shop balancing tolerances.pdf

7/22/2019 shop balancing tolerances.pdf

1/7

Balancing in the shop is analogous to the couplingalignment example, in that in both cases we must estab-lish certain criteria to be applied in a nonoperating statethat will insure acceptable performance when operatingn o r m a l l y.

The level of residual unbalance (heavy spot) that canbe tolerated on a rotor is re la ti ve to its mass and speed.The formula

Funb = 1.77 (RPM/1000)2 (oz-in)WHERE:Funb is the force due to unbalance in lbsoz-in is the residual heavy spot

dictates that the unbalance force is pro p o rtional to thesquare of the RPM and the amount of residual heavy spot.Since the goal of balancing is to reduce the in-place vibra-tion (Fun b), then either the speed must be slowed or theheavy spot reduced. There is normally only one option

open, since the speed is usually set as a design criteria ofthe machine and/or process. The only choice re m aining,therefore, is to control the amount of residual heavy spotremaining after balancing.

This is the justification for balance tolera n c e s theyestablish an acceptable residual heavy spot amount which,if achieved, will insure acceptable in-place vibration levels.

In days of yore, balance tolerances were established bythe capability of the balancing machine. In other words ,

the balance job was complete when the system reached itslevel of sensitivity to noise level interference from sourcesother than unbalance (sometimes referred to as signal-to-noise ratio). It is no longer reasonable to take a balancelevel down to loss of phase angle (unbalance location) dueto purely economic considerations. Most modern balanc-ing systems with digital filtering offering as much as 85 dBof dynamic range can achieve residual unbalance levels far

below those normally required for smooth rotor operationin the field.

While economics may dictate that performing microbalancing on some rotors is unnecessary, the reverse situa-tion may also occur when attempting to utilize a balancingsystem with less-than-desirable sensitivity. Phase loss in thiscase might occur prior to achieving acceptable re s i d u a lunbalance levels, making the application of a balance qual-ity level imperative.

ROTOR DYNAMICS - PRELIMS TO

CHOOSING BALANCE PLANESA typical rotor has planes containing unbalance. Sincemost rotors are somewhat asymmetrical in configuration, itis difficult to determine which of the planes contain thelargest amounts of unbalance. The unbalance could be inany plane or planes located along the axis of the rotor andit would be most difficult and time consuming to deter-mine exactly which plane(s) was guilty. Furthermore, it isnot always possible to make weight corrections in just any

B A L A N C I N G

Balanced scrawled in chalk across a rotor all too often goes unchallenged. Balance is

a very relative quantity - all rotors are balanced, but how well does the chalk message

relate to the final product in its field application? In the same way a coupling can be

described as aligned, but is it aligned closely enough to operate smoothly?

SHOP BALANCING TOLERANCES

A PRACTICAL GUIDEB Y E A R L M . H A L F E N


2/7

plane. Theref o re, the usual practice is to compromise bymaking weight corrections in the most convenient planes.It is possible to successfully make this compro m i s ebecause, for a rigid rotor, any condition of unbalance can

be compensated for by weight correction in any two bal-ancing planes. Again, this is true only if the rotor and shafta re r igid and do not bend or deflect due to the forc e scaused by the unbalance.

Classifying a rotor as either rigid or flexible depends onthe relationship between the rotating speed and the naturalfrequency (fn). Natural frequency can be simply defined asthe frequency at which the rotor likes to vibrate (a formaldefinition depends on such parameters as stiffness coeffi-cients, mass, molecular structures, etc. remember this is asimplified guide!) When the fn of a machine part is also

equal to the RPM or some other exciting frequency, a con-dition ofresonance exists. The rotating speed at which therotor goes into bending resonance is called acritical speed.

St a rting with a machine at rest, if the speed of themachine is increased while measuring vibration amplitude,a plot similar to Fi g u re 1 would be generated. Note theincrease in vibration followed by a decrease to a fairly con-stant level. The RPM at which the peak occurs is where thebending resonance occurs and is called the critical speed.

In actual practice, a plot of vibration amplitude versusRPM may show several peaks as illustrated in Figure 2. The

additional peaks may be due to resonance of the bearingsand supporting stru c tu re, which are different from shaftcriticals. The shaft and rotor may have more than one criti-cal speed, in addition to that reflected in Figure 2. In anycase, when discussing rigid versus flexible rotors, referenceshould be made to the shaft and rotor critical speed and notthe resonance of the supporting structure. As a general rule,rotors that operate below 70% oftheir critical speed are consideredri g id. When these rotors are bal-anced at one speed they will

remain balanced at any other nor-mal operating speed below 70% ofcritical. Since it is almost impossi-ble to accurately simulate fieldconditions of temperature, pre s-sure, speed, bearing stiffness, tor-sional loading, damping factors,etc. in the shop environment, thevast majority of shop balancing is

done at low speed (from 100 to 1,000 RPM). This meansthat the majority of rotors balanced in the shop are in a

rigid mode. A determination of the actual type of the rotormust be made because rotors which operate above 70% oftheir critical speed will actually bend or flex due to theforces of unbalance. These flexible rotors must be treatedwith extreme care when choosing the correction planes. Inreality, almost all rotors encountered will be rigid, with highspeed turbo-machinery being the exception. Most fans,electric motors, pumps, etc. are rigid and can be handled inthe appropriate manner. In fact, most flexible rotors areknown by the manufacturer and appropriate informationon balancing is available from them.

CHOOSING CORRECTION PLANES

A flexible rotor balanced at one operating speed may notremain balanced when operating at another speed. Thisalso means that a flexible rotor balanced at low speed inthe shop environment (in its rigid mode) may not remainbalanced when operating at field RPM.

To illustrate, consider theunbalanced rotor in Fi g u re 3A.The unbalance shown is a combi-nation of static and couple unbal-

ance by definition, dynamicunbalance. If this rotor were bal-anced in a normal shop balancingmachine with correction we i g h t sadded in the two end planes,these correction weights wouldcompensate for all sources ofunbalance distributed thro u g houtthe ro t o r. Howe ve r, when the

Fi g u re 2.

Fi g u re 1.


3/7

in-place RPM (operat-ing) is encount ere d ,with the rotor nowoperating above 70% of

its critical, the rotor willdeflect due to the cen-trifugal force of theunbalance located at thecenter portion asdepicted in Fi g u re 3B.As the rotor bends ordeflects, the weight ofthe rotor is moved out away from the rotating centerline,creating a new unbalanced condition. This new unbalancecan be corrected by re-balancing in the two end planes

(assuming that internal clearances where the maximumdeflection occurs would allow it). Howe ve r, the ro t o rwould then be out of balance at slower speeds where thereis no deflection. The only solution to insure smooth oper-ation at all speeds is to make the balance corrections inthe actual planes of unbalance (or, at least to cut down thec o m p romise, since we have already stated that there existan infinite number of unbalance planes in eve ry ro t o r ) .Thus the rotor in Fi g u re 3 would re q u i re balancing inthreepla n e s .

The flexible rotor in Figure 3 actually re p resents the

simplest type of non-rigid rotor. A rotor can deflect in sev-eral ways, depending on its operating speed and the distrib-ution of unbalance throughout the axis. For example,Figure 4illustrates the first, second and third flexural modesa rotor could experience. These are also called first, secondand third rotor critical speeds and are usually encounteredon high speed machines such as multi-stage centrifugal pumps and compressors,as well as many steam and gas turbines.

These machines may require that bal-ance corrections be made in several planes

and are often designed with multiple cor-rection planes. Howe ve r, not all flexiblerotors require multi-plane balancing. Thiscan only be determined by the normaloperating speeds of the rotor and the sig-nificance of rotor deflection on the func-tional re q u i rements of the machine.Flexible rotors generally fall into one ofthe following categories:

x If the rotor operatesat only one speed anda slight amount ofdeflection will not

accelerate wear orhamper the prod u c-tivity or safety of themachine, then bal-ancing in any twoc o r rection planes tom i n i m i ze bearingvibration is sufficient.

x If a flexible rotor operates at only one speed, but it isessential that rotor deflection be minimized, thenmulti-plane balancing may be required. For example,

excessive deflection of the rolls used in paper machinesmay result in variations in product quality. This makesit necessary to balance in multiple planes to minimizeboth bearing vibration and rotor deflection.

x If it is essential that a rotor operate smoothly over awide range of speeds where the rotor is changingbetween rigid and flexible modes, then multiplane bal-ancing is required.

Now, why all of this discussion of rotor dynamics? It iso bvious that the shop balancing machine operator must

consider the dynamic characteristics of the rotor he ischarged with balancing. And, furthermore, if a balance tol-erance is to be applied correctly, we must consider the num-ber of correction planes and the type of unbalance, such asstatic, couple or dynamic (see the Glossary for definitions).

CHOOSING A BALANCE

TOLERANCE PITFALLS &

PRACTICALITY

With the pre l i m i n a ry foundation estab-lished, lets explore several of the most com-

mon balance standards, keeping in mindseveral points:

x Some standards are tighter thanothers (as you will see in the follow-ing examples).

x The old adage If a little bit is good,then a whole lot should be better doesnot necessarily apply to balancing on a

Fi g u re 3A. Fi g u re 3B.

Fi g u re 4.

ROTOR UNBALANCE

BALANCECORRECTIONS

A. FIRSTCRITICAL

B. SECONDCRITICAL

C. THIRDCRITICAL


4/7

wide range of rotor types in other words, apply-ing the tightest toleranceavailable to all rotors may

be impractical from a timeand resources standpoint.

x Tighter tolerances call forbetter balancing equipmentand better rotor journalquality.

x Tighter tolerances call formuch better mechanicalfit-up on componentrotors dont blame thebalancing machine if a rotor is balanced to super fine

levels and then performs poorly due to a sloppy inter-ference fit upon reassembly.

For the purpose of this paper, four (4) shop balancing tol-erances will be discussed and applied to the rotor depictedin Fi g u re 5. The author will not attempt to rec o m men dany one tolerance.

CENTRIFUGAL FORCE


5/7

API (AMERICAN

PETROLEUM INSTITUTE)

The API tolerance is, in effectone half of the USN MIL-STD

specification, in, that it a llow sfor the formula to contain static

journa l load ing instead of to ta lrotor weight (assuming that therotor i s symmetrical and sup-p o rted by two journals).

Uper = 4W/N

where:

W = Static Journal Load

N = Maximum Continuous Operating RPMOur example:

Uper = 4(750)/4000 = 0.75 oz-in

SUMMARY OF TOLERANCES:

APPLYING UPERTO NARROW PLANE &

OVERHUNG ROTORS

If the balance correction planes do not exist

between bearings or if the correction planesare quite narrow in comparison to the jour-nal-to-journal distance, then some specialrules need to be applied. Referring to Figures6 and 7the following applies:

x Distance between correction planes is

Documents

shop balancing tolerances.pdf