Updating Shaft Alignment Knowledge

Defining permissible misalignment tolerances determines a reliability-focused approach.

In the majority of industrial facilities, worker and technician resources are probably stretched to the limit. They might be looking for ways to simplify some traditional work processes and procedures. Or the facilities may have had an experience that reinforces the contention that high-tech tools are not always the answer, and that back-to-basics thinking has considerable merit.

While no reasonable and experienced reliability professional will take issue with these statements, industry must be cautioned against drawing the wrong conclusions. A recent example of wrong conclusions involves claims that rotating equipment alignment is sufficiently accurate as long as the shaft centerlines in their cold, standstill condition are within 0.002 in. (0.05 mm) of each other. Those who believe this and blindly follow this questionable advice may soon find themselves among the repair-focused dinosaurs who are struggling to survive.

On the other hand, those who update their knowledge of shaft alignment and alignment tolerances are on the way to becoming reliability focused. Indications are that only the reliability-focused facilities will be around in a few years.

Expressing tolerances
The only correct way to express shaft alignment tolerances is in terms of alignment conditions at the coupling. This article will describe several ways to do this. It is incorrect to describe alignment tolerances in terms of correction values at the machine feet, and this will be examined also.

When two machines are directly coupled with a flexible coupling, any misalignment between their centerlines of rotation can result in vibration which, depending on its severity, can produce premature wear or catastrophic failure of bearings, seals, the coupling itself, and other rotating components. Misalignment of the centerlines of rotation has long been recognized as one of the leading causes of machinery damage.

Decades of well-documented observations attest that misalignment is responsible for huge economic losses. The more misalignment, the greater the rate of wear, likelihood of premature failure, and loss of efficiency of the machine. Misaligned machines absorb more energy and thus consume more power.

However, even excellent alignment of the shaft centers of rotation does not guarantee absence of vibration because there is still the possibility of imbalance of rotating components. Structural resonance, fluid flow turbulence and cavitation, or vibration from nearby running machines that is transmitted to adjacent machines through either foundation or piping could also exist.

Foot alignment does not work
Absolute perfection in the alignment of shafts is not realistically attainable, nor is it needed. The issue is quantification of alignment quality and allowable deviation—the alignment tolerance.

Misalignment is defined by visualizing the shaft centerlines of rotation as two straight lines in space. The trick is to get them to coincide to form one straight line. If they do not, then there must exist either offset misalignment or angular misalignment (Fig. 1), or a combination of both.

Since the shafts exist in three-dimensional space, these misalignments can exist in any direction. Dividing this space into two planes, the vertical and the horizontal, and describing the specific amount of offset and angularity that exists in each of these planes simultaneously helps define the four specific conditions of misalignment—vertical offset (VO), vertical angularity (VA), horizontal offset (HO), and horizontal angularity (HA). These conditions are described at the location of the coupling because that is where harmful machinery vibration is created whenever misalignment exists.

The magnitude of an alignment tolerance (the description of desired alignment quality) must be expressed in terms of these offsets and angularities, or the sliding velocities resulting from them. Attempts to describe misalignment in terms of foot corrections alone do not take into account the size, geometry, or operating temperature of a given machine. Accepting the simple foot corrections approach can seriously compromise equipment life and has no place in a reliability-focused facility. An illustration of the fallacy of the foot correction approach will be given later.

How much vibration and efficiency loss will result from the misalignment of shaft centers depends on shaft speed and coupling type. Acceptable alignment tolerances are functions of shaft speed and coupling geometry. High-quality flexible couplings are designed to tolerate more misalignment than what is good for the machines involved. Bearing load increases with misalignment, and bearing life decreases as the cube of the load increase, i.e., doubling the load will shorten bearing life by a factor of eight.

Why would high-quality flexible couplings be generally able to accommodate greater misalignment than what is good for the connected machines? A large percentage of machines must be deliberately misaligned—sometimes significantly—in the cold and stopped condition. As they reach operating speeds and temperatures, thermal growth is anticipated to bring the two shafts into alignment.

Alignment case history
A refinery has a small foot-mounted steam turbine enveloped in insulating blankets. The operating temperature of the steel casing is 455 F, and the distance from centerline to the bottom of the feet is 18 in. The turbine drives an ANSI pump with a casing temperature of 85 F; its centerline-to-bottom-of-feet distance is also 18 in. Both initially started up at the same ambient temperature. The differential in their growth is (0.0000065 in/in/deg F) × 18 in. × (455 – 85) deg F = 0.043 in.

If these two machines had their shafts aligned center to center, this amount of offset would cast the equipment train into the frequent failure category. Using the 80/20 rule, it would be safe to assume that 20 percent of the machinery population accounts for 80 percent of the maintenance money. This pump train would be in the 20 percent group.

As mentioned earlier, aligning center to center without paying attention to thermal growth is one of the factors that keeps practitioners in the repair-focused category.

Permissible tolerances
There are a number of acceptable ways to describe misalignment at the coupling and to define permissible misalignment tolerances. While many of these are of equal merit, describing alignment tolerances in terms of foot corrections is not acceptable in a reliability-focused environment.

Offset and angularity at the coupling (for short couplings) is one of the most common ways of correctly defining alignment tolerances. The offset tolerance simply describes the maximum separation that can exist between two machine shafts at a specific location along their shaft axes, usually the coupling center. The angularity describes the rate at which the offset between the shaft centerlines may change as they travel along the axes of the shafts.

The angularity may be described either directly, as an angle in terms of mils/in. (or milliradians), or as a gap difference at a particular coupling diameter. The latter method is popular because it relates directly to what the mechanic can detect with his feeler gages between the coupling faces. A modern laser shaft alignment system measures the angle between shaft centerlines; such a system can also be set to describe this angle as a gap difference at any desired diameter.

This approach can have two different interpretations, however. If the permissible offset between the driver and driven shafts is X, does this mean X in any direction, or X individually in both the horizontal and vertical planes? These two alternatives are not the same. The first example, vector tolerance, is more conservative. The second approach, standard tolerance, is the more common approach. If it is not desired to have more than X of offset to exist between the shafts in any direction, then standard tolerances should not be used. Doing so would, in some circumstances, lead to greater-than-intended offsets. Figure 2 illustrates this point.

Figure 2 (left) illustrates a case where applying standard tolerances results in an offset of 2.5 mils horizontally and 2.7 mils vertically that is deemed acceptable because the permissible limit for either of these offsets individually is 3.0 mils. However, the actual offset 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 (Fig. 2 right).

A good laser shaft alignment system will allow the user to make this distinction and to specify exactly which type of tolerance is desired.

Table 1 shows the values most widely accepted as the standard industry norm for short couplings.

Spacer coupling tolerances are generally expressed as limits to the angle that may exist between each machine shaft and the spacer piece between them. Since the spacer piece (or spool piece) connects to each of the machine shafts at either end, there should be no offset between the spacer and each of the machine shafts. Only the maximum angle allowed between the spacer shaft and each of the connected machine shafts needs to be specified. This angle may be specified directly in mils/in. (or milliradians) or in terms of the offset that each individual angle between spacer and machine shaft projects to the opposite end of the spacer. The first way is called the angle-angle method (sometimes called the alpha-beta method), and the second way is called the offset-offset (or offset A-offset B) method (see Fig. 3).

Since most flexible couplings have two flex planes (or points of articulation), the spacer coupling tolerances may safely be used, even with flex couplings. The best criterion to make the distinction is the relation between the diameter of the flex planes and the distance between them. Whenever the distance between flex planes (span) is greater than the diameter, it is called a spacer coupling. This will make achieving tolerances easier when performing alignment corrections in the field. Table 2 presents the values most widely accepted as the standard industry norm for spacer couplings.

Sliding velocity tolerance is the permissible limit of the velocity that the moving elements in a flexible coupling may attain during operation. This can be related to the maximum permissible offset and angularity through the formula:

Maximum allowable component sliding velocity = 2 X d X r X a X π

where:
d = coupling diameter
r = revolutions/time
a = angle in radians

When offset and angularity exist, the flexible or moving coupling element must travel by double the amount of the offset and angularity every half rotation. Because the speed of rotation is defined, so must be the velocity that is achieved by the moving element in accommodating the misalignment as the shaft turns. When the permissible sliding velocity is limited, the offset and angularity (in any combination) that can exist between the coupled shafts as these turn is also limited by definition.

For 1800 rpm, this limit is about 1.13 in./sec for excellent alignment, and 1.89 in./sec for acceptable alignment. A good laser alignment system will let reliability-focused users apply this approach as well.

Tolerances expressed as corrections at the machine feet are not acceptable. It is impossible to define the quality of the alignment between rotating shaft centerlines in terms of correction values at the feet alone, unless the exact dimensions related to these specific correction values are specified each time. This approach is too cumbersome and error-prone, since two machines will rarely share the same dimensions between the coupling and the feet, and between the feet themselves.

A tolerance that describes only a maximum permissible correction value at the feet without references to the operative dimensions involved makes no sense because the same correction values can yield vastly different alignment conditions between the machine shafts with different dimensions. Such a tolerance ignores the effects of rise over run, which is essentially what shaft alignment is all about. Such a tolerance does not take into account the type of coupling or the rotating speed of the machines.

These alignment tolerances specified generically in terms of foot corrections can have two equally bad 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 representing excellent alignment between the shafts. In the first scenario, the aligner may stop correcting his alignment before the machines are properly aligned. In the second case, the aligner may be misled into continuing to move machines long after they have already arrived in tolerance.

A couple of examples illustrate the fallacy of the generic foot correction approach. Assume that the specified alignment tolerance for an 1800 rpm machine is defined as a maximum 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—in tolerance by this method. If the distance between the feet is 8 in., this would imply an existing angular misalignment of 0.5 mil/in. If the distance from the front bearing to the coupling center is 10 in., the offset between the machine shafts at the coupling would be +7.0 mils.

This offset is considerably in excess of the ±3 mils of offset (either standard or vector) that is considered the maximum acceptable for an 1800 rpm machine at the coupling. Yet, with the improperly specified foot correction tolerances, this alignment would be in tolerance. This is a classic example of where small correction values at the feet do not necessarily reflect good alignment at the coupling (Fig. 4).

In addition, the opposite scenario is just as likely to occur. Assume a large machine (such as a diesel engine) is running at 1200 rpm with distance between the feet of 80 in. The distance from front feet to the coupling is 30 in. The machine has misalignment requiring foot corrections of –8 mils at the front feet and –26 mils at the back feet. The misalignment at the coupling is only 1.25 mils of offset and only 0.225 mil/in. of angularity. In reference to the coupling, both of these alignment conditions are already much better than required by the standard industry norms for 1200 rpm. However, using the improperly specified tolerance values of ±2 mils at the feet, the aligner would be misled into working much harder and longer than necessary to bring the machines to these values (Fig. 5).

Note: Figures 2, 4, and 5 were created with the assistance of Alignment Explorer software by Prüftechnik, Ismaning, Germany, and input from Ludeca, Miami, FL. MT

Heinz P. Bloch, P.E. is a consulting engineer with over 40 years of experience in chemical process plants and oil refineries. He may be reached at 5459 Ponderosa Dr., West Des Moines, IA 50266; (515) 225-0668; e-mail hpbloch@mchsi.com

MISALIGNMENT

Fig. 1. Misalignment can be offset (left), angular (right, or a combination of both.

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TOLERANCE EVALUATION

Fig. 2. Depending on circumstances, using the standard tolerance (left) instead of vector tolerance (right) can lead to greater than intended offsets.

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SPACER COUPLING TOLERANCES

Fig. 3. Specifying the maximum angle allowed between the space shaft and each of the connected machine shafts sets the spacer coupling tolerances.

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GENERIC FOOT CORRECTION

Fig. 4. Small correction values at the feet do not necessarily reflect good alignment at the coupling.

Fig. 5. Improperly specified tolerance values may show misalignment where none exists.

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