Misalignment


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generally be able to accommodate greater misalign-

ment than what is good for the connected

machines? Well, a large percentage of machines

must be deliberately misalignedsometimes signifi-

cantly soin the cold and stopped condition. Asthey reach operating speeds and temperatures, ther-

mal growth is anticipated to bring the two shafts

into alignment. The following case history illus-

trates the point.

A refinery has a small steam turbine, foot-

mounted and enveloped in insulating blankets. Theoperating temperature of the steel casing is 455 F,

and the distance from centerline

to the bottom of the feet is 18

inches. The turbine drives an

ANSI pump with a casing tem-

perature of 85F; its centerline-to-

bottom-of-feet distance is also 18

inches. Both initially started up at

the same ambient temperature.

The differential in their growth is

(0.0000065 inches per inch per

deg. F) x 18 inches x (455 85)

deg. F = 0.043 inches.

If these two machines had

their shafts aligned center-to-cen-

ter, this amount of offset would

be certain to cast the equipment

train into the frequent failure cat-egory. Using the 80/20 rule, it

would be safe to assume that 20

percent of our machinery popula-

tion eats up 80 percent of our

maintenance money. This pump

train would be a problem child

in the 20 percent group.

As we mentioned earlier,

aligning center-to-center without

paying attention to thermalgrowth is surely one of the factors

that keeps its practitioners in the

repair-focused category. Using

alignment tools and procedures

that take into account all of the

above is mandatory for reliability-

focused companies.


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Describing PermissibleTolerances

There are a number of acceptable ways todescribe misalignment at the coupling and to definepermissible misalignment tolerances. While manyare of equal merit, describing alignment tolerancesin terms of foot corrections isunacceptable in a reliability-focused environment. Letslook at the most prominent ofthe various approaches wemight use, then summarizewhat not to do.

Offset and angularity at thecoupling (for short couplings)

Offset and angularity atthe coupling is one of themost common ways of correct-ly defining alignment toler-ances. The offset tolerance sim-ply describes the maximumseparation that can existbetween two machine shafts ata specific location along theirshaft axes, usually the coup-ling center. The angularity

describes the rate at which theoffset between the shaft center-lines may change as we travelalong the axes of the shafts.Figure 3 serves as a typicalillustration of such a toleranceenvelope.

Figure 3. Typical toleranceenvelope

The angularity may bedescribed either directly, as anangle in terms of mils per inch(or milliradians), or as a gapdifference at a particular cou-

pling diameter. The latter method is popularbecause it relates directly to what a mechanic candetect with his feeler gages between the couplingfaces. A modern laser shaft alignment system mea-sures the angle between shaft centerlines; such a sys-

tem also can be set to describe this angle as a gapdifference at any desired diameter (Fig. 4).

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Figure 4. Measuring the angle between shaft centerlines

This approach can have two different interpre-tations, however. If we describe a permissible offsetbetween the driver and driven shafts as beingxamount in any direction, orx amount individuallyin both the horizontal and vertical planes? Thesetwo alternatives are notthe same! The first exampleis called vector tolerance and is more conserva-tive. The second approach, called standard toler-ance, is more commonly used. If you dont wantmore thanxamount of offset to exist between theshafts in any direction, standard tolerances shouldnot be used. Doing so would, in some circum-stances, lead to greater-than-intended offsets.Figures 5 and 6 illustrate this point.

Figure 5. Results of using standard tolerances

Figure 5 depicts a case where applying standardtolerances results in an offset of 2.5 mils horizontallyand 2.7 mils vertically being deemed acceptable,because the permissible limit for either of these off-sets individually is 3.0 mils. However, the actual off-set between the shafts is 3.7 mils ( ),which

is unacceptable if your absolute limit is 3.0 mils.This result can be seen as a vector tolerance,shown in Figure 6. A good laser shaft alignment sys-tem will allow the user to make this distinction andto specify exactly which type of tolerance is desired.

Figure 6. Results of using vector tolerances

As regards tolerances, Table 1 presents the val-ues most widely accepted as the standard industrynorm for short couplings:

Spacer Coupling TolerancesSpacer coupling tolerances are generally

expressed as limits to the angle that may existbetween each machine shaft, and the spacer piecebetween them. Since the spacer piece (or spool

piece) connects to each of the machine shafts ateither end, by definition, there should be no offsetbetween the spacer and each of the machine shafts.Consequently, all that needs to be specified is themaximum angle allowed between the spacer shaftand each of the connected machine shafts. Thisangle may be specified directly in mils per inch (ormilliradians), or in terms of the offset that eachindividual angle between spacer and machine shaft


Table 1


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Figure 7. Methods for specifying maximum angle allowed

between spacer shaft and connected machine shafts

projects to the opposite end of the spacer. The firstway is called the angle-angle method (also some-times called the alpha-beta method), and the sec-

ond is called the offset-offset (or offset A-offset-B) method. Figure 7 illustrates this rather well.

Since most flexible couplings have two flexplanes (or points of articulation), the spacer cou-pling tolerances may safely be used with all suchcouplings, even the ones usually considered shortflex couplings. The best criterion to make the dis-tinction is the relation between the diameter of the

flex planes and the distance between them.Whenever the distance between flex planes (span) isgreater than the diameter, reliability professionalscall it a spacer coupling. This will make achievingtolerances easier when performing alignment cor-

rections in the field. It in no way diminishes theconservatism of these values. Table 2 presents thevalues most widely accepted as the standard indus-try norm for spacer couplings.

Sliding velocity tolerancesAnother approach for specifying tolerances is to

describe the permissible limit of the velocity thatthe moving elements in a flexible coupling mayattain during operation. This can be easily relatedto the maximum permissible offset and angularity


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Table 2


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through the formula:Maximum allowable component sliding velocity =

2xdxrxaxwhere: d= coupling diameter, r= revolutions/time,and a= angle in radians.

When offset and angularity exist, the flexible ormoving coupling element must travel by double the

amount of the offset and angularity, every half-rota-tion. Since the speed of rotation is defined, so mustbe the velocity that is achieved by the moving ele-ment in accommodating this misalignment as theshaft turns. In essence, when we limit the permissi-

ble sliding velocity, we haveby definitionalsolimited the offset and angularity(in any combination) that canexist between the coupled shaftsas these turn. For 1,800 RPM,typically this limit is about 1.13inches per second for excellentalignment, and 1.89 inches persecond for acceptable alignment.A good laser alignment systemwill let reliability-focused users

apply this approach as well.

Not Acceptable: TolerancesExpressed as Corrections at theMachine feet

Again, and for emphasis, thisapproach is wrong. It is impossi-ble to define the quality of thealignment between rotating shaftcenterlines in terms of correctionvalues at the feet alone, unless onealso specifies the exact dimensionsrelated to these specific correction

values, and does so each time.This approach is too cumbersomeand error-prone, since twomachines will rarely share thesame dimensions between thecoupling and the feet, andbetween the feet themselves.

A tolerance that only de-scribes a maximum permissiblecorrection value at the feet with-out references to the operativedimensions involved makes nosense. This is because the same

correction values can yield vastlydifferent alignment conditionsbetween the machine shafts withdifferent dimensions. Such a tol-erance simply ignores the effectsof RISE over RUN, which isessentially what shaft alignment isall about. Furthermore, such a tol-erance does not take into accountthe type of coupling or the rotat-


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ing speed of the machines.These alignment tolerances specifiedgenerical-

lyin terms of foot corrections can have two equallybad consequences: the values may be met at the feet,yet allow poor alignment to exist between the shafts;or, these values may be greatly exceeded, while rep-resenting excellent alignment between the shafts!This means that, in the first scenario, the aligner maystop correcting his alignment before the machines areproperly aligned, and in the second case may be mis-led into continuing to move machines long after theyhave already arrivedin tolerance.

Lets review a couple of examples that illustratethe fallacy of the generic foot correction approachand assume that the specified alignment tolerancefor an 1,800 RPM machine is defined as a maxi-mum correction value for the machine feet, 2 mils.A machine is found to require a correction of 2

mils at the front feet and +2 mils at the back feet,therefore it is deemed by this method to be in tol-erance. If the distance between the feet is 8 inches,this would imply an existing angular misalignmentof 0.5 mils per inch. If the distance from the frontbearing to the coupling center is 10 inches, the

resulting offset between the machine shafts at thecoupling would be +7.0 mils! This offset is consid-erably in excess of the 3 mils of offset (either stan-dard or vector) that is considered the maximumacceptable for an 1,800 RPM machine at the cou-pling. Yet, with the improperly specified foot cor-rection tolerances, this alignment would beerro-neouslyconsidered to be in tolerance! This is aclassic example of where small correction values atthe feet do notnecessarily reflect good alignment atthe coupling. Figure 8 explains this quite well.

Figure 8. Small correction values at the feet do not neces-sarily reflect good alignment at the coupling.



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An equally bad consequence of this approachis that the opposite scenario is just as likely tooccur. Assume you have a large machine (such asa diesel engine), running at 1,200 RPM, whosedistance between the feet is 80 inches. Further,

the distance from front feet to thecoupling is 30 inches. Such amachine is found to have misalign-ment requiring foot corrections of8 mils at the front feet and 26mils at the back feet. In this case,the resulting misalignment at the

coupling is only 1.25 mils of offsetand only 0.225 mils per inch ofangularity. Referring to the cou-pling, both of these alignment con-ditions are already much better thanrequired by the standard industrynorms for 1,200 RPM. However,using the improperly specified toler-ance values of 2 mils at the feet,the aligner would be misled intoworking much harder and longer

than necessary to bring themachines to these values. This canbe observed in Figure 9. P&S

Figures 5, 6, 8 and 9 createdwith the assistance of AlignmentExplorer software by Prftechnik,Ismaning/Germany, and input fromLudeca, Miami, FL.

Heinz P. Bloch, P.E. is a consult-ing engineer with over 40 years of

experience in chemical process plantsand oil refineries. He advises industryon reliability improvement andmaintenance cost reduction issues andis the author or co-author of 13 com-

prehensive texts and over 280 papersor articles on reliability-related sub-

jects. His Pump Life ExtensionHandbook will be available in

2004.


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Figure 9. Using improperly specified tol-erance values at the feet

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