Maintenance Engineering -- Rotor Balancing





Mechanical imbalance is one of the most common causes of machinery vibration and is present to some degree on nearly all machines that have rotating parts or rotors. Static, or standing, imbalance is the condition in which there is more weight on one side of a centerline than the other. However, a rotor may be in perfect static balance and not be in a balanced state when rotating at high speed.

If the rotor is a thin disc, careful static balancing may be accurate enough for high speeds. However, if the rotating part is long in proportion to its diameter, and the unbalanced portions are at opposite ends or in different planes, the balancing must counteract the centrifugal force of these heavy parts when they are rotating rapidly.

This section provides information needed to understand and solve the majority of balancing problems by using a vibration/balance analyzer, a portable device that detects the level of imbalance, misalignment, etc., in a rotating part based on the measurement of vibration signals.

SOURCES OF VIBRATION CAUSED BY MECHANICAL IMBALANCE

Two major sources of vibration caused by mechanical imbalance in equipment with rotating parts or rotors are (1) assembly errors and (2) incorrect key length guesses during balancing.

Assembly Errors

Even when parts are precision balanced to extremely close tolerances, vibration caused by mechanical imbalance can be much greater than necessary because of assembly errors. Potential errors include relative placement of each part's center of rotation, location of the shaft relative to the bore, and cocked rotors.

Center of Rotation

Assembly errors are not simply the additive effects of tolerances but also include the relative placement of each part's center of rotation. For example, a ''perfectly'' balanced blower rotor can be assembled to a ''perfectly'' balanced shaft and yet the resultant imbalance can be high. This can happen if the rotor is balanced on a balancing shaft that fits the rotor bore within 0.5 mils (0.5 thousandths of an inch) and then is mounted on a standard cold-rolled steel shaft allowing a clearance of over 2 mils.

Shifting any rotor from the rotational center on which it was balanced to the piece of machinery on which it's intended to operate can cause an assembly imbalance four to five times greater than that resulting simply from tolerances.

For this reason, all rotors should be balanced on a shaft having a diameter as nearly the same as possible as the shaft on which it will be assembled.

For best results, balance the rotor on its own shaft rather than on a balancing shaft. This may require some rotors to be balanced in an overhung position, a procedure the balancing shop often wishes to avoid. However, it's better to use this technique rather than being forced to make too many balancing shafts. The extra precision balance attained by using this procedure is well worth the effort.

Method of Locating Position of Shaft Relative to Bore

Imbalance often results with rotors that don't incorporate setscrews to locate the shaft relative to the bore (e.g., rotors that are end-clamped). In this case, the balancing shaft is usually horizontal. When the operator slides the rotor on the shaft, gravity causes the rotor's bore to make contact at the 12 o'clock position on the top surface of the shaft. In this position, the rotor is end-clamped in place and then balanced.

If the operator removes the rotor from the balancing shaft without marking the point of bore and shaft contact, it may not be in the same position when reassembled. This often shifts the rotor by several mils as compared with the axis on which it was balanced, thus causing an imbalance to be introduced. The vibration that results is usually enough to spoil what should have been a precision balance and produce a barely acceptable vibration level. In addition, if the resultant vibration is resonant with some part of the machine or structure, a more serious vibration could result.

To prevent this type of error, the balancer operators and those who do final assembly should follow the following procedure. The balancer operator should permanently mark the location of the contact point between the bore and the shaft during balancing. When the equipment is reassembled in the plant or the shop, the assembler should also use this mark. For end-clamped rotors, the assembler should slide the bore on the horizontal shaft, rotating both until the mark is at the 12 o'clock position, and then clamp it in place.

Cocked Rotor

If a rotor is cocked on a shaft in a position different from the one in which it was originally balanced, an imbalanced assembly will result. If, for example, a pulley has a wide face that requires more than one setscrew, it could be mounted on-center but be cocked in a different position than during balancing. This can happen by reversing the order in which the setscrews are tightened against a straight key during final mounting as compared with the order in which the setscrews were tightened on the balancing arbor. This can introduce a pure couple imbalance, which adds to the small couple imbalance already existing in the rotor and causes unnecessary vibration.

For very narrow rotors (i.e., disc-shaped pump impellers or pulleys), the distance between the centrifugal forces of each half may be very small. Nevertheless, a very high centrifugal force, which is mostly counterbalanced statically by its counterpart in the other half of the rotor, can result. If the rotor is slightly cocked, the small axial distance between the two very large centrifugal forces causes an appreciable couple imbalance, which is often several times the allow-able tolerance. This is because of the fact that the centrifugal force is proportional to half the rotor weight (at any one time, half of the rotor is pulling against the other half) times the radial distance from the axis of rotation to the center of gravity of that half.

To prevent this, the assembler should tighten each setscrew gradually-first one, then the other, and back again-so that the rotor is aligned evenly. On flange-mounted rotors such as flywheels, it's important to clean the mating surfaces and the bolt holes. Clean bolt holes are important because high couple imbalance can result from the assembly bolt pushing a small amount of dirt between the surfaces, cocking the rotor. Burrs on bolt holes also can produce the same problem.

Other

Other assembly errors can cause vibration. Variances in bolt weights when one bolt is replaced by one of a different length or material can cause vibration. For setscrews that are 90 degrees apart, the tightening sequence may not be the same at final assembly as during balancing. To prevent this, the balancer operator should mark which was tightened first.

Key Length

With a keyed-shaft rotor, the balancing process can introduce machine vibration if the assumed key length is different from the length of the one used during operation. Such an imbalance usually results in a mediocre or ''good'' running machine as opposed to a very smooth running machine.

For example, a ''good'' vibration level that can be obtained without following the precautions described in this section is amplitude of 0.12 in./sec (3.0 mm/sec).

By following the precautions, the orbit can be reduced to about 0.04 in./sec (1 mm/sec). This smaller orbit results in longer bearing or seal life, which is worth the effort required to make sure that the proper key length is used.

When balancing a keyed-shaft rotor, one half of the key's weight is assumed to be part of the shaft's male portion. The other half is considered to be part of the female portion that's coupled to it. However, when the two rotor parts are sent to a balancing shop for rebalancing, the actual key is rarely included. As a result, the balance operator usually guesses at the key's length, makes up a half key, and then balances the part. (Note: A ''half key'' is of full-key length but only half-key depth.) To prevent an imbalance from occurring, don't allow the balance operator to guess the key length. It is strongly suggested that the actual key length be recorded on a tag that's attached to the rotor to be balanced. The tag should be attached in such a way that another device (such as a coupling half, pulley, fan, etc.) can't be attached until the balance operator removes the tag.

THEORY OF IMBALANCE

Imbalance is the condition in which there is more weight on one side of a centerline than on the other. This condition results in unnecessary vibration, which generally can be corrected by the addition of counterweights. There are four types of imbalance: (1) static, (2) dynamic, (3) couple, and (4) dynamic imbalance combinations of static and couple.

Static

Static imbalance is single-plane imbalance acting through the center of gravity of the rotor, perpendicular to the shaft axis. The imbalance also can be separated into two separate single-plane imbalances, each acting in-phase or at the same angular relationship to each other (i.e., 0 degrees apart). However, the net effect is as if one force is acting through the center of gravity. For a uniform straight cylinder such as a simple paper machine roll or a multi-grooved sheave, the forces of static imbalance measured at each end of the rotor are equal in magnitude (i.e., the ounce-inches or gram-centimeters in one plane are equal to the ounce-inches or gram-centimeters in the other).

In static imbalance, the only force involved is weight. For example, assume that a rotor is perfectly balanced and therefore will not vibrate regardless of the speed of rotation. Also assume that this rotor is placed on frictionless rollers or ''knife edges.'' If a weight is applied on the rim at the center of gravity line between two ends, the weighted portion immediately rolls to the 6 o'clock position because of the gravitational force.

When rotation occurs, static imbalance translates into a centrifugal force. As a result, this type of imbalance is sometimes referred to as ''force imbalance,'' and some balancing machine manufacturers use the word ''force'' instead of ''static'' on their machines. However, when the term ''force imbalance'' was just starting to be accepted as the proper term, an American standardization committee on balancing terminology standardized the term ''static'' instead of ''force.'' The rationale was that the role of the standardization committee was not to determine and /or correct right or wrong practices but to standardize those currently in use by industry. As a result, the term ''static imbalance'' is now widely accepted as the international standard and therefore is the term used in this document.

Dynamic

Dynamic imbalance is any imbalance resolved to at least two correction planes (i.e., planes in which a balancing correction is made by adding or removing weight). The imbalance in each of these two planes may be the result of many imbalances in many planes, but the final effects can be characterized to only two planes in almost all situations.

An example of a case in which more than two planes are required is flexible rotors (i.e., long rotors running at high speeds). High speeds are considered to be revolutions per minute (rpm) higher than about 80% of the rotor's first critical speed. However, in over 95% of all run-of-the-mill rotors (e.g., pump impellers, armatures, generators, fans, couplings, pulleys, etc.), two-plane dynamic balance is sufficient. Therefore, flexible rotors are not covered in this document because of the low number in operation and the fact that specially trained people at the manufacturer's plant almost always perform balancing operations.

In dynamic imbalance, the two imbalances don't have to be equal in magnitude to each other, nor do they have to have any particular angular reference to each other.

For example, they could be 0 (in-phase), 10, 80, or 180 degrees from each other.

Although the definition of dynamic imbalance covers all two-plane situations, an understanding of the components of dynamic imbalance is needed so that its causes can be understood. Also, an understanding of the components makes it easier to understand why certain types of balancing don't always work with many older balancing machines for overhung rotors and very narrow rotors. The primary components of dynamic imbalance include the number of points of imbalance, the amount of imbalance, the phase relationships, and the rotor speed.

Points of Imbalance

The first consideration of dynamic balancing is the number of imbalance points on the rotor, as there can be more than one point of imbalance within a rotor assembly. This is especially true in rotor assemblies with more than one rotating element, such as a three-rotor fan or multi-stage pump.

Amount of Imbalance

The amplitude of each point of imbalance must be known to resolve dynamic balance problems. Most dynamic balancing machines or in situ balancing instruments are able to isolate and define the specific amount of imbalance at each point on the rotor.

Phase Relationship

The phase relationship of each point of imbalance is the third factor that must be known. Balancing instruments isolate each point of imbalance and determine their phase relationship. Plotting each point of imbalance on a polar plot does this. In simple terms, a polar plot is a circular display of the shaft end. Each point of imbalance is located on the polar plot as a specific radial, ranging from 0 to 360 degrees.

Rotor Speed

Rotor speed is the final factor that must be considered. Most rotating elements are balanced at their normal running speed or over their normal speed range.

As a result, they may be out of balance at some speeds that are not included in the balancing solution. As an example, the wheels and tires on your car are dynamically balanced for speeds ranging from zero to the maximum expected speed (i.e., 80 miles per hour). At speeds above 80 miles per hour, they may be out of balance.

COUPLED

Coupled imbalance is caused by two equal non-colinear imbalance forces that oppose each other angularly (i.e., 180 degrees apart). Assume that a rotor with pure couple imbalance is placed on frictionless rollers. Because the imbalance weights or forces are 180 degrees apart and equal, the rotor is statically balanced.

However, a pure couple imbalance occurs if this same rotor is revolved at an appreciable speed.

Each weight causes a centrifugal force, which results in a rocking motion or rotor wobble. This condition can be simulated by placing a pencil on a table, then at one end pushing the side of the pencil with one finger. At the same time, push in the opposite direction at the other end. The pencil will tend to rotate end-over end. This end-over-end action causes two imbalances ''orbits,'' both 180 degrees out of phase, resulting in a ''wobble'' motion.

Dynamic Imbalance Combinations of Static and Couple

Visualize a rotor that has only one imbalance in a single plane. Also visualize that the plane is not at the rotor's center of gravity but is off to one side.

Although there is no other source of couple, this force to one side of the rotor not only causes translation (parallel motion caused by pure static imbalance) but also causes the rotor to rotate or wobble end-over-end as from a couple. In other words, such a force would create a combination of both static and couple imbalance. This again is dynamic imbalance.

In addition, a rotor may have two imbalance forces exactly 180 degrees opposite to each other. However, if the forces are not equal in magnitude, the rotor has a static imbalance in combination with its pure couple. This combination is also dynamic imbalance.

Another way of looking at it's to visualize the usual rendition of dynamic imbalance-imbalance in two separate planes at an angle and magnitude relative to each other not necessarily that of pure static or pure couple.

For example, assume that the angular relationship is 80 degrees and the magnitudes are 8 units in one plane and 3 units in the other. Normally, you would simply balance this rotor on an ordinary two-plane dynamic balancer and that would be satisfactory. But for further understanding of balancing, imagine that this same rotor is placed on static balancing rollers, whereby gravity brings the static imbalance components of this dynamically out-of-balance rotor to the 6 o'clock position.

The static imbalance can be removed by adding counterbalancing weights at the 12 o'clock position. Although statically balanced, the two remaining forces result in a pure couple imbalance. With the entire static imbalance removed, these two forces are equal in magnitude and exactly 180 degrees apart. The couple imbalance can be removed, as with any other couple imbalance, by using a two-plane dynamic balancer and adding counterweights.

Note that whenever you hear the word ''imbalance,'' mentally add the word ''dynamic'' to it. Then when you hear ''dynamic imbalance,'' mentally visualize ''combination of static and couple imbalance.'' This will be of much help not only in balancing but in understanding phase and coupling misalignment as well.

BALANCING

Imbalance is one of the most common sources of major vibration in machinery.

It is the main source in about 40% of the excessive vibration situations. The vibration frequency of imbalance is equal to one times the rpm ( 1 x rpm ) of the imbalanced rotating part.

Before a part can be balanced with the vibration analyzer, certain conditions must be met:

_ The vibration must be caused by mechanical imbalance; and ,

_ Weight corrections can be made on the rotating component.

To calculate imbalance units, simply multiply the amount of imbalance by the radius at which it's acting. In other words, 1 oz. of imbalance at a 1-in. radius will result in 1 oz.-in. of imbalance. Five ounces at a 0.5-in. radius results in 2.5 oz.-in. of imbalance. (Dynamic imbalance units are measured in ounce-inches [oz.-in.] or gram-millimeters [g-mm].) Although this refers to a single plane, dynamic balancing is performed in at least two separate planes. Therefore the tolerance is usually given in single-plane units for each plane of correction.

Important balancing techniques and concepts to be discussed in the sections to follow include in-place balancing, single-plane versus two-plane balancing, precision balancing, techniques that make use of a phase shift, and balancing standards.

IN-PLACE BALANCING

In most cases, weight corrections can be made with the rotor mounted in its normal housing. The process of balancing a part without taking it out of the machine is called in-place balancin g. This technique eliminates costly and time-consuming disassembly. It also prevents the possibility of damage to the rotor, which can occur during removal, transportation to and from the balancing machine, and reinstallation in the machine.

SINGLE-PLANE VERSUS TWO-PLANE BALANCING

The most common rule of thumb is that a disc-shaped rotating part usually can be balanced in one correction plane only, whereas parts that have appreciable width require two-plane balancing. Precision tolerances, which become more meaningful for higher performance (even on relatively narrow face width), suggest two-plane balancing. However, the width should be the guide, not the diameter-to-width ratio.

For example, a 20-inch-wide rotor could have a large enough couple imbalance component in its dynamic imbalance to require two-plane balancing. (Note: The couple component makes two-plane balancing important.) Yet if the 20-inch width is on a rotor of large enough diameter to qualify as a ''disc-shaped rotor,'' even some of the balance manufacturers erroneously would call for a single-plane balance.

It is true that the narrower the rotor, the less the chance for a large couple component and therefore the greater the possibility of getting by with a single-plane balance. For rotors over 4-5 in. in width, it's best to check for real dynamic imbalance (or for couple imbalance).

Unfortunately, you can't always get by with a static- and couple-type balance, even for very narrow flywheels used in automobiles. Although most of the flywheels are only 1-1.5 in. wide, more than half have enough couple imbalance to cause excessive vibration. This obviously is not caused by a large distance between the planes (width) but rather by the fact that the flywheel's mounting surface can cause it to be slightly cocked or tilted. Instead of the flywheel being 90 degrees to the shaft axis, it may be perhaps 85 to 95 degrees, causing a large couple despite its narrow width.

This situation is very common with narrow and disc-shaped industrial rotors such as single-stage turbine wheels, narrow fans, and pump impellers. The original manufacturer often accepts the guidelines supplied by others and performs a single-plane balance only. By obtaining separate readings for static and couple, the manufacturer could and should easily remove the remaining couple.

An important point to remember is that static imbalance is always removed first.

In static and couple balancing, remove the static imbalance first and then remove the couple.

PRECISION BALANCING

Most original-equipment manufacturers balance to commercial tolerances, a practice that has become acceptable to most buyers. However, because of frequent customer demands, some of the equipment manufacturers now provide precision balancing. Part of the driving force for providing this service is that many large mills and refineries have started doing their own precision balancing to tolerances considerably closer than those used by the original-equipment manufacturer. For example, the International Standards Organization (ISO) for process plant machinery calls for a G6.3 level of balancing in its balancing guide. This was a calculated based on a rotor running free in space with a restraint vibration of 6.3 mm/sec (0.25 in./sec) vibration velocity.

Precision balancing requires a G2.5 guide number, which is based on 2.5 mm/sec (0.1 in./sec) vibration velocity. As can be seen from this, 6.3 mm/sec (0.25 in./sec) balanced rotors will vibrate more than the 2.5 mm/sec (0.1 in./sec) precision balanced rotors. Many vibration guidelines now consider 2.5 mm/sec (0.1 in./sec) ''good,'' creating the demand for precision balancing. Precision balancing tolerances can produce velocities of 0.01 in./sec (0.3 mm/sec) and lower.

It is true that the extra weight of non-rotating parts (i.e., frame and foundation) reduces the vibration somewhat from the free-in-space amplitude. However, it's possible to reach precision balancing levels in only two or three additional runs, providing the smoothest running rotor. The extra effort to the balance operator is minimal because he already has the ''feel'' of the rotor and has the proper setup and tools in hand. In addition, there is a large financial payoff for this minimal extra effort because of decreased bearing and seal wear.

TECHNIQUES USING PHASE SHIFT

If we assume that there is no other source of vibration other than imbalance (i.e., we have perfect alignment, a perfectly straight shaft, etc.), it's readily seen that pure static imbalance gives in-phase vibrations and pure coupled imbalance gives various phase relationships. Compare the vertical reading of a bearing at one end of the rotor with the vertical reading at the other end of the rotor to determine how that part is shaking vertically. Then compare the horizontal reading at one end with the horizontal reading at the other end to determine how the part is shaking horizontally.

If there is no resonant condition to modify the resultant vibration phase, then the phase for both vertical and horizontal readings is essentially the same, even though the vertical and horizontal amplitudes don't necessarily correspond. In actual practice, this may be slightly off because of other vibration sources such as misalignment. In performing the analysis, what counts is that when the source of the vibration is primarily from imbalance, then the vertical reading phase differences between one end of the rotor and the other will be very similar to the phase differences when measured horizontally. For example, vibrations 60 degrees out of phase vertically would show 60 degrees out of phase horizontally within 20%.

However, the horizontal reading on one bearing will not show the same phase relationship as the vertical reading on the same bearing. This is caused by the pickup axis being oriented in a different angular position as well as the phase adjustment caused by possible resonance. For example, the horizontal vibration frequency may be below the horizontal resonance of various major portions of machinery, whereas the vertical vibration frequency may be above the natural frequency of the floor supporting the machine.

First, determine how the rotor is vibrating vertically by comparing ''vertical only'' readings with each other. Then determine how the rotor is vibrating horizontally.

If the rotor is shaking horizontally and vertically and the phase differences are relatively similar, then the source of vibration is likely to be imbalance. However, before coming to a final conclusion, be sure that other l x rpm sources (e.g., bent shaft, eccentric armature, misaligned coupling) are not at fault.

BALANCING STANDARDS

The ISO has published standards for acceptable limits for residual imbalance in various classifications of rotor assemblies. Balancing standards are given in ounce-inches or pound-inches per pound of rotor weight or the equivalent in metric units (gram-millimeters per kilogram). The ounce-inches are for each correction plane for which the imbalance is measured and corrected.

Caution must be exercised when using balancing standards. The recommended levels are for residual imbalance, which is defined as imbalance of any kind that remains after balancing.

Ill. 0-8.1 and Table 8.1 are the norms established for most rotating equipment.

Additional information can be obtained from ISO 5406 and 5343. Similar standards are available from the American National Standards Institute (ANSI) in their publication ANSI S2.43-1984.

So far, there has been no consideration of the angular positions of the usual two points of imbalance relative to each other or the distance between the two correction planes. For example, if the residual imbalances in each of the two planes were in phase, they would add to each other to create more static imbalance.

Most balancing standards are based on a residual imbalance and don't include multi-plane imbalance. If they are approximately 180 degrees to each other, they form a couple. If the distance between the planes is small, the resulting couple is small; if the distance is large, the couple is large. A couple creates considerably more vibration than when the two residual imbalances are in phase. Un-fortunately, there is nothing in the balancing standards that takes this into consideration.

Ill. 0-8.1 Balancing standards. Residual imbalance per unit rotor weight. Balancing of Rotating Machinery There is another problem that could also result in excessive imbalance-related vibration even when the ISO standards have been met. The ISO standards call for a balancing grade of G6.3 for components such as pump impellers, normal electric armatures, and parts of process plant machines. This results in an operating speed vibration velocity of 6.3 mm/sec (0.25 in./sec) vibration velocity.

However, practice has shown that an acceptable vibration velocity is 0.1 in./sec and the ISO standard of G2.5 is actually required. As a result of these discrepancies, changes in the recommended balancing grade are expected in the future.

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Table 8.1 Balance Quality Grades for Various Groups of Rigid Rotors

Balance | Quality | Grade | Type of Rotor

G4,000 Crankshaft drives of rigidly mounted slow marine diesel engines with uneven number of cylinders G1,600 Crankshaft drives of rigidly mounted large two-cycle engines G630 Crankshaft drives of rigidly mounted large four-cycle engines; crankshaft drives of elastically mounted marine diesel engines G250 Crankshaft drives of rigidly mounted fast four-cylinder diesel engines G100 Crankshaft drives of fast diesel engines with six or more cylinders; complete engines (gasoline or diesel) for cars and trucks.

G40 Car wheels, wheel rims, wheel sets, drive shafts; crankshaft drives of elastically mounted fast four-cycle engines (gasoline and diesel) with six or more cylinders; crankshaft drives for engines of cars and trucks G16 Parts of agricultural machinery; individual components of engines (gasoline or diesel) for cars and trucks G6.3 Parts or process plant machines; marine main-turbine gears; centrifuge drums; fans; assembled aircraft gas-turbine rotors; fly wheels; pump impellers; machine-tool and general machinery parts; electrical armatures G2.5 Gas and steam turbines; rigid turbo-generator rotors; rotors; turbo-compressors; machine-tool drives; small electrical armatures; turbine-driven pumps G1 Tape recorder and phonograph drives; grinding-machine drives

G0.4 Spindles, disks, and armatures of precision grinders; gyroscopes

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