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Metallic flexible element couplings use various metal alloys to achieve high power density. Their flexibility is gained through either loose fitting parts which roll or slide against one another (gear, grid, chain) -sometimes referred to as "mechanical flexing"-- or through flexing/bending of a membrane (disc, flex link, diaphragm, beam, bellows).
Those with moving parts generally are less expensive, but need to be lubricated and maintained. Their primary cause of failure is wear, so overloads generally shorten their life through increased wear rather than sudden failure. Membrane types are generally more expensive do not need lubrication and require little maintenance. Generally, their primary cause of failure is fatigue. They can fail quickly (in a short cycle time) when overloaded. They can be very long lived if kept within their load ratings. They may even out last their connected equipment.
Lubricated Coupling Types
Gear-couplings are the king of coupling types. They can do things that many other couplings cannot do, can only do with difficulty or with expensive modifications and de-rating. Gear couplings are more power intensive, offer more modifications, and a wider size, torque bore range
than any other type, and can perform at extremely high speeds. Gear
couplings have axial slide capability, low speed high torque capability,
shifter capability and spindle capability not found in other couplings.
They are easily modified for shear pin service, floating shaft type, vertical
type, insulated type, limited end float, and can have a brake drum or
disc added. While those latter items may be available on other
couplings, it is usually easier and less costly to modify the gear
coupling. With all these advantages the gear coupling is used on twice
as many applications versus the nearest competitor type.
Gear couplings can also perform at extremely high rates of speed. As
implied by the name, gear couplings use the meshing of gear teeth to
transmit the torque and to provide for misalignment. External gear teeth
are cut on the circumference of the hub. Both toothed hubs fit inside the
ends of a tubular sleeve that has matching gear teeth cut around its
interior circumference, with each tooth extending axially the full length of
the sleeve. Hub and sleeve teeth mesh, so torque transfers from the
driving hub's teeth to the sleeve teeth and back to the driven hub's
Gear couplings achieve their misalignment capability through backlash
in the teeth, crowning on the tooth surfaces, and major diameter fit.
Backlash is the looseness-of-fit that results from gear teeth being narrower
than the gaps between the teeth. In addition to contributing to the
misalignment capabilities, the backlash provides space for the lubricant.
The loose fit provides misalignment capability by allowing the sleeve to
shift off-axis without binding against the hub teeth. Some gear couplings
have more backlash than others. Those with the least (roughly one-half of the backlash present in those with the most) are known as "minimum backlash" couplings. Some users prefer this type, most prefer normal backlash.Crowning, or curving the surface of the hub teeth, further enhances this
capability. The crowning can include tip crowns, flank crowns, and
chamfers on the sharp edges. This also helps improve tooth life by
broadening the contact area along the "pitch line" (where the teeth mate
and transfer torque), thereby reducing the pressure of torque forces. In
addition, it prevents the sharp squared edges of the tooth from digging
in and locking the coupling. Vari-Crown, which varies the curvature
radius along the tooth flank, maintains greater contact area between
teeth during misalignment compared with standard crowning, and
reduces those stresses that cause wear. Note that crowning applies to
hub teeth only; sleeve teeth are straight except for a chamfer on the
minor diameter edge.
While the hub and sleeve teeth are cut to fit loosely side to side, they
are cut to fit closely where the tip diameter of the hub teeth meet the
root diameter of the gaps between the sleeve teeth. That is called a
major diameter fit. When the coupling is not rotating, those two surfaces
rest upon each other if it is a horizontal installation. Minor diameter fits
(where the tips of the sleeve teeth meet the root diameter of the hub
teeth) are purposely avoided, because a close fit here would preclude
suitable misalignment capability and torque transmission capability.
It was noted earlier that gear couplings are power intensive. That
means more torque transmitted per pound of coupling weight and per
cubic inch of space consumed than other couplings. In many cases the
gear coupling has more torque capability than the shaft can transmit.
The resulting relatively small size of the gear coupling allows the addition
of attachments without having the coupling grow to impracticable
proportions. It also allows the OEM designer more latitude to locate the
coupling in small, out-of-the-way places with confidence that it will be
reliable. Gear couplings eventually wear, but rarely to a catastrophic
failure. They can be sized to make sure that wear life is consistent with
the rest of the machine design.
Gear coupling sleeves can be a single piece, termed a "continuous sleeve", or can be split laterally (radially) into two half sleeves, one on
each hub. The split version is termed a "flanged sleeve", because each
half has a flanged end, drilled for bolt holes, which allows them to be
Because the continuous sleeve needs neither flanges nor bolts, it provides
the advantage of making the coupling lighter and smaller in diameter
than comparably rated flange types. With that comes lower inertia
values, which helps lighten motor load during start-up. Bolt stress,
which can be a weak point in some applications, is eliminated. The
absence of bolts is an advantage in high-speed applications, because
bolts add potential points of unbalance and bolted connections can be
another point of non-concentricity.
When two halves of a flanged sleeve are bolted together, the bolting
becomes an important part of the power transmission path. Best
designs have the power transmitted across the face by friction, in which
case the bolts simply provide enough clamping force to provide face
friction. Other designs could allow the bolts to carry the load in shear,
but those are in the minority. Both cases require a proper analysis of the
multiple loads on the bolts. In addition the bolt bodies may provide the
centering action to pilot the two halves of the coupling.
Bolts can either be exposed, or shrouded for safety reasons. However,
with the advent of OSHA coupling-guard requirements, shrouding
becomes unnecessary. The two types also have different windage loss and that affects high speed applications. Windage losses cause a heat generation inside the coupling guard.Note that flange bolts are specially made for their purpose, and should
never be replaced with common hardware-store bolts. Flanged sleeve
gear couplings built to American Gear Manufacturers Association
(AGMA) dimensional standards will mate half-for-half with all other gear
couplings made to those same standards. While AGMA standards are
U.S. based, many European manufacturers build to match the dimensions.
However, matching dimensions include the interface only, such
as outside flange diameter, number of boltholes, bolt hole size, bolt circle,
and flange thickness. Although length-through-bore of the hub is
often identical as well, torque and bore capability are likely to be different
and should be compared carefully.
Flex Planes and Misalignment Capability
Planes of flexibility ("flex planes") are those pivot points along the shaft-to-shaft connection where rigid components engage but can move independently
of each other The standard gear coupling (two toothed hubs
engaging opposite ends of the same rigid sleeve) has two flex planes,
one at each hub-to-sleeve gear mesh. When both flex planes work
together in series, flexing in the same direction, they give the gear coupling
an angular misalignment capability of up to 1½° at each flex plane.
This standard configuration is called "full flex" or "double engagement"
The full-flex gear coupling, with two flex planes in series flexing in opposite
directions, allows for parallel (radial) misalignment of 0.055 to 0.165
inches in standard models with short sleeves. The longer the sleeve
(i.e. the greater the axial distance from one flex plane to the other), the
greater the parallel misalignment. The greatest parallel capability results
from floating shaft, spacer and spindle versions, described later, which
greatly lengthen the distance between flex planes.
Gear couplings can be configured with only one flex plane, for applications
where parallel misalignment capability is unwanted. In flanged type
couplings, this is accomplished by using a single-piece flanged hub with
no teeth, as the rigid half, bolted to a flexible half that uses a standard
flanged sleeve with teeth and a standard hub with crowned teeth. These
are called "flex-rigid", "single-engagement" or sometimes "half couplings".
In continuous-sleeve couplings, a flex-rigid configuration is
accomplished by mating the sleeve at the rigid end with a hub having
straight teeth that fits into the sleeve like a spline shaft into a spline hub.
While the full flex design is the most popular in gear couplings, flex rigid
designs are often useful in systems with three bearings or floating
shafts. Sometimes one flex-rigid coupling is used in series with another
flex-rigid coupling at a distance to allow much more parallel misalignment.
While gear couplings will normally provide from ½° to 1½° of angular
misalignment per flex half, they can be designed for up to 6° with
reduced load capability and with accommodating grease seals.
Gear couplings naturally accommodate axial (in-out) shaft movement better than other competing designs, because their hub teeth easily slide along their sleeve teeth with no effect on coupling operation or torque load capability. Axial movement often results from thermal expansion/contraction of the shaft, as in hot applications, or a rotor seeking its magnetic centers (floating rotor). Thrust bearings can limit or prevent shaft movement at the coupling end, but if positioned at the far end of the machine, they can force the shaft movement back toward the coupling. The amount of axial displacement the gear coupling can handle depends primarily on the length of the sleeve, and specials are available for long sliding application.
The Gear Coupling Tooth
The gear coupling tooth has evolved over many years. The first gear couplings had straight teeth, and depended purely on backlash to
achieve misalignment. Later improvements included tooth crowning that
increased misalignment capability and coupling life. The original tooth
form followed the spur gear form with modification. Various pressure
angles were used that walked the line between life and strength. The
40° pressure angle tooth was chosen for strength. It proved to have
problems with wear life and with reactionary loading on the machinery.
Eventually the 20° pressure angle tooth became the standard, and it still
is the standard. Some 25° teeth are used to achieve added strength for
special designs. The additional strength of today's materials alleviates
the need for 40° teeth and still provides low sliding friction.
The gear coupling tooth, like the spline tooth, is not a full height tooth.
Where the spline is 50% height, the gear coupling tooth is about 80%.
Gear coupling teeth do not need full height because the torque load is
carried at the pitch line of the tooth and many teeth are in contact with
each other in the hub and the sleeve to carry the load. The number of
teeth in contact is a function of the true form of the teeth. If all teeth in
the hub and sleeve are identical the maximum number will be in contact.
As the teeth wear into place the more teeth come into contact.
Therefore initial tooth wear makes the coupling stronger, but can
increase the friction loading too.
The strength of the gear tooth is the subject of many questions in determining
the amount of load to be carried. The tooth is the strongest of all
the elements of a gear coupling. The tooth strength is calculated as a
bending moment at the root of the tooth, the shear strength at the pitch
line, and the Hertzian loading at the contact surface. All of these forces
The most likely failure mode of a gear coupling tooth is that which
comes from wear rather than any other factor. As the teeth wear, they
move from being the strongest element to being the weakest element.
Severe misalignment that causes a lock up of the teeth will also result
in premature failure. Most other loading on the coupling will not result in
Lubricant must always be available in the tooth mesh. The lack of lubricant
will, of course, cause the coupling to fail almost instantly. The gear
coupling is fitted together so as to prevent the lubricant from leaking.
Most gear couplings are lubricated with grease. The sleeve to hub interface
at the boundaries will need elastomer O-rings, gaskets, or
labyrinths to prevent grease leakage. (Note that O-ring material might
limit the coupling's ambient temperature capability.) When oil lubrication
is used, it is usually a continuous flow through the tooth mesh, but can
be a batch lube in some applications. Oil lubrication is a special case.
Misalignment may allow grease to leak out the seal surface, or some
modifications may need a wiper seal rather than an O-ring. One type of
flange coupling uses a high misalignment seal with more flex than the
regular seal. The seals can be held in place by several means. The Oring
is the simplest; it fits into a groove in the sleeve.
The continuous sleeve coupling seal is held in place by a spiral ring.
The seal has stiffeners molded into the inside face. It is a U or C shape
that stays closed under load. It also provides the movement limit for the
coupling and is actually rated to withstand an axial force.
Sometimes the seal holder is bolted to the coupling sleeve. This is
always the case on couplings larger than size 9. It makes the assembly
of the coupling to the shaft easier, and makes replacement of seals easier.
The couplings with bolt on seal carriers are designated heavy duty
(HD). Flange series couplings size 7 through 9 can be either the "HD"
version or the plain version.
Remember that the coupling grease is not ordinary grease but is specially
formulated so the oils do not separate from the soaps. The result
is that the lubricant is contained within the needed space and sludge is
not allowed to accumulate. Oil and soaps separate in ordinary lubricants
because of centrifugal forces on the heavier particles. Use only coupling
grease for best results.
Variations to Gear Couplings
1. Fill the Space between Shafts
Couplings often must fill a space between shafts as one of their primary attributes. It would seem a simple enough task, but not all couplings
offer flexibility doing that job. This is another reason why the gear coupling
is very popular.
2. "BSE" Dimension
The distance between shaft ends (BSE) will vary with different machine systems to accommodate design standards, product line alternatives,different motor frames and maintenance needs. The "BSE" dimension is important for all couplings. Gear couplings have the advantage of allowing a variable "BSE". That variation can be achieved by machining the hub face or can be achieved by reversing one or both of the hubs. An infinite number of possibilities can be obtained from catalog minimum to catalog maximum. Note this gap (BSE) does not always affect the distance between flex planes unless the hubs are reversed. A combination of facing and reversing is possible too. All couplings have a certain "BSE" dimension variability, but few are able to tolerate as great a variance as gear couplings can.
3. Spacer Couplings
Spacer couplings consist of two flexible hub and sleeve assemblies i.e. a half coupling on both the driving and driven shafts. These are connected
by a tubular center section of various lengths that can easily be
removed to allow space for removal of the hub or other components on
one side of the system without disturbing the hub or component mountings
on the other side. The tubular center section can have flanged
ends for bolting to hub flanged sleeves, or toothed ends that mate with
hubs using continuous sleeves. Spacers are built to the standards of the
rotating machinery builders. Pumps have several standard spacers such
as 3½ inches, 7 inches and others. Compressors could have a different
set of standard spacers. Spacers can serve to separate the flex planes
and can be part of the torsional tuning of a coupling.
They have practical
limits on length in regard to cost, weight and critical speed. The
flanged hollow tube is machined to varying tolerances depending on
speed and balance. As the tube gets longer, deflection of the unsupported
center section forces the cylinder walls to be made thicker. As the
walls get thicker the cost grows more and so does the weight. The
weight then reduces the critical speed. That is a cross combination of
events that eventually makes the spacer a poor choice. When the spacer
becomes impractical, the next step is to use a floating shaft coupling
to achieve the necessary spacing.
4. Floating Shafts
Floating shaft couplings consist of flex rigid couplings on both driving and driven shafts connected by a piece of solid shafting between the couplings. Usually the coupling hubs on the equipment ends are rigid while the two center hubs connected to the floating shaft are flexible.
While these two can be used to provide service spacing, the primary reason for a long floating shaft is to allow for greater radial misalignment between shafts. The secondary reason is to reach a long distance between the driver and the rotating equipment. Weight and critical speed are important considerations for floating shafts. They are found on bridge cranes and steel rolling mills.
The couplings and center shaft are designed as a unit to suit their specific application. The parameters include the usual torque and bore, but must include length and speed because, as in any spacer, critical speed and deflection are interrelated. These issues may require a larger diameter center shaft to reduce deflection. In that case the rigid hubs are on the floating shaft, taking advantage of the rigid hub's greater bore capability, to accommodate the oversized center shaft.
Otherwise, the center shaft may need to be necked down (reduced) to fit a flex coupling hub. The rigid hub could also be placed on the outside to fit a shaft that is made larger than is necessary to carry torque, as would be the case with bending problems. The center shaft would be smaller to carry torque only and thus fit the flex hub. When the flex hub is on the center shaft it is called a marine style coupling. When the flex hub is on the equipment shaft it is called a reduced moment style. The floating shaft designer must always balance the effects of weight (which causes deflection) and diameter (which determines torque capacity and resists deflection but increases weight and cost).
5. Limited End Float Couplings
Gear couplings can be modified to allow shaft growth in the axial direction or to limit movement in the axial direction. Limiting the movement calls for a plate and possibly a button to be inserted between the coupling halves. As the shaft tries to move in the axial direction, it is stopped after moving a predetermined distance.
These are called limited end float couplings. They are necessary with sleeve bearing motors, a design commonly found in larger sizes of 200 horsepower or more. The same plates and buttons are used on vertical couplings as explained below.
In addition to thermal growth, gear couplings can be arranged to slide great distances. Extra long sleeves enable the hub to slide 10 inches or more, either at rest or while in operation, to serve applications where equipment must be temporarily removed from the system and the coupling is the most suitable point of movement. Refiners, Jordan machines, and roll winders found in paper mills utilize this sliding capability. The Jordan coupling is a special variation that can move its hub relative to its shaft with a clamping mechanism.
Two dimensions are important when considering the slider coupling. One is the minimum BSE and the other is the total amount of slide. Those are in addition to the usual gear coupling requirements. If a Jordan is involved the amount of clamping movement is necessary to know.
7. Spindle Couplings
Spindle couplings are special floating shaft gear couplings that are used in rolling mills. They are designed for high torque, shock loads, and high angular misalignment. They have replaceable wear parts and customized accessories. Spindle couplings also have some slide capability to adjust to the installation or operational requirements of rolling mills. The spindle coupling uses the continuous sleeve principal to reduce the overall outer diameter.
8. Insulating Couplings
Gear couplings can also be equipped to block galvanic (electrical) currents, which can cause pitting and corrosion at the close running fits of gear teeth and other mechanical components. One half of the coupling is electrically insulated from the other half by adding insulating plates and bushings. It is not necessarily a high voltage insulator as found in wiring systems.
Modification for Special Needs
Both continuous sleeve and flanged sleeve gear couplings can operate in the vertical position with the addition of a vertical kit, which is a limited end float plate or plate-and-button that supports the loose weights above the coupling. The button is rounded to allow the load to transfer under misalignment. Therefore, load is transferred to the lower shaft and ultimately supported by a thrust bearing in that equipment. Since a gear coupling is normally a shrink or interference fit, the upper hub is fixed to the shaft as is the lower one. In a vertical-floating shaft coupling, where both outer hubs are rigid and inner hubs are flexible, the entire center rotor is loose weight that needs to be supported by a plate in the upper coupling and a plate and button in the lower coupling.
A special vertical coupling is the rigid adjustable pump coupling. This coupling is designed for use with vertical circulating pumps that need clearance adjustments in the impeller. As indicated in the coupling name, it is a rigid coupling with no teeth in either half, with no provision for misalignment. The entire rotor weight is hung from the motor or driver bearings. Special designs of hanging load gear couplings can provide misalignment capability.
Other Gear Coupling Special Configurations
Gear couplings can be configured to do special jobs. Possibilities include the shear pin, cutout shifter, and brake coupling. Shear pin couplings disconnect when subjected to predetermined torque overloads thus protecting other equipment. Torque overloads could come from stalls or cyclic overloads.
Cutout couplings allow the driving/driven halves to be disengaged without disassembling the coupling. They use a special sleeve in which the teeth are interrupted at one end by a flat-bottomed annular groove. When the sleeve is shifted axially to align the groove with the teeth of one hub that hub spins freely, disengaged from the torque transmission path. A cutout pin (set screw) holds the sleeve in engaged or disengaged positions. They can be used on a dual drive machine to isolate the unused driver, or for a turning gear that rotates heavy equipment when it is off line, and helps prevent a permanent set in the shaft. Automatic cutout is available for temporary disconnect "on the fly" to allow adjustment of relative position between driving/driven halves.
Brake Drums and Brake Discs
Gear couplings are easily modified for the attachment of a brake, which saves system space by eliminating a separate brake. In other situations putting the brake at the coupling prevents the high cyclic torque from reaching low torque shafts. Brake wheel couplings are often attached near the gearbox shaft since high gear inertia is in the box. The brake drum or disk is a piece of metal, machined to standard brake sizes and clamped between the coupling's bolted flanged sleeves, requiring longer bolts. The coupling manufacturer does not include the brake and actuator.
Whenever a brake is installed in the system, it flags the need to check the stopping torque requirements. Stopping torque, like starting torque, depends on the amount of time that is available to stop or start. See the section on torque for a torque formula.
Moderate and High Speed Applications
As noted earlier, gear couplings are capable of very high speeds and high torque together. The limits have always been the need for lubrication of the mating gear surfaces and the need for balance. While high speeds increase the wear rate and can be the cause of high stresses within the coupling, the bigger issue is balance. Couplings operating at high RPM or high rim speed will cause vibration problems if they are not in balance.
A full discussion of balance will be found in another section of this handbook so only a few issues that relate specifically to gear couplings will be referenced here. Balance concerns itself with how the weight of the rotating mass or inertia is positioned or displaced relative to the center of rotation. If that weight is perfectly distributed around the center of rotation, the coupling is in balance. Since nothing is perfect in couplings, there is always a potential unbalance.
Coupling balance is achieved through design, manufacturing and remedial balancing machines. Off center bores, out of round circumferences, non-parallel sides, or even loose fits lead to mass displacement. In castings some of the potential unbalance could come from voids or air space internal to the casting. When a coupling consists of an assembly, the component design and the assembly process can result in an unbalance condition.
If the hub OD is not perfectly concentric with the hub bore the center of mass and center of rotation will be different. This means the gear teeth must be carefully cut with a pitch diameter concentric with the bore. That is controlled by the arbor or mandrel used on the hobbing machine and the concentricity of the pilot bore. The hub face must be perpendicular to the bore or to the hub OD. If it is not it becomes a trapezoid. Trapezoidal hubs have poor weight distribution and therefore unbalance.
Sleeves must likewise be concentric with the hub bore at the pitch diameter, the OD and at the pilot fits if any exist. Flanged sleeves must have a concentric bolt circle as well as a proper hole size and location. Flange-to-flange alignment before bolting will have a big effect on the balance of the assembled coupling.
Once the equipment is designed and the tolerances are established, it is possible to calculate the mass displacement of each component. The mass displacement of each component is added algebraically by a method that is called the square root of the sum of the squares. The total mass displacement can then be called the potential unbalance of the coupling. That total unbalance of the coupling could then be compared to recognized standards to see if it is acceptable. Refer to AGMA standard 9000-C90 for more on this subject.
Component & Assembly Balancing
It is unlikely that calculation of the mass displacement would be sufficient to satisfy a high-speed specification. That leads to the next process. Each component or piece of the coupling could be subjected to a balancing procedure on a balancing machine. Single-plane or two-plane balancing is also a consideration. If the coupling's width-to-diameter ratio is 1:1, or greater in diameter, single-plane balancing is sufficient. If width (axial dimension) is greater, two-plane balancing is needed. (See chapter on balancing for more information.) Machine balancing results in adding or subtracting weight from the piece to counter the unbalanced weight and lessen the unbalance. The remaining unbalance of the part while on the balance machine is called the residual unbalance. The coupling can be assembled after component balancing and left at that potential unbalance. The total unbalance of the assembly at that point would depend on the distribution of the individual high points within the assembly. The worst case would be to end up with all heavy points in one quadrant.
For further reduction of unbalance, the coupling could be assembled and returned to the balance machine, again with corrective adding or subtracting of weight. The result would be called an assembly balanced coupling. With these, all individual pieces are match marked before the coupling is disassembled so they can be reassembled exactly the same way on the users equipment.
A gear coupling is not easily assembly balanced. First the coupling must be assembled with tight fits between the hub teeth and sleeve teeth so that loose parts will not fool the balancing machine. After the coupling is balanced, the teeth are relieved so the coupling can be installed in a system with possible misalignment.
The final balance, after the coupling is removed from the machine, will be affected by the concentricity run-out and bearing surfaces of the mandrels, arbors and mounting devices of the balance machine. The coupling, unlike machinery rotors, is not balanced on its own shaft. A half coupling might be balanced on the equipment rotor, but then the two half couplings from the two different rotors must be joined together. Why should one worry so much about balance? The balance is critical on high-speed applications to prevent destructive vibration. Different applications have different definitions of high or low speed, but generally for couplings, anything greater than 3000 RPM is high speed.
High Speed Gear Couplings
Most high-speed gear couplings are spacer types, that would acknowledge the need for maintenance on the connected equipment. Two important attributes of high-speed couplings are lightweight and low inertia. If the coupling is to be accelerated from zero to 10,000 or 15,000 rpm the torque required to reach those speeds quickly is substantial if inertia is allowed to be too high. High-speed machines are sensitive to overhung weights too. Everything is built for speed, which means small, light and precise.
We mentioned that high-speed couplings are precision made to tight tolerances. They are also made with ground bores, body fitted bolts and reamed holes in the flanges. Since the couplings are highly stressed the materials are magnetic particle inspected to make sure of the integrity of the piece. The material may be standard 4140 steel, but it often has papers to prove its strength and chemical composition. Hubs are attached to the shaft by hydraulic fits on a taper in the really high-speed units. That eliminates keys and keyways that could affect the balance. Other methods might include an integral flange on the rotor that bolts up to a marine style spacer coupling.
Sometimes the need for maintainability or rigidity forces the coupling to be a marine type of spacer coupling. Marine style refers to the tooth location not the application. In a marine type unit the gear teeth are on the spacer section not the hub section. This increases the overhung moment so a trade off is being made.
Materials for High Speed Units
While balance is most important to high speed gear couplings, it must also be noted that high speed has the potential for high wear of the teeth. For that reason extreme high-speed units utilize hardened teeth to extend the coupling life. However, this requires material that will be compatible with induction hardening, carbonization, or nitride hardening. The hardened tooth must retain its strength to carry the torque. Iron carbides and carbon or other nitrides provide the surface hardness. While AISI 1045 carbon steel is the most popular for gear couplings, AISI 4140 high alloy steel is used on high-speed units. Coupling materials and hardening will be discussed more thoroughly a bit later in this chapter.
Lubrication of High Speed Units
High-speed couplings are lubricated with oil rather than grease. The oil, which is circulated through filters and coolers, is sprayed into the sleeve on one side of the teeth and drained from the sleeve on the other side of the teeth. Circulating oil has the advantage of constant renewal, but even with the circulation it is necessary to prevent sludge build-up in the coupling. Sludge will prevent oil from reaching the necessary surfaces that need lubrication. Anti-sludge features in a coupling prevent the build up by putting drains and dams in the passages.
Grease-lubricated high-speed couplings are limited in their application possibilities. Even though grease labeled as "coupling type", will resist separation of soaps and oils, it is not enough for the true high-speed application. Another problem with grease is temperature build up. Oil that is circulating is also cooled. Grease that is static would heat up from the rubbing friction at the high speeds.
Mounting the Gear Coupling in a Shaft System
Metric Versus English Units
The metric and English systems of size and tolerance were developed without a desire to interchange with each other. Simple conversions are not satisfactory because different bore dimensions are used, along with different tolerances and different formulas defining tight and loose fits. Metric bores are defined in ISO standards while English bores are defined in AGMA and ANSI standards. Those standards are also summarized in coupling manufacturers' catalogs.
Hub to Shaft Interface
There are several methods to fasten the hub to the shaft. In all cases the objective is to have a joint that facilitates the transfer of torque from shaft to hub, is easy to install or remove, and does not make the alignment more difficult.
Clearance or Loose Fits
Loose fits are easiest to manufacture and to install. But, loose fits are not the first choice for gear couplings, except low torque applications or some nylon sleeve applications. The loose fit does not provide sufficient restraint for the forces found in gear couplings, so interference fits are used. Loose or clearance fit hubs use a keyway and a loose fit key to transmit torque, with a setscrew to hold the hub tight to the shaft and key to prevent wobble and fretting wear. The key and setscrew also help if some cyclic loading is present. Since that is the only means of transferring the torque, the length through bore for clearance fits is longer than that of other fits. The preferred length is 1.25 to 1.5 times the diameter of the bore. Keyways on clearance fit bores are a square cross section. Key sizes are matched to shaft sizes to ensure sufficient surface is available for the torque transfer. The key also has a loose fit within the keyway.
Interference (or Shrink) Fits
The interference or shrink fit is the hub mounting choice in the majority of gear couplings. It utilizes a hub bore diameter that is slightly smaller than the shaft diameter under all tolerance combinations. There are many combinations to the amount of interference, but a popular number is .0005 inches per inch of shaft diameter.
The interference fit installation is accomplished by heating the hub to the point where it expands enough to fit over the shaft. Heating can be done in ovens, oil baths or by induction. The induction method is popular as a hub removal method too. A temperature of 300° F to 350° F is sufficient to do the job. Excess heat may change the metallurgical properties of the hub, and excess shrink or interference may split the hub.
The interference fit hub has a straight bore with a keyway so the friction between shaft and hub and the key are not used to transmit torque. The key is the main means of torque transfer, and may be either a loose or interference fit. Again a square key is used, and most times a radius is included in the keyway and on the key to reduce stress concentrations.
Reduced keys, known as shallow, half height or rectangular keys can be used to allow greater shaft diameters within the hub limits. All are wider than they are tall. Metric keys are of the reduced or rectangular key variety. When using reduced keys, torque capability must be carefully assessed. On large couplings and shafts two half-height keys are sometimes used to strengthen torque transmission. Interference fit hubs use a 1 to 1 ratio between the hub contact length and the shaft diameter. That ratio may vary in applications prone to high cyclic loads or sudden peaks in the torque from transitory conditions.
Tapers and Mill Motor Bores
Two types of taper bores are also common on gear couplings. One type is the tapered and keyed mill motor bore. This hub fits a standard mill motor shaft that has a like taper. As the hub slides up the shaft it forms a tight fit with the shaft. A shaft end nut is used to hold it in place. This method achieves good torque transfer, with a tight fit. It is an easy assembly or disassembly feature. Tapered shafts of this type can be used with machinery other than mill motors.
Another type of taper bore is the shallow taper hydraulic type. In this type there is no key. The hub is expanded by hydraulic pressure and pushed up the shaft to a predetermined point. When the pressure is removed, the hub shrinks to the shaft. The shaft can have a nut or plate attached to the end for retention of the hub. Removal also is accomplished by hydraulic pressure. The hubs have oil grooves machined in the bore to facilitate the application of oil pressure. Taper bore shaft hub combinations require a very complete match between the hub and shaft. The contact area of the hub bore to a gage acting as a shaft is measured in the manufacturing of the hubs to make sure a proper fit will be obtained when the hub is mounted on the shaft. Standards have been established to use as a guide for percentage of contact.
Shrink fit and hydraulic fit hubs are the choice for the heavy torque applications. One of the weak points in the power transmission train is the interface between hub and shaft. It is also the place where cyclic loads and peak loads can cause slippage or fretting damage. The tightness of the fit contributes to a more secure connection for torque transmission.
Sleeve to Sleeve Interface
Gear couplings from size 1 to size 9 will match up half for half with other flange type gear couplings made to AGMA standard dimensions. However, while the dimensional standard ensures compatibility of the face to face match between sleeve flanges, it does not assure matching torque capability or bore. This should always be checked. When a labyrinth seal coupling is matched to an O-ring-sealed coupling, the bore capability and torque may both be different despite the fact that their flanges match and bolt together.
Bolts and Torque
Flange bolting is important to coupling reliability, as bolting can be a potential weak point. Most designs use a friction basis for transferring the load across the face to face match of the two coupling halves. Bolts are designed for tension loading, and primarily serve the purpose of clamping the two flanges together to enable face friction to transfer torque. In fact, the maximum outer diameter of the flange on flanged sleeve couplings is partially determined by the needs of space for bolts and surface for friction. Although friction is the main means of torque transfer, if the coupling is overloaded to the point of overcoming friction, it becomes a shear load on the bolts before becoming a coupling failure. Since the bolts are loaded by several types of forces one must be sure the bolt threads are not in the shear plane between the flanges.
Other specifications could allow body fit bolts to carry the load in shear, although from an engineering standpoint the concept of carrying load on bolts in shear is not favored. The body fit bolt has a tight fit to the bolt holes that keep the two halves concentric. To carry that method to the extreme one would drill and ream the boltholes at assembly and then match mark the two coupling halves.
Bolting will also affect and be affected by balance requirements. Balanced couplings may require weigh-balanced bolts. In addition, bolting can provide a means of piloting the two half couplings. To use the bolts as a pilot, the boltholes must be drilled to a close tolerance or line reamed at assembly.
Remember that the continuous sleeve coupling is not affected by any of the issues associated with bolting. The continuous sleeve coupling provides a bolt-free method of transferring torque through a continuous cylinder of metal with the additional advantage of a smaller outside diameter.
Although alignment is covered in another section of this handbook, the gear coupling has some special alignment considerations that should be noted here. As mentioned in the bolting section, it is necessary for the two halves of flanged type to have some sort of piloting for best alignment practice. That can be achieved by piloted bolts or better achieved by pilot rings or rabbet fits. The alignment needs depend on the connected machinery and the speed of operation. High-speed operation always needs close alignment. Always refer to the machinery specification's first, not the coupling specifications, when setting the alignment parameters. Since continuous sleeve couplings do not have bolts, alignment is done hub face to hub face.
Once in a while there is a call for an "indexing" coupling. That type of coupling aligns two shafts in a rotational circular position that is the same each time. To accomplish that, the hub keyway is cut to be in line with a tooth or a space. The second hub is cut the same way. If it is a continuous-sleeve coupling, the continuous sleeve might be marked to identify the same tooth or space on both ends of the sleeve. The procedure on flanged sleeve couplings is more complex. In addition to the keyway meeting the hub tooth or space, a bolthole on the flange also must be lined up with a tooth or space. The mating flange must be drilled the same way so that when it is assembled the unit will be aligned or indexed. Of course, to make this work, the shaft keyway must also be aligned with a significant part of the machinery. Indexing is done to a specified tolerance on the location of that alignment.
Additional indexing is accomplished with floating shaft couplings when the coupling on each end of the unit has a different number of teeth. The indexing can then have a number of set points equal to the product of the two numbers of teeth.
Selecting Gear Couplings
Gear coupling selection parameters include two very important items and many more secondary items. The most important items are the bore and torque capabilities, in that order. Bore refers to the nominal shaft size where the coupling will be used. Torque in this case refers to the normal operating torque that the coupling must transmit. The secondary items can include a whole host of things like speed, misalignment, weight, spacer length, inertia, etc.
Bore and Torque: First Pass Selection
The gear coupling size in most cases will be determined by the nominal shaft size. The nominal shaft size is a mixed number of units and fractions that represent a specific diameter of shafting. The actual shaft is the decimal equivalent of that number plus .000 minus .0005 or .001 inches. Nominal sizes are not just any number, but are chosen from a list of preferred numbers. Preferred numbers can also be metric in origin. This is part of our discussion is limited to inch numbers. That nominal number would also be the coupling bore with the actual size as a function of the class of fit.
Gear couplings typically use interference fits, so the coupling bore usually is smaller than the shaft size. The amount of interference varies by the designer's requirements, but a value of .0005 inches per inch of bore diameter is often used. For details on shaft size for interference or clearance fits refer to AGMA 9002-A86. That is an inch series document; if metric is of interest, refer to "Preferred Metric Limits and Fits" ANSI B4.2 1978 reaffirmed 1984.
If the nominal shaft size is equal to or less than the published coupling bore capability, the gear coupling is usually okay for the service. "If it fits it is okay" is the gear coupling motto. For example smooth running, 1800 RPM, machinery without high starting torque or stopping requirements can use bore size to select the coupling.
The second step in gear coupling selection is to check the torque requirement of the application vs. the torque rating of the coupling. Normal operating torque is used unless a peak or cyclic torque is known. If the application calls for peak torque or cyclic torque, more care must be taken. The application description is also important to see if further investigation is needed. At this point, the nominal torque requirements of the system times an application factor that could be used to select the coupling.
The normal or continuous operating torque of the system is that torque value that is required for design point operation on a continuous basis. Coupling ratings are sometimes listed as HP per 100 RPM, but torque and horsepower can be derived from one another if the speed in RPM is also known.
Service factors (sometimes called Application Factors) are applied to the normal torque to account for variations that are typical of specific applications. They are based on a combination of empirical data and experience, and provide a quick reference to guide selection of a coupling for torque, and perhaps life, without going into the details of the application. Service Factor tables usually are provided in coupling catalogs, and will be different for different types of couplings. Another source of Service Factors (application factors) is AGMA standard 922-A96. Factors of Safety and Service Factors should not be confused with each other or interchanged. The former is for design work and the latter is for applications work.
Stretching the Bore
This subject is included to highlight the fact that it is not recommended. Never exceed the bore associated with the coupling size and the key type. Square keys have a maximum bore, rectangular keys have another and metric has its own. Do not mix them. When extra shrink is requested, or an over bore is requested for low torque applications, engineering should review the application. The gear coupling is the most power intensive coupling as it is designed, but the shaft to hub connection can be the weak point of the coupling. Stretching the limits can result in machinery failure as well as coupling failure.
Bigger Than Size 7
There are several magic numbers when it comes to gear couplings. One is the size cut-off between big and small. That number is arbitrarily set at 7, but could be 9. The AGMA dimensional interchange goes to size 9 for gear couplings, but once the size rises to 7 and above, the number of applications become very limited. A size 7 gear coupling has a bore capability of nine or more inches (depends on key size too) and a torque of one million inch pounds. That torque corresponds to 16,000 horsepower at 1,000 RPM. Not many applications go that far and when they do the situation is special or low speed. Gennerally, big gear couplings are used on very low RPM and very high torque applications such as those found in the steel and aluminum rolling mills, crushers, rubber processors or mine concentrators.
For an idea of how big the gear couplings can be made, the catalogs will show gear couplings up to size 30. Loosely, the number equates to half the pitch diameter for flanged sleeve couplings. That means the coupling overall diameter will exceed sixty inches. Continuous sleeve coupling numbers are roughly equal to the maximum bore.
When the coupling size reaches the double-digit numbers, the torque rating is nebulous. Couplings are often re-rated based on improved materials, heat treating, and hardening. In reality the user and designer are trading wear life for torque rating. The torque rating can be used as a peak load or cyclic high and not always as the normal operating torque.
Not many modifications are made to these large coupling sizes. At this size, added functions are too expensive to build into the coupling and may be available as a separate device. Torque limiters fall in the latter category as they replace shear pins. The weight of the coupling and the other pieces of the rotating system also may preclude the desire for modifications. We should point out that large coupling bores are not always the ordinary bore and keyway because they may have special shapes and non-standard dimensions.
Catalog ratings are often accompanied by speed limits in RPM. It is possible to increase the RPM limit by balancing the coupling to minimize vibration. Balancing combined with special manufacturing tolerances can increase the speed even more. However, a perfectly balanced coupling will eventually have a speed limit set by stress, friction between the teeth, and lubricant breakdown.
All couplings have a misalignment limit. The standard gear coupling is capable of 1½° angular misalignment per mesh. Specially designed gear couplings can push that limit to 6° or more, per mesh. However, high misalignment limits can reduce the torque capability of the coupling. Misalignment accelerates tooth wear, because it causes the hub and sleeve to rub harder against each other. Sometimes high misalignment capability is sought for and limited to non-operational conditions, such as moving a shaft aside for maintenance.
Modifications used to achieve high misalignment capability in gear couplings include increases in backlash (tooth gap), additional crowning, 25° or more tooth pressure angles, hardened wear surfaces, modified grease seals, increased clearance between sleeve and hub (makes the teeth look taller), and a torque de-rating. High misalignment couplings may also have modifications to make coupling maintenance easier or less expensive such as replaceable wear surfaces.
Materials of Construction
Gear couplings are typically made of two common steels, AISI 1045 carbon steel, and AISI 4140-alloy steel. Alloy steel means elements other than carbon have been added to give additional properties to the steel.
Standard gear couplings use AISI 1045 steel. It can be bar stock or forging depending on the size and the component. Couplings needing higher strength or hardness for greater wear resistance are made from AISI 4140 which also can be bar stock or forging.
Gear couplings can be specified in 303 SS, but that is expensive and usually done only when required for the food processing or the pulp and paper industry.
Steel can be treated in many ways to improve hardness and strength. Hardness is the key to improving wear resistance for longer life under increased friction from high speed or misalignment, because gear couplings typically wear out under load rather than break. Strength provides resistance to the impact and cyclic loads.
The terms heat treatment, hardening, annealing, quenching and tempering are used in conjunction with the materials. Each of these terms represents a process that conditions the steel. Heat-treating is the general description that includes variations of all the others. Heat-treating does not have to mean hardening of the steel although it is usually taken in that context. Hardening of steel can mean in-depth hardened or surface hardened, which is also called casehardening. Hardness is measured in Brinnell units or Rockwell units, abbreviated as Bhn or Rc. The Rockwell Rc method of measurement is more popular on hardened surfaces of gear couplings while Bhn is used for overall hardness of a batch of steel.
For AISI 1045 steel, expected properties of strength for gear couplings would require a range of 190-260 Bhn. For AISI 4140 the range would extend up to 300 Bhn in the higher strength versions of the steel. The basic process in simple terms is that the steel is heated to a critical temperature held for a period of time and then rapidly cooled. After the rapid cooling the steel has a very hard structure that may need further tempering or annealing to trade hardness for strength. Rapid cooling is called quenching. Tempering or annealing is heating to a temperature and then cooling at a predetermined rate that is slower than a quench. The intent of these processes is to obtain a strong hard material that is ductile and tough.
For wear resistance we want to increase the surface hardness to 50 Rc or better. That requires an additional process known as hardening, case hardening, or nitriding. The process is to load the surface with iron carbides by exposure to carbon and heat or carbon nitrides and other nitrides by exposure to nitrogen and heat. The heat is provided by a heat treating furnace and the other elements are provided by the atmosphere in the case of nitriding or by packing the piece in carbon in the case of carburizing. The base steel has to be suitable for the process. In the case of nitriding the end product retains the original dimensions, but in the case of carburizing the end product grows and needs to be ground if the original dimensions are to be held. There are many methods, beyond these mentioned, which can harden steel surfaces. It is a complex subject. The process of hardening the surface of gear coupling teeth can extend the useful life of gear couplings.
Gear Coupling Applications
Reduced Moment, Three Bearing and Four Bearing Systems
The weight of the coupling and any reactionary forces all act at the center of the flex plane and cause a bending moment on the equipment shaft. When the coupling is placed close to a support bearing, the close support reduces that bending moment arm and the coupling can be called a "reduced moment" coupling. Reduced moments mean smaller loads and less wear on the equipment bearings. Placing the flex point close to a bearing also helps keep the system stable. Increasing the distance between flex point and bearing invites vibration, or wobble. For the most part, a three-bearing system has one bearing in the driven equipment and two bearings in the driver. The one-bearing side of the equipment is given a rigid half coupling without a flex plane. The two bearing side of the equipment, which is more stable, is given a flexible half coupling. With only one flex plane, this type of system can only have angular misalignment. Three bearing systems are commonly found in motor generator sets, and a long-shaft situation such as bridge crane traction drives.
The more common system is the four-bearing system with two bearings each in the driving and driven equipment. The system is more expensive and usually needs two flex planes because two bearings on each shaft make shaft locations rigid, usually in parallel misalignment.
Standard Couplings vs. Spacers
The simplest application for a coupling is a pump, compressor or centrifuge or the input side of a gearbox. These usually involve an electric motor drive mounted on the same base plate as the driven equipment. The coupling connects the two shafts and the most complicated issue is usually the BSE dimension. As the gear coupling has some range in BSE, the equipment designer can use a common size base plate for many different models of his equipment. The torque requirement of this type of rotating equipment is usually a smooth curve from zero to full speed and does not have any cyclic content. The coupling can be selected by torque and bore with a minimum service factor.
When the designer wants to make his equipment easier and cheaper to maintain, a spacer is installed between the two flex halves of the coupling. When the designer needs to span a long gap between driving and driven equipment (as when reaching up to a big-diameter roll, removing a large piece of equipment from an on-line position, or extending through a wall or bulkhead) a floating shaft is needed. This arrangement is often used with pinion stands, where the output is a double shaft that drives a meshing pair of rolls or mixers that are part of a large machine, such as a rolling mill.
Separating the Driver & Driven
Rotating equipment such as fans, pumps and compressors can have two separate drivers on the same piece of equipment. The drivers might be an electric motor for start-up and a steam turbine for running. That occurs on co-generation applications where steam is available and the operator wants to conserve electricity or use the electricity for other purposes.
Sometimes the equipment has an electric motor for normal purposes and some other device like an internal combustion engine for emergency operation. Other times the equipment sits idle but the driver runs. While these sound like applications for clutches, they also can be places where cut-out gear couplings might be the wiser choice. The gear coupling in many cases is less expensive and takes less space in the system than a clutch.
Save the Equipment from Torque
Rotating equipment shafts are often oversized because they are designed to limit deflection, which can lead to oversize couplings. Motors are sized as the next larger standard unit compared to the application requirements. Those issues plus a service factor can result in a drive system that has torque capability well in excess of the driven equipment needs. In such systems, torque spikes or overloads are easily passed to components that are not designed to withstand them and may be severely damaged. To prevent that, a torque limit device is installed in the drive train. The gear coupling, which probably is needed in the system for other reasons anyway, can provide the same protection at much lower cost than many devices sold as torque limiters, with the simple addition of a shear pin.
Grid Spring Couplings
The Grid coupling was designed in 1919 and has since found favor in applications involving pumps, gear boxes, cranes, mixers/agitators, bulk
material handling systems, mining equipment, compressors, and
fans/blowers. Grid couplings are most popular in the Steel, Pulp &
Paper and Mining Industries.
This design is characterized by two flanged hubs with slots cut axially
into the perimeter of their flanges, forming a ring of narrow, closely set
teeth around each flange. When the hubs are brought together, their
slots match, and a single serpentine grid of spring steel is wrapped
around both flanges so that the spring loops back and forth between the
two hubs, nesting in the slots. Lubrication is required, so a collar-type
cover with seals and gaskets is used to hold the lubricating grease in
place. It also keeps the grid spring from migrating out of the hub slots
when misaligned. Because the covers can be removed, worn or broken
grids can be replaced easily without disturbing the positions of either
driver or driven equipment. Two cover designs will be described later.
Torque is transmitted from one hub to another through the grid spring.
The flexible nature of the grid absorbs impact energy by spreading it out
over time, thus reducing the magnitude of peak loads. This is possible
because of the progressive contact that occurs between the flexible grid
and the curved profile (axial crowning) of the sides of the hub teeth. As
the load increases, more of the tooth comes into contact with the grid,
allowing the coupling to handle shock loads that occur within the
The Grid coupling teeth do not mesh between the hubs, therefore, if the
grid spring were to fail the coupling would no longer transmit torque.
The most common type of driver is an electric motor. However, the ability of the grid coupling to damp vibrations allows it to be used with reciprocating engines (4 or more cylinders), as long as the proper service factors are applied in sizing the coupling.
Standard grid couplings can be used in vertical-axis drives without any
modifications. Many modifications or adaptations are possible for special
applications. Examples of the variations of grid couplings that are
possible include floating shaft designs, disc brake couplings, spacer
couplings, high speed applications, engine flywheel adapters and
The grid coupling typically competes with either elastomeric or gear
couplings. Due to its high torque-to-outside diameter ratio the grid coupling
may be used in applications where elastomeric couplings are too
large to fit into the space constraints. It also becomes a less expensive
alternative where torque levels start to require comparably rated elastomeric
couplings to be much larger. The downside is that the grid coupling
tends to be higher maintenance than an elastomeric type coupling
due to its lubrication requirement, lesser misalignment capacity, and
The grid coupling excels in applications where an all-metal coupling is
desired but moderate vibration damping is required. The resilience of
the grid gives this design a damping capability that is typically not available
with an all-metal coupling.
Most designs have backlash or free play between the fit of the grid
teeth and spring and are not suited for motion control applications.
Temperature capacity is usually no greater than 220°F (120°C), due to
the rubber seals, which keeps them out of some applications.
To the users advantage, grid coupling manufacturers have kept their
couplings directly interchangeable. For the most part, components of
one can be used with another. The only real difference between manufacturers
is the shape of the seal and gaskets. Each manufacturer uses
a different shape of seal and gasket that fits with their cover assembly.
Therefore, it is required that you use the corresponding seal and gasket
of the cover manufacturer. If a different manufacturers seal and gasket
are used, there is the possibility that a seal will not form and there could
be loss of lubrication.
Being an all-metal coupling the only purpose of the cover is to hold the
necessary grease in place to provide proper lubrication. Some manufacturers
shot-blast the tooth area of the hubs to remove any burrs and
sharp edges that may cut into the grid spring. This allows for a
smoother transmission from the grid teeth to the grid spring.
Torque Transmission and Torsional Flex
Accommodation of Misalignment and Axial Displacement
Angular and parallel misalignment are restricted by the design of the covers and seals, and this restricted misalignment will introduce fairly significant reactionary loads on the shafts.
Parallel: The movement of the grid in the hub grooves accommodates parallel misalignment and still allows full functioning of the grid-groove action in damping out shock and vibration. Maximum parallel misalignment ranges from 0.012" to 0.022" depending on the size of the coupling.
Angular:Under angular misalignment, the grid-groove design permits a rocking and sliding action of the grid and hubs without any loss of power through the resilient grid. Maximum angular misalignment is ¼°.
Axial:End float is permitted for both driving and driven members because the grid slides freely in the grooves. Maximum allowable end float ranges from 0.198" to 0.571" depending on the size of the coupling.
Grid Coupling Types
There are two designs for the grid coupling: tapered grid and straight grid. The straight grid design is the original style developed in 1919. In the 1950's, the design was enhanced with a tapered grid, which supports higher torque ratings than comparably sized straight grids. This is due to the sliding action of the grid spring on the hub grooves versus the twisting action of the straight grid. Additionally, the hub teeth of the tapered grid style are stronger due to the base radii of the teeth.
The grid coupling is available with two different cover styles: horizontal and vertical. The horizontal design splits the cover axially so it can be installed or removed and replaced externally, in a wrap-around fashion.The two halves are joined by bolts inserted in tangential orientation. The vertical design splits the covers radially into two flanged circular halves, with one half pre-fitted over each hub from the shaft end and
joined by bolts oriented axially around the perimeter of each half. Both
cover styles allow the coupling to be opened for service or grid replacement
without disturbing the installation of driver or driven equipment.
The horizontal cover is manufactured from die-cast aluminum. Do to it
having a smaller outside diameter than the vertical cover for a comparably
sized coupling, it is ideal for limited space applications. Also, it is
well suited for reversing applications due to a lug that is molded into the
inside of the cover. The lug fits in between the grid spring and does not
allow the cover to spin in the opposite direction of the hubs. If the cover
were to spin it would break the seal between the seal and the hub and
the lubrication could leak out. Horizontal covers are always required for
The vertical cover is manufactured from stamped steel. It can be used
in higher speed applications than the horizontal cover because of the
cover shape. Vertical covers cannot be completely removed without demounting
the hubs, but for simple grid maintenance they are typically
just moved back over the hubs. Adequate clearance must be available
to do so.
Vertical covers do not have the lug like the horizontal version, so reversing
applications can cause the covers to spin in the opposite direction of
The horizontal cover is typically the more popular cover. It is believed that this has to do with installation and maintenance. When using the vertical covers, the covers have to be put on the shafts before the gaskets and hubs are installed. It is easy to overlook this installation sequence and force installed gaskets and hubs to be taken off again so the cover halves can be put in place. The horizontal cover avoids this mistake by allowing installation after the hubs and gaskets are in place.
The pump industry (primarily ANSI chemical pumps) has long desired spacer couplings for ease of maintenance. Spacer design couplings
allow for a standardized gap between the ends of the driver and driven
shafts. The spacer allows the coupling to be opened up with a gap
wide enough to let the pump casing be removed from the "wet end" of
the pump for servicing of components such as the impeller, seals,
bearings, etc. without disturbing motor or pump mountings.
The grid spacer coupling meets standard ANSI spacing requirements for
pump disassembly. One of the benefits of the grid spacer coupling is
that various components can be mixed/matched in combinations to
achieve dozens of other shaft separations beyond the ANSI
standards of 3 ½", 5" and 7".
The grid spacer coupling is achieved by using a shaft hub that is bolted
to a spacer hub. There are usually three or more lengths of spacer
hubs available in each size coupling. The shaft hub has the finished
bore and keyway for the driver/driven shaft. The grid spring wraps around the grid teeth of the spacer hub. The horizontal cover fits over the spacer hubs. Each half of the full spacer coupling uses a spacer
hub/shaft hub combination.
Half-spacer couplings are made possible by using a standard hub on
one side and the spacer hub/shaft hub combination on the other side.
The spacer hub is bolted to the shaft hub with four to twelve hex head
cap screws depending on the size of the coupling. By removing these
screws, the center section of the coupling can be dropped out.
The grid springs are made of a high tensile alloy steel that is formed to shape, then hardened and tempered under controlled conditions. The grids are then shot-peened, compressing the surface molecules and leaving a residually stressed surface. This process creates a stronger surface in compression.
Examples of Failures
There are three main causes of grid spring failure. The first, mode of failure, has to do with misalignment. When the grid coupling is misaligned
beyond the specified catalog limits, the grid seal will fail, causing
the lubrication to leak out. This loss of lubrication will cause the grid
spring to fail at the curves of the grid. This is not a catastrophic failure
and the grid coupling will continue to transmit torque. The user may not
even realize that the grid spring has failed until the next time maintenance
is done on the coupling.
The second mode of failure mimics the first mode, but occurs even in
correctly installed and aligned couplings if not given proper and sufficient
lubrication. When this situation occurs, the grid spring will fail at
the curves of the grid. This is not a catastrophic failure and the grid
coupling will continue to transmit torque. The user may not even realize
that the grid spring has failed until they do maintenance on the coupling.
The third mode of failure has to do with an over-torque situation. If the
driving equipment transmits more torque than the coupling can handle,
the grid spring will fail at the center of the grid. When this occurs, the
grid coupling will no longer transmit torque from the driving to driven
equipment. The grid spring would need to be replaced immediately in
order to continue operation.
Adequate lubrication is essential to prolong the life of the grid coupling and to obtain trouble free service. Grid coupling manufacturers specify
that the couplings be re-lubed annually when using a common industrial
lubricant. Special lubrication can be used to extend lube intervals. A
coupling that is exposed to extreme temperatures, excessive moisture,
frequent reversals or grease leakage may require more frequent
It is recommended that the coupling covers be removed to check lubrication
condition, alignment and the general condition of the grid and
teeth every year. Couplings used in ambient temperatures greater than
158ºF, at high speeds and/or frequent reversing applications may
require more frequent inspection, re-lubing and possible grid
Some manufacturers offer further adaptations of the grid coupling for special applications. These include; controlled torque, disc brake, brakewheel, engine flywheel, piloted, and high-speed designs.
The chain coupling is used extensively on unsophisticated applications such as agricultural equipment and machinery because it is an all metal,
rugged, lightweight and economical method for connecting two
The coupling consists of two sprockets (hubs with chain-matching teeth
on the periphery) connected with a double roller chain. It is easy to
install, easy to maintain and easy to rough align. The chain coupling
allows quick shaft disconnection by removal of the chain. A cover is
used to contain grease and keep out dirt. The coupling can be supplied
with taper bore bushings for easier installation, and can be upgraded
with hardened sprocket teeth and a precision chain to improve wear life.
The chain is the replacement element although the sprocket may also
Misalignment is accommodated by the loose fit of the chain to the
sprockets. Therefore, the coupling is a wearable unit. Couplings of this
type are suitable for a maximum of 700 HP at 1800 RPM, or 145 HP at
3600 RPM. The maximum bore in the largest size is 4 inches and maximum
angular misalignment is ½°. Typically, their applications present smaller values.
General Coupling Characteristics
An understanding of the general characteristics of couplings will help in selecting the correct coupling for an application and will help evaluate competitive coupling proposals.
Finite Life and Infinite Life Couplings
All flexible couplings fall into one of two categories, "finite life" or "infinite life.”
Finite-life couplings are those that wear in normal operation, because of using sliding or rubbing parts to transmit torque and compensate for misalignment. This group includes jaw, gear, grid, sleeve (shear), Nylon sleeve gear, chain, offset and pin & bushing types. All usually have lower purchase costs than infinite-life couplings. They won't last as long, but their life span may be sufficient for the life expectancy of the application. Periodic maintenance is required.
Infinite-life is something of a misnomer, as these couplings do not necessarily last forever. It simply means the couplings do not wear in normal operation and by design the acceptable loads do not exceed the fatigue life of the parts. This is because they transmit torque and compensate for misalignment through distortion within their flexing elements rather than by the sliding or rubbing movement of loosely fitted parts. This group includes tire, disc, diaphragm, some donut types, wrapped-spring, flex-link, and most motion-control types. (Sleeve types are excluded here because their torque and misalignment capabilities are served by the flexing of their elastomeric element, the interface between the element and its hubs is a loose gear-like fit that wears.)
Distortions of the flexible element results in fatigue stresses rather than wear. Infinite life in couplings remains infinite only as long as the operating load stresses, considering misalignment is kept within the fatigue capabilities of the coupling's material. Elastomeric couplings do not have the same fatigue capabilities as metal couplings and they also experience reduced load capability as time passes. For that reason, the shelf life of the elastomer must be factored into the couplings design rating
An overload that will fail an infinite-life coupling might only reduce the life of a comparably rated finite-life coupling. Accordingly infinite-life designs are most often used on maintenance-free systems where maximum torque requirements (including transient, cyclic and start-up torque) are known.
Single Flex, 3 Bearing Systems and Two Flex Plane 4 Bearing Systems
We have previously mentioned that some couplings, the metallic flex type in particular can be built as single flex element or two flex element couplings. Single flex element couplings were noted to be limited to angular misalignment and possibly axial displacement whereas the two flex element units were needed to achieve the additional parallel (radial) misalignment capabilities. Couplings of the single flex element type could be expected to have a lower cost. Elastomeric couplings may provide parallel misalignment in a single element through distortion of the element.
There are applications that require two elements and therefore two flex planes, and other applications that either allow or require the single element.
The coupling installed in a four bearing system will be of the two flex plane type or of a type that allows radial misalignment. The four bearing system consists of two pieces of rotating equipment, each having a set of two bearings. Each set of bearings will hold the associated shaft in a straight line when the equipment is installed on its foundation. Alignment of the two pieces of equipment will make the shafts close to coaxial and coplanar. However, since some misalignment will occur, a flexible coupling is needed. Unless one piece of equipment can swivel about a point, parallel misalignment will eventually show up in the system and a coupling with that capability will be required.
The coupling installed in a three bearing system will be of the single flex type. The single bearing is a self aligning type which provides the swivel possibilities. (It does not have to be a three dimensional swivel.) The system only needs an angular misalignment capability as associated with a single flex coupling. There are two types of single bearing systems. Note that both types have a radial load that is carried across the coupling. Most elastomeric couplings will not be able to carry that radial load and should not be used in the system unless checked for radial capabilities. Two flex plane couplings will be unstable in these systems and cause vibrations or wobble.
The first type of a single bearing system is one that places the load between the coupling and the outboard single bearing. This is typical of a three bearing generator coupled to a driver. The load can be heavy. Usually the flex half of the coupling is mounted near to the driver bearing to reduce the overhung moment.
The second type of single bearing system is the overhung pulley application. The pulley has one side open to allow for easy changing of the belts. The bending moment caused by the pulley load has to be passed through the coupling. The bearing between the coupling and the pulley is a pivot point and a load carrying position.
The floating shaft or floating tube coupling is a special case of using two single flex couplings in a four bearing system. The connection between the two single flex couplings is a long unsupported shaft or tube. The length of the shaft or tube is limited by critical speed, the diameter is a function of the torque. Tubes are used to lighten the weight and improve the critical speed. The flex halves can be on the center shaft, a marine type, or next to the equipment bearing, a reduced moment type. Mixed coupling designs with one side being a reduced moment and the other a marine style is acceptable. Vertical floating shafts are available. Floating shafts that use full flex couplings on both ends of the floating shaft are unstable when operating and should be avoided. Elastomeric floating shaft couplings are possible but must be reviewed and approved by the coupling manufacturer. Floating shafts are used when the connected equipment has offset shafts and space is available for the long shaft. They are found in paper and steel mills.
Some systems have a mix of floating shafts and semi-floating shafts. Usually single flex couplings are needed to provide stability in the rotating system. The flex half should be mounted nearest to a bearing for best results.
Torque - Limited and Bore - Limited Couplings
Some coupling designs are limited by the torque capability of the flexing element. They are called "torque-limited" couplings. Other couplings are limited by the hub bore size because the flex element is capable of transmitting all the torque that shaft size will normally deliver. These are termed "bore-limited".
The elastomeric coupling is considered a torque limited coupling device because the flexing element uns out of torque carrying capacity before the connecting shafts reach their full torque potential. This is becauseelastomers have much lower tensile and compressive strength than the metals otherwise used in flexing elements. Consequently, elastomeric couplings must become larger in diameter to achieve higher torque-carryingcapabilities. The hub naturally follows the elastomer in becoming large, giving the hub a bore capability that is unnecessarily large in relation to the torque capability of the coupling. Enterprising designers use that extra bore capability to fit tapered bushings and other easy-assembly devices into the hubs. (For more discussion on these devices see the chapter on mounting the coupling hubs to the shafts.) All this means that elastomeric couplings must always be checked for both torque and bore capability.
Metallic element couplings tend to keep a close relationship between hub, bore and torque capability. One notable exception is the gear coupling which is truly bore-limited because it can transmit more torque than its maximum shaft size will normally deliver.
Some composite materials offer strength capabilities somewhere between elastomers and metals. These materials sometimes offer weight-to-strength advantages that can be important.
Coupling selection always needs consideration of torque, speed, misalignment, connecting shaft sizes, and appropriate service factors. However, in old installations needing coupling replacement, the real torque values might be unknown or uncertain. In such situations the gear coupling could be selected purely on shaft diameter and speed, with limited risk.
Selecting elastomeric couplings purely on shaft diameter and speed is very risky. In some cases, however, that risk can be an advantage. When overloaded, the elastomeric coupling will fail before the rotating equipment shaft fails, provided the overall design is correct, thus sacrificing the less expensive coupling to protect the more expensive rotating equipment. When using this strategy, the overall design considerations should include the wear life of the coupling and the damping energy to be absorbed by the coupling.
The fact that the coupling can be the weakest element does not necessarily mean that the coupling will provide a fusible link. To have the coupling serve that function, it is necessary to pick the type of coupling with that feature. These types are termed non-failsafe. In failsafe design, coupling failure does not automatically disconnect the two rotating shafts, but will require the coupling to be maintained as soon as possible. If the coupling is a wearable device, as are most elastomeric couplings, both load and misalignment are factors in total life. The end of usable life of the coupling might not be the result of equipment problems when wear is a consideration.
1. Determining Torque Requirements
Coupling torque requirements can be defined many ways, and specifiers need to decide which definition to use. We will first review the various definitions, then discuss how they are used in coupling selection.
Normal Operating Torque
The steady state torque required by the system when operating at normal design conditions. This is usually the level at which the equipment designer certifies the equipment performance.
Starting TorqueThe torque needed when the system starts its operation. This torque can be greater or less than the normal operating torque.
Peak TorqueThe maximum torque required by the system. This torque is normally a one time event or limited to a specified number of occasions. In torsional vibration coupling systems it is the maximum vibratory response torque that could pass through the coupling.
Cyclic TorqueIt is any torque requirement of the system that varies with time. It can be of a smooth, periodic variation like a sine wave or could be an erratic variation. It does not go through zero to a negative value, but can be equal to zero. In torsional vibrating systems it is the vibratory torque that occurs at the operating speed.
Reversing TorqueThis is a cyclic torque that passes through zero and becomes negative or "reverses" to the opposite direction.
Transient TorqueA transient torque is of short duration, not necessarily expected, not happening on a regular basis but occurring when a system is upset. It may or may not be equal to or greater than peak torque.
Normal Braking TorqueIt is the torque used to decelerate or reduce the speed of the equipment when the brakes are applied in a normal manner. The torque is time dependent, and moves through the system.
Emergency Braking TorqueIn this case the brakes are applied to stop the equipment in a very short time. The torque will exceed the normal braking torque by the inverse ratio of the time required to stop in each case.
Stall or Lockup TorqueThis is the torque that passes through the system when the system stalls or otherwise come to a stop because of some activity within the driven system.
Shutdown TorqueThe torque required to bring the equipment from operating conditions to a shut down condition. This can be the normal braking torque or could be a result of friction or load in a system that is coasting to a stop.
Torque to Accelerate or to DecelerateThe torque required to increase or to decrease the equipment operating speed. In the case of acceleration, the available torque for acceleration is the difference between the driver capability and the system requirements at the current speed. Decelerating torque comes from braking devices or from frictional drag or other energy drains within the system that cannot be overcome by the driver. A formula for calculating this torque is found at the end of this section.
Driver Horsepower (Torque)
Nameplate Rated Horsepower (Torque)
The torque value is derived from the driver capability shown on the nameplate as a horsepower and a speed. It is based on specific inputs to the driver such as voltage, and amps or kVA if it is an electric motor. A formula to convert horsepower and speed to torque is found at the end of this section.
Service Factor Rated Horsepower (Torque)Some drivers have additional capabilities beyond the name plate rating. The nameplate capabilities are multiplied by the service factor. The service factor is also shown on the nameplate.
Start-Up TorqueThe driver torque capability at start-up available to accelerate the driven equipment to operating speed. Some drivers have a fixed percentage of the rated torque available at start-up. It can be greater than 100%.
Peak TorqueThis is the maximum torque available from the driver, it may not be able to operate for extended time periods at this torque.
Stall TorqueThis is the system torque requirement that will cause the driver to come to a stop.
While all the torque values defined previously may exist within the system at some point in time, the torque requirements of the driven equipment are the primary consideration. The driver will not supply moretorque than the driven equipment will absorb or the driver can produce. Under some conditions the maximum torque within the system may exceed the driver capability, for example when brakes are energized.
A piece of driven equipment operating at its full speed capability requires a certain amount of torque. If the driven equipment is not operating at full speed the driver will supply additional torque until equilibrium is reached for torque and speed. The driver could still have additional capability, but it will not transmit it to the driven equipment. Other than at start up, the speed variation is small and subject to driver speed limits.
Drivers have speed limits that are imposed by physics or by trip devices. The physical limits can include the effects of frequency on an electric motor, or the effects of fuel restriction on an internal combustion engine, or steam availability to a steam turbine. Trip devices can include governors and over speed switches. The speed-torque capabilities of the driver are fixed by the design of the driver and the inputs to the driver.
Coupling Selection Torque
Using the Driver TorqueThe coupling can be selected based on the driver capabilities, using nameplate values or start-up torque. The capability requirements can be increased by an application service factor before choosing the coupling. This method of coupling selection usually results in a coupling that is oversized for the application, even if the service factor is 1.0. This translates into high cost and other problems. The reason oversizing results is twofold. First the driven loads may include equipment service factors that have already increased the torque value. Second, the driver is usually oversized. Drivers such as electric motors, come in standard sizes. If a piece of driven equipment requires a horsepower that is in between two standard sizes, the larger is chosen. Even when the requirements are right on the nose, the designer will usually pick a larger size out of conservatism.
When the coupling is chosen by driver horsepower one must be sure there is no gear reducer between the driver and the driven load. Gearboxes are constant horsepower devices that increase the torque or decrease the torque depending on the gear ratio input to output. Other power transmission devices may do the same. In any case those types of devices must be accounted for in the coupling selection. Couplings selected using driver torque are normally mounted with one half on the driver shaft.
Using Driven Equipment Torque and a Service FactorThe coupling can be selected by using the normal operating torque of the driven equipment, adjusted for coupling location, multiplied by a service factor. Service factors are used to account for unknowns in the driven equipment system.
Service FactorsSometimes "Service Factors" are called "Application Factors" or "Experience Factors". They have been empirically developed for most applications, or are known by their designer based on experience with their systems. Coupling manufacturers publish Service Factors based on their experience with their couplings on various systems. The factors are listed in coupling catalogs. Manufacturers may publish different service factors by product line. The catalog service factors will include factors for the application and the type of driver. Elastomer couplings sometimes include an environmental temperature service factor, and if intended for dampening vibratory torque, will have a frequency service factor as well.
AGMA Standard 9922-A96 lists service factors for many different applications. Service Factors are not the same as design Factors of Safety. Service factors deal with the unpredictable nature of the application, not with unknowns in the design of the coupling.
Depending on the selection of service factor this method also could result in an oversize coupling. Oversize couplings cost more and can result in bearing overload, excess inertia, premature wear and more maintenance.
Using System Torque with Little or No Service FactorA coupling can be selected based on the exact requirements of the system. In this case the requirements must include all the torque values to be transmitted through the coupling. That can include starting requirements, braking requirements, peaks, transients and any others listed at the beginning of this section. Check the coupling manufacturers catalog as the coupling can have various torque capabilities. It may have one rating based on normal operation with another simultaneous rating for low cycles of peak torque. It may be acceptable to compare the peak system torque with a 1.15 Service Factor, to the yield strength of the coupling, and allow that as part of the acceptable selection. The coupling manufacturer should be consulted when the coupling selection is based on peak torque, emergency torque or a high transient that comes along only once. The coupling may already have sufficient reserve to satisfy those requirements on a limited number of occurrences.
If the coupling has only one published torque value, the coupling would have to be selected to meet the maximum torque expected in that part of the system. However, some types of couplings have torque ratings based on wear life, maximum misalignment combined with torque, or conservative considerations.
If the coupling is subject to cyclic torque or reversing torque, the selection should be based on those torque values. In the case of cyclic torque, use the high value. In the case of reversing torque, it will be necessary to check the coupling's fatigue life against the torque peaks and the acceleration/deceleration requirements associated with reversing operation.
The most economical selection will be based on the exact torque requirements of the system including peak, transients, breaking, or other expected torque values. Of course, this approach requires that all of the torque values be known with certainty. When the coupling has been sized to meet torque, it must also be checked for bore capability. Some bore-limited couplings might have the needed torque capacity but not enough bore capability to accept the shaft that will deliver it. Likewise, some torque-limited couplings might have sufficient bore capability to accept the shaft but not be able to carry all the torque that the shaft will deliver.
Using Torque Information Coupling CatalogsCoupling manufacturers have several methods of listing the coupling capabilities in their catalogs. These capabilities vary on each manufacturer's experience, design requirements, and testing capabilities. Torque capabilities found in the catalogs may have to be factored or reduced for misalignment, vibration frequency, temperature (elastomers), life (including elastomeric shelf life), or maximum torque. Such factors, if they are to be used, should be shown in the same catalog.
Most often the value shown in the catalog is the normal torque capability that the coupling can transmit over its design life. Some couplings have a listed maximum torque. Usually that maximum value is used when the application might involve short cycle fatigue on couplings that have infinite life. Couplings that wear over time, such as a gear coupling, may have maximum capabilities that are quite large as long as their application keeps wear low. Couplings that wear may also offer alternate materials for reduced wear and longer life or for higher torque.
Because some coupling torque capabilities are limited by wear of the flex element and others limited by on fatigue of the flex element, it is best to understand the type of coupling that is to be used in the system before selecting the size. In addition to the flex element, coupling torque capability is affected by the method of securing shaft to hub, and any other joint in the unit, bolted or otherwise. Usually it is the flex element that is the limiting factor for the catalog torque values.
Useful torque equations:Converting horsepower to torque
T = BHP x 63025 / RPMWhereT = the torque in inch-poundsBHP = the motor or other horsepowerRPM = the operating speed in revolutions per minute63025 = a constant used for inch-pounds, use 5252 for foot-pounds, and 7121 for Newton-meters
Converting kW to torqueWhereT = BHP x 84518 / RPMT = the torque in inch-poundskW = the motor or other kilowattsRPM = the operating speed in revolutions per minute84518 = a constant used for inch-pounds, use 7043 for foot-pounds, and 9550 for Newton-meters
Determining the acceleration or deceleration torque
T = (Wk^2 x N) / (307 x t)
T = the torque to accelerate or decelerate in foot-poundsWk2 = the inertia of the piece to be accelerated or decelerated in poundfeet squaredN = the absolute change of speed in RPMt = the time for the speed change in seconds307 = a constant that allows the speed to be in RPM, the time to be in seconds and the torque and inertia to be in pounds and feet. It is a common form of the equation.
2. Motion ControlWhile all couplings claim constant RPM transfer, not all can meet the rigorous demands of motion control, for two shafts to be synchronized, moving at exactly the same speed to exactly the same location. Such demands are found in applications such as shaft encoders, resolvers, all forms of servo devices, linear and ball screw actuators, robots, stepmotors, light duty pumps and metering devices, plotters, medical equipment, positioning tables, computers and radar.
Coupling characteristics most important for motion control include high torsional stiffness but low radial stiffness, low inertia, constant velocity, no wind-up, zero backlash in coupling components, shaft interface and corrosion resistance. Highly accurate machine tools, robotic systems and printing presses need extremely stiff couplings, as do encoders, which provide positioning feedback to the system.
Coupling torsional stiffness often is related more to the torque used by the system than to the capability of the coupling. Lightly loaded couplings can act like very stiff couplings, but this usually means they are oversized for the job. This oversizing effect is often seen when low backlash curved jaw couplings are used for motion control.
Another required attribute for the motion control coupling is to withstand peak torque loads that are high relative to the nominal torque and size of the coupling. High peak loads result from reversals. In these applications couplings should be sized for the reversing torque rather than the normal torque. The reversing torque may in fact be the normal operating torque. One way to reduce the start-up and reversing torque peaks is to reduce the inertia of the two halves of the system, including the couplings. For this reason motion control couplings are designed for low inertia.
Fortunately, most of the high-precision end of motion control applications call for small equipment with small shaft diameters and small couplings. As a result, nominal torque values are low in absolute terms even if not in relative terms.
Small shafts must be aligned properly too, because reactionary loading from misalignment will be detrimental to both shaft and bearing systems. Motion control couplings are designed to minimize reactionary loads even though they are stiff couplings. That might seem contradictory since stiffer couplings generate higher reactionary forces. However, bellows and beam couplings provide a good example of being radially resilient while torsionally stiff.
Finally the system could be subject to lateral critical (vibratory) speeds if the coupling is not stiff enough in that direction. Once again, just because the systems are usually small does not mean they can't experience these problems.
3. Torsional Vibrations
One class of coupling applications is unique in that a secondary load is transferred through the power transmission system and the equipment connected to the system. That secondary load is torsional vibration. Torsional vibrations are associated with internal combustion engines, reciprocating (piston) type compressors, vane passing frequencies of some centrifugal pumps, grinding mill drives, kiln drives, rolling mill drives, variable speed motors, and the start up of synchronous motors. Diesel engines represent the most significant unit volume of torsional coupling applications, and will be discussed in separate detail later.
Torsional vibrations cause equipment breakdowns such as wear or chatter on loose connections like spline pump shafts, or complete fatigue failure of the shaft or some other element. These harmonic torsional pulses are difficult to detect, because they do not bounce the equipment up and down, as would a lateral vibration. Nor can they be felt by touching the equipment. Usually, the result of the vibrations is known before the vibration is known. Often something else is blamed.
If the torsional vibratory frequency matches a system torsional natural frequency, the system reaches a harmonic or becomes resonant. That's because the natural frequency is an energy balance point at which additional forces will set off uncontrolled vibration. From a technical standpoint, it is the frequency at which the kinetic energy of spinning inertia blocks is equal to the potential energy of the torsional spring connecting the inertia blocks.
In such systems, inertia blocks can be impellers, pistons, mill rolls, motor rotors or any other device that is mounted on the shaft, which all rotate together as a single wheel. The torsional spring is a combination of the shaft and coupling's flexible element, plus other potentially flexing components such as a spacer or floating shaft.
When the wheel and spring rotate as parts of the same system, the inertia of the spinning wheel is balanced against the windup of the spring. Any additional forward pulsing force on the wheel will cause the spring to windup more and that will in turn react with a reverse force to the wheel. Between pulses, when that additional force is removed the spring unwinds, and the wheel surges forward. When the pulsing force returns, the spring winds up again, reapplying the reverse force to the wheel, etc. That pulsing force which is being applied at some time cycle or frequency at or related to the operating RPM, is the torsional vibration. If the timing is right, the winding and unwinding of the spring and the energy changes in the wheel resonate back and forth. The point where the timing is right is the system's natural frequency.
Determining the Natural FrequencyAll rotating systems have a torsional natural frequency. It is a function of the driven inertia, driver inertia and the torsional stiffness of the shaft, spacer and/or coupling connecting the two. There is a natural frequency for each combination of inertia and spring. Aside from the kinds of torsionally sensitive systems discussed in this section, most systems have torsional natural frequencies so high as to be inconsequential. By itself, the natural frequency is harmless and does not generate torsional vibration, but is simply a sensitive spot along the systems RPM curve. It is a "forcing frequency", i.e. it is not self-initiating or self-sustaining, rather it must be triggered by a vibratory force pulsing at that frequency.
Many systems can be reduced to a two-mass system. For a two-mass system, the frequency can be determined mathematically from the following equation.
CPM= (60/2π) SqRt (Ctdyn x (JA + JL) / (JA x JL))CPM is the frequency in "cycles per minute".JA is the polar inertia of the driverJL is the polar inertia of the drivenCtdyn is the dynamic torsional stiffness of the coupling.60/2π is a constant.
Reducing a system to a two-mass system is done by lumping inertias connected by torsionally stiff shaft elements. For example the lumped polar moment of inertia of the driver JA and the lumped polar moment of inertia of the driven equipment JL are determined by adding all the individual inertias that are connected by stiff shafts. When a gear reducer or increaser is involved, the downstream inertia must be factored by the square of the gear ratio (speed). It is an inverse function.
A coupling is between the driver and the driven. The coupling stiffness Ctdyn is obtained from the coupling manufacturer. It is called the dynamic torsional stiffness, which is higher than the static torsional stiffness.
Inertia is marked by the symbol "J", the units in the English system are inch-pounds second squared. It is related to WR2 by "g" the acceleration due to gravity. For a method of calculating the inertia value and the stiffness of connecting pieces refer to AGMA Standard 9004.
In a multi-mass system that includes more than two inertias connected by torsionally soft shafts, couplings, or sections, the natural frequency can be determined using the Holzer method. Refer to a textbook for an example of the Holzer analysis.
As long as the torsional natural frequency is more than 40% above or 30% below (.7 Nc to 1.4 Nc, where Nc is the critical numerical value) the system's operating frequency or idling frequencies (RPM) or associated torsional vibration frequencies (CPM) no resonance problems should occur. If it is in between those values there is a good chance the system vibratory response will cause damage to one or more components. If it is close to any of those frequencies, resonance is likely to occur.5000Campbell diagrams are graphic plots of operating speeds and pulse frequency. They are used to identify the potential trouble spots where operating or idle RPM is equal to a torsional pulse frequency in CPM, (cycles per minute).
If it is decided to operate the system normally at an RPM above the torsional critical speed, (natural frequency) then the driver must have enough torque available to accelerate the load quickly through the critical speed zone (RPM). Comparing the speed torque capabilities of the driver and the load will determine the system's ability to accelerate through the critical zone quickly enough.
Using the Coupling to Tune Critical FrequencyIn the torsionally sensitive system, couplings take on an important extra role beyond the transfer of driving torque and the handling of misalignment. They have the ability to move the natural frequency away from those levels that will be occupied by the torsional vibratory frequency at normal operating or idling speeds. This is called "tuning" the critical frequency. It works as long as the coupling is the controlling element for the critical frequency. That is not the case when long slender shafts are in the torque path. They also can be used to damp the energy of torsional vibration to reduce its potential for damage. The coupling torsional stiffness/softness is an attribute that is important in serving these functions.
Couplings with the highest levels of torsional stiffness are not used here. Those designs primarily serve systems that must transfer motion without windup or backlash, as previously discussed under motion control. Torsionally soft systems have a normal operating speed above the torsional critical speed while torsionally stiff systems have a normal operating speed well below the torsional critical speed. Because the coupling is usually the softest torsional element in either system, the system tends to be stiff when the coupling has a relatively high torsional stiffness and soft when the coupling has relatively high torsional softness.
Stiff couplings have elastomers of the Zytel® and Hytrel® type of plastic or their flex elements are metal. Soft couplings are rubber elastomers in compression or in shear.
Changing to a torsionally stiffer coupling raises the system's natural (critical) frequency, and reduces or eliminates the coupling's capacity to damp vibratory energy.
Using stiffer couplings to drive the critical frequency above the operating speed is as effective on simple systems like a single hydraulic pump driven by a diesel engine as it is on sophisticated high-speed couplings that are found on turbine driven rotating equipment.
When the coupling is used in the regimen of keeping the critical frequency high, it is usually just a matter of making sure the coupling is sufficiently torsionally stiff. That could be accomplished by using a stiff spacer piece with a metallic-element coupling, or by using a very stiff elastomeric element on a flywheel coupling. The coupling manufacturer can provide the necessary information on coupling and spacer piece stiffness.
An exception would occur when a system has a long slender shaft, which usually means the lowest critical frequency would be the result of that shaft. That type of system can become complex because the coupling is no longer the element that controls the stiffness.
The torsionally stiff elastomeric coupling and the torsionally stiff metallic element coupling offer no damping between the driver and the driven equipment. That means torsional vibrations are passed into the driven system. In such stiff systems, loose parts or parts with backlash will vibrate and rattle, and may have wear problems. Typically spline shafts on hydraulic pumps and gears with backlash suffer the wear.
Changing to a torsionally softer coupling lowers the system's natural (critical) frequency, but also may increase the coupling's capacity to damp vibratory energy, so that function and the heat it will generate through hysteresis needs to be considered in the selection. Elastomeric couplings expected to damp torsional energy must be designed to reject the resulting heat to a heat sink. Otherwise the heat will fail the elastomer by melting from the inside out.
Elastomeric torsional couplings can be either compression type or shear types. The more common compression types are of the donut or torus configuration. However, some use elastomer blocks or elastomer cylinders. The compression block types are most often found in the high torque applications. The shear types are shaped to equalize stresses from torque and misalignment.
Torsional softness and torque capabilities are opposite coupling characteristics. A soft coupling tends to have lower torque capabilities than similar sizes of stiff couplings. The softer, lower-torque couplings generally are used on applications that require 100 HP at 2,000 RPM or less. Torque capability increases with torsional stiffness of the flex elements.
The coupling designer must balance the various attributes to achieve the desired coupling for the specific application, or to devise a coupling with broad capabilities as a standard unit that serves many applications. Dual stage torsional couplings can also be obtained. They incorporate two different elements. One is soft for low or idle speed and a stiffer one for high or operational speed.
Some torsional coupling types utilize viscous friction damping This method is found in hydraulic torque converters, which mechanically isolate the driven system from the driver, and transmit torque between them through the motion of a viscous fluid. When a system uses a torque converter, it becomes two separate torsional systems. Torsional vibrations do not pass through the torque converter, except when a lock up device is engaged to mechanically connect the two halves. Hydraulic torque converters are not included in this handbook's discussion of flexible couplings.
Refer to the torsional coupling section and the metallic element section of this handbook for a more detailed description of the couplings used to damp torsional vibration and/or tune critical frequencies. Also refer to the bibliography for more publications on this subject.
Torsionally Sensitive SystemsTorsional vibration problems appear primarily in four types of applications briefly discussed here.
High Speed MachinesHigh-speed machines have torsional pulses or vibrations at high frequencies, therefore the natural (critical) frequencies must be kept even higher. A discussion about high-speed special purpose couplings andthe associated equipment torsional problems can be found in many of the coupling textbooks. They are a sophisticated coupling application, which is not covered in this handbook.
Variable Frequency DrivesThe VFD will produce a torsional pulse at low speeds that is larger than those generated at faster speeds. Keeping the operating speed above 10% of the maximum speed, i.e. no lower than 90% below maximum will alleviate the problem in that type of system.
Synchronous MotorsSynchronous motor start-up is a unique situation. At start-up the motor produces torque pulses at a frequency equal to two times the slip frequency. (Slip frequency is numerically equal to full synchronous speed minus operating speed.) The magnitude of the pulse is related to the torque developed by the motor. As the unit accelerates to full speed, the torque pulses drop in frequency reaching zero at full synchronous speed. The torsional vibrations or vibratory torque ceases at that point. A problem will occur if a torsional natural frequency is less than two times the AC power line frequency, as the start-up torque pulse frequency must then pass through the critical frequency. When going from startup to the running speed the driver must accelerate the load through the critical speed quickly. Acceleration through the critical frequency is a function of the torque available from the motor at starting.
High torque synchronous motors will also have high vibratory pulses that need the damping of torsionally soft couplings, but soft torsional couplings have a high vibratory response when passing through the critical speed, as well as difficulty carrying the high torque loads. The relationship between the two functions therefore must be a compromise. The coupling must be soft enough at startup to dampen some torsional vibration energy, but stiff enough to carry the high torque at running speed.
Reciprocating Internal Combustion Engine DrivesThere are three main types of engines in common use. They are gasoline engines, gas engines (natural or LPG or propane or other), and diesel engines. Gasoline and natural gas engines are spark-ignited low cylinder pressure types as compared to the compression-ignited diesel engine, which requires very high cylinder pressure.
The diesel is the most efficient of the three so it is very popular for continuous duty applications in those regions of the world that have high fuel prices. Gas engines (Natural, LPG or Propane) are most popular where these gases are readily available or where air pollution is a serious problem like the inner cities.
All internal combustion engines generate a torsional vibratory pulse. The magnitude of the pulse is a function of the cylinder pressure, turbo charging, engine's displacement, internal damping, the engine geometry, and whether it is a two or four stroke engine. The diesel drive rotating system accounts for the majority of the torsional vibration problems due to its high cylinder pressures and resulting high magnitude of torsional harmonic pulse compounded by its widespread popularity. The engine itself is designed to tolerate its internally generated forces from torsional vibration and may include some internal damping. The problems start when these harmonic vibrations pass to the driven equipment. Special attention must then be given to selecting couplings that can help reduce these problems
Diesel drives range from the simple low-horsepower single-unit hydraulic pump to a marine installation in which the diesel will drive the propeller and generators through a gear reducer. The preponderance of diesel drive systems can utilize a simple analysis to select the right coupling, however the marine system should be analyzed by an expert in that field.
While the magnitude of the torsional pulse is important, it is also necessary to know the frequency of the pulse. Like magnitude, pulse frequency is dependent on many factors. Those factors can include the number of cylinders, the configuration, such as "V" or inline, the stroke (two or four) and the firing order.
Also note that diesels typically have several torsional pulse frequencies, established at harmonic intervals. A 6-cylinder 4-stroke inline engine will have major harmonic orders of 3 and 6. Pulse frequencies are the RPM multiplied by the order. For example an engine running at 2100 RPM will have pulse frequencies of 6300 and 12600 CPM. If the natural frequency were also 6300 or 12600 CPM, the engine should not be operated at 2100 RPM. Note the frequency in CPM and speed in RPM is the same units in this case.
If any of the torsional frequencies are equal to a natural frequency the system will vibrate at resonance.
Coupling Selection for Torsional SystemsIn addition to the damping possibilities, the coupling is selected with three torque values in mind. The first would be the continuous running torque. The coupling should be capable of handling this torque under all environmental conditions of the applications. The second is continuous vibratory torque. The coupling should damp this torque without a meltdown from heat generation. The third is the maximum torque pulse or peak torque. The pulse occurs as a vibratory response torque at critical speed. The coupling rating is a fatigue life in that case. The manufacturer will publish the torque value at 100,000 non-reversing cycles.
Continuous Running TorqueThis is the design torque for the system. Usually it is the driver horsepower and operating speed. That value is normally in excess of the load requirements or is tied closely to the load requirements. Since the coupling also will be judged against peak or maximum transients in the system, a service factor is redundant except for high starting torque. Couplings that are oversized by using service factors can also be too stiff. Elastomer couplings may require derate factors on the coupling capabilities for temperature or frequency or speed. A derate factor is not a service factor. For more discussion on the various torque values found in an operating system see the chapter called "Applications ".
Maximum or Peak and Vibratory TorqueIt is important when analyzing the torsionally sensitive coupling application to know the value for the generated vibratory torque. That value becomes the forcing torque that puts the natural frequency into critical resonant vibration. In a diesel drive system the torque pulse is a function of cylinder pressure, number of cylinders, use of turbocharging, number of strokes for firing, etc.
Lloyds Register of Shipping publishes a pamphlet that is a good source of harmonic pulse factors used to determine the harmonic vibratory torque of diesel engines. The engine manufacturers could also provide the information. The manufacturers of other drivers or driven equipment should be able to provide the similar forcing torque values for their equipment.
Once the initial pulse torque is known, it is then possible to calculate the values for the vibratory response.
If the values are plotted on a graph of torque vs. speed, the peak will occur at the critical RPM. All values to the left of peak (below critical) are higher than the initial value. Once past peak the coupling will dampen the vibratory response starting at about 1.4 times critical RPM. Actually there are several damping possibilities in a system, but the damping type coupling is the best bet. Because of the damping, the torque pulse transmitted downstream to the system is reduced from the original value.
The torque pulse transmitted down the system can trigger other forced responses or vibrations. Thus an undamped pulse, like the type transmitted through the stiff coupling system, has the potential to damage downstream components. Loose connections, such as a hydraulic pump spline shaft, are susceptible to this damage unless they are protected. It is also likely that more than one pulse is generated in the operating range. This is very true of diesel engines. All the vibrations must be accounted for if they are in or near the operating range. The Campbell diagram shows the ones in the operating range.
At continuous operating speeds the vibratory torque capability of the coupling must be greater than the vibratory response torque pulse. The damping is energy absorbed by the coupling through hystersis, which results in heat generation. The coupling must dissipate the heat to survive. Its capability to dissipate heat is reflected in the published continuous vibratory torque rating.
The peak torque generated at critical speed must not exceed the maximum torque capability of the coupling. That value is shown in the coupling capabilities as the Tkmax for 100,000 cycles or 50,000 reversing cycles.
ConclusionsTorsional systems are a special case for coupling applications. Equipment that is driven directly off the diesel flywheel or by synchronous motors should always be given an extended system analysis for torsional vibrations.
Mounting the Coupling on the Shaft
In order for the coupling to be effective it must first be secured to the shafts of the connected equipment. There are several ways to do this. Most shafts are cylindrical at the point of coupling attachment, but it is not uncommon to find taper shafts, flanged shafts, polygon shafts, spline shafts and others that the coupling must match. The interface of the coupling hub and the shaft must be able to transmit torque and reactionary loads without slipping, without backlash, without contributing to unbalance and without causing vibration. Therefore, the coupling hub-to-shaft interface is an important part of the coupling design.
Most couplings mount with one of two basic types of interface, either clearance fit or interference fit. In clearance fits, the coupling bore is sized so the shaft will slide snugly but freely into the hole. Obviously, this does not provide meaningful friction so additional devices are needed to keep the hub in place and transfer torque. Interference fits, also known as shrink fits, mean the hub bore is slightly smaller than the shaft, and binds tightly to the shaft by means of the size difference. Both types of fit are discussed below in more detail under "Coupling Attachment and Torque Transmission". Before delving into that topic, it's helpful to understand the basics of shaft sizing.
Shaft SizeShaft size is one of the first considerations in coupling selection. Equipment designers, through their trade associations or via common practice, develop shaft standards in the U.S. market. The metric marketplace uses ISO standards for the most part, or JIS standards in Japanese-influenced marketplaces. JIS standards are similar to metric standards. In U.S. standards, the hole-to-shaft fit is expressed on a shaft basis. For the metric markets it is expressed as a hole basis. Shaft basis means the maximum shaft size is basic. Hole basis means the minimum hole size is basic. Basic means that it is the starting dimension to which tolerances are applied. Which one should be used? To quote from the ANSI B4.2 standard "normally the hole system is preferred, however, when a common shaft mates with several holes, the shaft basis system should be used." Most rotating equipment has a shaft that mates with several holes, such as wheels, impellers and bearings as well as couplings. Of course the coupling has to mate with the shafts of two different rotors. Typically the shaft system in the US market is defined by the motor manufacturer through NEMA, by the gear box manufacturer or by the manufacturer of equipment such as pumps or compressors.
On a shaft basis system, if the preferred shaft diameter is 1¼ inches, its maximum diameter is 1.250 inches. The tolerance, if chosen from the motor manufacturers NEMA standard, would be +0.0000 -0.0005. Under these conditions, the acceptable shaft diameter range becomes 1.2500/1.2495 inches. A less precise standard allowing a tolerance +0.000-0.001 would accept a shaft diameter range of 1.250/1.249 inches. The coupling hub bore would take its beginning dimension from the shaft depending on the type of fit being used. It would have a hole (bore) diameter that starts with 1.250 inches, but could be as big as 1.252 inches for a clearance fit or is as small as 1.249 inches for an interference fit. Those coupling hub bores would be defined by a standard such as AGMA 9002-A86 Bores and Keyways for Flexible Couplings (Inch Series).
When the shaft basis system is used in metric measurements, if the preferred shaft diameter is 30mm, its maximum diameter is 30mm. The tolerance (h6 is usually chosen) would be +0.0000 - 0.0130 mm. The coupling hub bore would take its beginning dimension from the shaft, depending on the type of fit being used. For a clearance fit it would have a hole (bore) diameter that starts at 30.000 mm, but could be as big as 30.021 mm (H7). If an interference fit is desired, a P7 hub tolerance is usually chosen. That would result in a hub that is bored to 29.986/29.952 mm. Note that shaft fits are defined with lower case letters such as h6 and bore fits are defined with upper case letters such as H7. The complete description would be 30h6 for a shaft diameter and 30H7 for a bore diameter.
Standard metric coupling hubs are bored, by the manufacturer, to an H7 tolerance unless specified to a different dimension. Other common hub bore tolerances are H6 and P7. They are matched to an equipment shaft that is h6, j6, k6, or m6. The fit of shaft to hub would then depend on the shaft size and would range from clearance to interference. If a specific fit is desired the coupling manufacturer and the equipment manufacturer must agree on the hub bore tolerance. These shaft and bore dimensions would be defined by a standard such as ANSI B4.2-1978 Preferred Metric Limits and Fits or (when available) an AGMA standard.
Coupling manufacturers' catalogs have charts and tables that define shaft/bore combinations for their products. New applications of couplings on newly designed equipment follow the lead of the equipment designer's shaft system and fit requirements. Some difficulty might occur when replacing an old coupling with a new one, especially when the new coupling is a different type or from a different manufacturer.
When facing this situation, first determine whether shaft dimensions are metric or inch type. If measuring with an inch type instrument, leads to a strange dimension, measure again with a metric instrument to see if perhaps it meets a preferred metric size. Nameplates can also provide assistance. It is usually a good bet to assume that the designer used a preferred shaft size, then selected a coupling bore to match. An exception would occur when a shaft has been dressed or repaired.
Shaft shapes other than cylindrical or spline present a problem for coupling selection. The equipment designer and the coupling manufacturer must agree on the applicable standard or the proper interface dimensions. The spline shaft is also a torque transmission device and is discussed later in this section.
Coupling Attachment and Torque TransmissionSeveral different techniques are used to hold the coupling hub to its shaft, secure against reactionary loads, gravity, speed and/or other forces, and to assure effective torque transfer between shaft and hub. In most cases, the functions are interrelated so we will discuss them together. Although several parts of the coupling system affect the amount of torque transmitted (including the flexible element, the way it is attached to the hubs , and any bolted connections between the hubs). This discussion will focus on hub-to-shaft mounting considerations only.
The capability of the shaft/hub juncture to transmit torque is indicated by the length of hub covering the shaft (called length through bore or LTB) and the ratio hub OD to the bore diameter. The length of hub covering the shaft usually ranges from 0.5 to 1.5 times the bore diameter. This depends on the type of interface, with splines able to use small values and clearance fits tending toward the higher values. The hub OD to bore diameter ratio typically ranges from 1.3 to 1.5. That ratio assures that the hub will not split from the loading.
Over-bored clearance hubs fail from fatigue at the corner of the keyway or at the setscrew hole, whereas over-bored shrink fit and bushing-type hubs will split in the process of mounting on the shaft. Sometimes a larger bore (smaller ratio) is safe when the torque is low in relation to shaft diameter. It is best to have the coupling manufacturer check the stress levels on all over-bored specifications.
Most industrial applications use keys to transmit torque from shaft to coupling hub. The number and type of coupling hub keyways vary widely. A single square keyway is most popular on industrial coupling applications up to 6.5 inches diameter, a rectangular key is used on larger bores. Woodruff keys are used on small-diameter shafts. Woodruff keys are thin and have a circular keyseat in the shaft. Two rectangular keyways are also used on very large hub bores. Polygon shafts, D shaped shafts, square shaft ends and hexagonal shafts can all transfer the torque directly to the hub without keys. They are chosen by rotating equipment designers for many reasons including ease of assembly and ease of maintenance.
Clearance fit hubs with keys commonly use a setscrew tightened radially against the key to hold both key and hub in place. In clearance fit hubs used without keys, the setscrew is usually tightened against a flat seat machined into the shaft circumference. Dual setscrews, placed one over the key way and one around the circumference of the shaft in the same axial plane, improve the effect of limiting backlash and holding the hub tight to the key under stop-start conditions. The clearance fit with setscrews generally can be used on shafts up to 4 inches in diameter. In applications with round shafts and no keys, splines or setscrews, torque transmission from the shaft to the coupling hub relies on friction. One frictional method used on some small or low torque devices is the axially split hub drawn together with clamping screws. Another method used on high torque transmission applications is a clamping ring or ashrink disc. These are described in a following section.
Interference (shrink) fits offer a third method for securing the hub to the shaft. They are used with gear and disc and other high torque couplings on both straight and tapered shafts. Interference fits are used with keyed shafts and heavy shrink fits are used with keyless shafts. For keyed fits the torque is transmitted by the key. The friction resulting from the interference keeps the hub in place when subject to various loads such as reactionary forces. Shrink fits involve heating the hub to expand the bore so it will fit over the shaft, then letting it cool and shrink into a secure grip. The temperature differential used for shrink-fit installation usually ranges from 200°F to 400°F. It can be calculated at 160°F per 0.001 inches of shrink per inch of shaft diameter plus 50° to 75°.Differential here refers to the relative temperatures of hub and shaft. Shrink fits are also accomplished by using hydraulic force to expand the hub and slide it on the shaft.
Shrink fits with keys to provide torque transfer use a shrink of 0.003 to 0.00075 inches per inch of shaft diameter. Shrink fits without keys, which rely solely on the fit friction to transmit torque, use shrink rates of 0.0015 to 0.003 inches per inch of shaft diameter. Known as a heavy shrink fit, the keyless type offers an obvious advantage for applications where balance is important. It should also be noted that heavy shrink fits are used with hub materials of sufficient strength. Consult the coupling manufacturer when contemplating this type of fit.
The tapered shaft uses an interference fit. It can be the keyed type or the keyless type. For light interference it is accomplished by pushing a hub (with a matching tapered bore) on to the tapered shaft until the desired interference is reached. The hub can be pushed onto the shaft using a end plate and nut. Medium interference will require that the hub be heated to be installed as the friction can not be overcome otherwise. Heavy interference requires heating or hydraulic force. The hydraulic force expands the hub and pushes the hub onto the shaft. The end plate and bolts or nut will serve to prevent axial movement under loads. Tapered shafts are used for ease of maintenance. A popular tapered shaft device is the mill motor that uses a standard tapered shaft with key.
Shaft Locking Devices: Bushings, and Shrink DisksShaft locking devices can use a combination of friction and keys to transfer the torque from shaft to coupling hub. Shrink discs are more popular in metric applications, but are growing in use on inch applications. All these devices were developed to make the attachment of the coupling hub to equipment shaft easy while still transmitting the proper torque. The devices also prevent fretting, wobble, sliding, and backlash in the hubs.
When extra bore capacity is available, the coupling can use a tapered bushing. When a bushing is used, the shaft size capability of the hub is reduced because the bushing takes up some of the hub's bore space normally available for the shaft. Since the hub must maintain the hub OD-to-bore-diameter ratio range of 1.3 to 1.5, adding a bushing makes the hub-OD-to-shaft-diameter raise to a range of 1.9 to 3.0 and sometimes more.
Bushings require that a taper bore be machined into the hub. The taper can be in either direction depending on the installation. The bushing, essentially a split ring with a matching taper, is inserted from the wide side of the taper and drawn into the bore by setscrews or cap screws threaded into the hub. As the bushing moves into the tapered bore, the taper forces the split halves of the bushing together, which clamps the bushing tightly to the shaft and wedges it into the tapered bore of the hub. Torque is transferred by friction from shaft to bushing to hub. Setscrews and cap screws are tightened to a specified torque value.
There are external and internal frictional locking devices that use opposing tapers to develop radial forces that lock hubs to shafts without using keys. The devices can be installed and removed without heat, are free of backlash, do not require a taper bore in the hub and are able to transmit high torque. The devices are locked to the hub and shaft until deliberately released.
They are commonly called shrink discs, locking assemblies, locking elements or clamping bushings. Shrink discs and clamping bushings are external hub locking devices.Locking assemblies or locking elements are internal devices mounted into a counter bore in the hub. Both use built-in opposing tapers or inclined planes as the means to squeeze the hub onto the shaft or to expand into an internal locking position between the hub and the shaft. The locking assembly is activated by cap screws that are tightened to provide an axial load which then becomes a radial pressure between hub and shaft. Another version uses hydraulic pressure to activate the locking force.
The dimensions of the hub are modified to allow the device to work properly. The devices are a self-contained part that is added to the basic coupling. The hub must be checked to determine if it is strong enough to accommodate the device at the expected torque transmission values.
Flywheel AttachmentsCouplings used for diesel engine applications mount on the engine flywheel. This can be accomplished by either attaching the hub to a flat plate or by making one hub in the form of a flat plate. The plate thickness is designed to transmit the required torque from the flywheel to the coupling. The plate has a pilot OD and holes of the proper size on a bolt circle that matches the engine flywheel. These are in accordance with standards for the dimensions on flywheels that were originally made for use with mechanical clutches. The more popular standard in North America and Europe is the SAE flywheel standard J620d.
Misalignment on these applications is usually limited to a tolerance stackup of the connected parts, all of which are attached to the engine block.
Some OEM applications involve enough production quantity to justify having the flywheel and the coupling hub matched to each other. In these cases the flywheel is not dimensioned for a mechanical clutch, but is sized for a specific inertia. The coupling hub has bolt holes that match a set of holes on the flywheel eliminating the need for a mounting plate or a special plate type hub.
SplinesSplines can be considered as parallel keys cut axially around the end of the shaft, which fit into corresponding grooves, or teeth, cut axially inside the bore of the coupling hub. Their big advantage is ease of assembly, even blind assembly. The spline is used to locate shaft and hub with respect to each other, and to transmit torque from one to the other without intentional relative motion or intentional misalignment. Splines can be involute tooth or straight-sided teeth. Splines are also flat root or fillet root.
Involute teeth are considered to have greater torque carrying capacities than straight-sided teeth. Spline involutes can have 30°, 37½°, or 45° pressure angles (PA), with 30° by far being the most common. That pressure angle is chosen for strength of tooth as well as manufacturing considerations. The pitch diameter (Pd) is usually dictated by the geometry of the parts to be assembled, such as the shaft OD. Many teeth rather than fewer teeth ensure multiple tooth shaft-to-hub contact under load. Splines utilize seventeen specific pitches from 2.5 to 128 teeth per inch of pitch diameter. Pitch is designated with a numerator and denominator such as 4/8 with the ratio for spline pitches as 1 to 2. This means the spline tooth is one-half the height of a standard gear tooth. Under this system, splines always have stub teeth that allow easier assembly. Spline tooth length, on the shaft, is a function of torque-carrying needs and tooth contact. Full contact of all teeth normally would allow a length equal to 1/3 the Pd. Only 50% contact can be counted on, so to ensure sufficient torque-carrying capability, the effective length usually is 2/3 the Pd and can equal the Pd. For comparison, a key-driven shaft-to-hub assembly usually uses a key length of 1 to 1½ times the shaft OD on clearance fits.
Splines are usually slide fits and can be locational as side fit or major diameter fit. Some machine tool and automotive splines fit on the minor diameter because grinding equipment more easily accesses that diameter. These are usually SAE straight-sided splines.
Splines have 4 tolerance levels, designated Class 4, 5, 6, and 7 with Class 5 considered as the base. The hub spline is usually a class 5 tolerance, while the external spline (shaft spline) is modifiable across the other tolerance classes. While these tolerances effect the precision of the teeth they do not dictate the fit between two pieces. Splines have major diameter fits that are always loose. Depending on the shaft spline tolerance the hub will be very loose or tight, but not line to line or tighter. Hub splines always have the same tolerance. Torque is always transmitted side to side between teeth. All tolerances are fixed so that the spacing between the teeth and the width of the teeth are equal at the maximum tooth and minimum space. The shaft is always supposed to fit into the hub no matter the tolerance match.
Splines present a problem for holding the hub in place because their normal tolerances tend to promote axial hub movement and backlash, which can destroy the spline profile, stress other system components and reduce operating efficiency. There is a need to limit axial movement and backlash movement. A collar and/or a snap ring might be used to limit axial movement. Setscrews are used, but tend to cause the shaft spline to be scored, distorted or otherwise damaged. Spline hubs can use the split hub clamping technique described previously, or other proprietary methods.
The coupling should never be the basis for allowable misalignment because it can accommodate far more misalignment than can be tolerated by the other connected pieces of equip. The coupling may have ten times the capability for misalignment that can be tolerated by the rotating equipment. System alignment should be based first on the minimum requirements of the driven equipment, or the driver, and then the coupling.
It is always important that the equipment be aligned as close as possible, keeping within the economics and sophistication of the system. Misalignment is a leading cause of bearing and seal failures, vibration, oil leakage from bearing frames, broken shafts and coupling failures.
Planes of FlexibilityWe could look at the plane of flexibility as pivot points within the coupling. A full flex coupling has two pivot points with one attached to each of the connected shafts. A coupling with a pivot point on one side and a rigid shaft attachment on the other is called a single flex, flex rigid, or half coupling. Note that the pivot point can be in the loose fit between separate parts, such as the hub tooth to sleeve tooth interface in a gear coupling, or in the bending of a continuous flexing element such as used in disc or diaphragm, or link coupling. The flex plane of an elastomeric coupling is within the elastomer itself. For other types of couplings you should refer to the literature, as no two coupling types are alike.
There are three variations to shaft misalignment. They are parallel offset (radial) misalignment, angular misalignment and the combination of angular and parallel offset. Axial displacement is considered a form of misalignment that the coupling may need to deal with.
Radial or Parallel MisalignmentParallel (radial) misalignment occurs when the driving and driven shafts are parallel, but with some offset between their axial centers. Accommodating such offset requires either a full flex coupling (with two flex planes) or two single flex couplings in series. In either case, the greater the axial distance between the two flex planes, the greater the coupling's parallel (radial) capability. Typical full flex couplings include gear, grid, and dual element disc or diaphragm types. Spacer couplings are good for extra radial displacement and close-coupled couplings such as a standard flanged sleeve gear coupling provide the minimum. Spindle couplings and floating shaft couplings provide the maximum capability by further spreading the flex planes. Although the elastomeric type has only one flex plane, the elastomer can distort enough in some cases to provide significant parallel offset capability if it has sufficient resilience. Elastomeric couplings can also be made as spacer or floating shaft types to a limited extent.
Angular MisalignmentAngular misalignment occurs when the axial centers of driving and driven shafts intersect. Flex-rigid or half couplings provide only angular misalignment, because there is only one flex plane. Single element disc or diaphragm couplings provide for angular misalignment only. Single element couplings are used on three bearing systems and on one end of floating shaft systems.
Axial DisplacementAxial misalignment or in-out movement is often associated with thermal shaft growth and floating rotors. Thermal growth is the result of high temperature in the rotating equipment causing an unconfined growth along the length of its shaft. Sometimes a thrust bearing at the coupling end of the shaft will direct axial movement or growth away from the coupling, but sometimes the thrust bearing is at the other end of the shaft. The location of a thrust bearing is a factor in the determination of axial containment. Another well-known source of axial movement is the rotor that seeks its magnetic centers. The coupling must either accommodate axial movement or contain it by transferring the thrust to the bearing system of the rotor. Those that contain it are called limited end float couplings. Sometimes axial thrust is deliberately transferred to another machine through the coupling. Limited end float may or may not be invoked in such a case.
Gear couplings exhibit the best capability to handle axial misalignment or movement. In the gear coupling the hub teeth are free to slide axially within the sleeve while in mesh. Axial displacement is available from either the full-flex or the flex-rigid unit. The amount available depends on many factors, and in fact specials are available for long sliding applications. Gear couplings can be made to limit end float with the addition of a plate and/or button between the coupling halves. The elastomeric coupling is not very tolerant of axial displacement.
Other types such as the diaphragm coupling can flex or stretch to allow some axial displacement. The disc coupling can also, but to a lesser degree than the diaphragm coupling. In both the disc and the diaphragm coupling, axial movement is met with resistance that increases as the displacement increases.
The elastomeric coupling is not a good unit for axial float or axial displacement. Sometimes the unit can slide in one direction, but there are no limiters that stop the slide and the coupling disengages. In other types of elastomeric coupling axial distortion would overload the elastomer when combined with normal misalignment and nominal torque transmission.
Misalignment Comparison by Coupling TypeThe coupling manufacturers' full line catalog is a good starting point for a comparison of misalignment capabilities. Coupling misalignment is usually given in terms of angular misalignment that can be converted to radial or parallel when two flex planes are used.
The conversion of angular to radial misalignment capability is a matter of plane geometry. The radial offset distance is the product of the tangent of the angle of misalignment and the distance between flex points.
In the case of elastomeric couplings, where parallel offset capability results from distortion within a single flex plane, parallel capability is listed in conjunction with the angular.
It is true that misalignment, torque capabilities and coupling life are intertwined. The torque capability of a coupling is reduced when the coupling is misaligned. The reduction in life can come from higher wear and the reduction in torque from high fatigue forces. Misalignment will cause fatigue forces. Some manufacturers publish their torque ratings as a maximum value and require that the user de-rate by some factor to determine usable nominal torque. Others publish their torque ratings at rated full misalignment. When the coupling is selected it should be able to carry the nominal torque while misaligned per the application.
Gear couplings are capable of 1½° of misalignment per gear mesh, although the gear coupling needs some misalignment to push the lubricant to the friction surfaces of the teeth. With tooth modification that misalignment can be increase to as much as 6°. At that much misalignment you may decide to change the pressure angle to strengthen the tooth. That would be the method for spindle couplings. High-speed gear coupling units might be reduced to ¼° per mesh.
Other metallic element flexible couplings have a wide range of angular capabilities. The flexing link coupling is suitable for 6° while the torsional spring coupling can be used up to 4½° of angular. Published data for the disc and diaphragm coupling types range from ¼° to 1°.
The elastomer couplings such as the jaw types and the unclamped donut elastomer in shear type couplings are limited to 1° or 1½°. Donut type elastomer couplings are suitable for 3° of angular misalignment.
The coupling manufacturers' catalog or published information should always be consulted for actual values of misalignment and torque derating requirements. Misalignment capabilities and torque capabilities are interrelated.
Acceptable MisalignmentA high speed, high power system requires close alignment, general-purpose equipment is aligned to a looser specification. For example high speed equipment (running at 3000 RPM or more), needs alignment to .0005 inch (or better) per inch of flex point separation. General-purpose equipment can be acceptable at .001 inches per inch of separation. Smaller values will improve the operation but should be consistent with the equipment manufacturer recommendation.
The preceding recommendations are usually used with "close coupled" equipment. When spacers and floating shafts are chosen for the ability to allow radial displacements of two pieces of equipment, these rules of thumb would not necessarily apply.
It is sometimes necessary to have large amounts of parallel displacement built into the equipment installation. Special design alterations to the couplings and the connected equipment can also be required in those cases. The drivers and driven equipment must have sufficient design strength to deal with the increased reactionary load. Equipment with spacer couplings or even floating shafts may be designed for ease of maintenance rather than for the increased misalignment capability that would be possible. In such situations, always remember that acceptable system misalignment is still dictated by equipment capabilities, not the coupling's ability to withstand the misalignment.
Exceeding acceptable misalignment contributes to vibration problems. Severe misalignment will impose a heavy vibratory force on the equipment. Large amounts of parallel misalignment are acceptable only on machines operating at slow speed. High-speed machines, even those with spacers and floating shafts must be aligned closely to limit the vibration and the reactionary loads.
Reactionary Loads from MisalignmentShaft to shaft misalignment causes couplings to impose reactionary loads on the connected equipment. The greater the misalignment the greater the reactionary load. Different types of couplings produce different reactionary loads.
In gear couplings, reactionary forces result from the sliding friction of the tooth to tooth movement. The sliding friction is considerable when dealing with metal-to-metal contact. There is some lubricant but it is a very thin film.
When a gear coupling is misaligned, the friction or drag forces become a bending moment on the system. The bending moment reactionary load can be 10% of the value of the torque transmitted. At that value it is many times the reaction load of a disc coupling.
Disc and diaphragm couplings utilize the flexing of thin metal elements to handle the misalignment, both angular and axial. The thinner the element the less force it takes to bend the metal. The force to bend becomes the force of the reaction in the classic physics truth about equal and opposite reaction forces. Link couplings and spring couplings also provide a reactionary force in proportion to the loading force.
Elastomeric couplings exhibit different kinds of forces depending on the type of coupling. Elastomeric couplings in shear will impose a thrust (axial) load on the adjacent machine parts. Distortion of a elastomeric donut or misalignment in a jaw coupling will impose a bending moment. Tire type couplings exert a thrust load also as the centrifugal force acts on the tire. Reaction forces in the elastomer coupling are a function of the resiliency and may also have some friction components. The friction is not unlike the sliding of gear teeth and occurs on couplings like the jaw units.
The reactionary loads include the weight of the coupling. Because weight usually reflects size, the more power intensive couplings (those with higher torque capacity for smaller size, such as gear couplings) will have the advantage of lower reactionary forces at comparable torque loads. Large size couplings, (less power intensive types) may also be made from light materials to reduce their weight and their inertia.
The loads imposed by the coupling are related to the point of loading. The pivot point or flex plane is the location of the loading for the coupling. Those couplings that move the pivot point closer to the next available bearing are termed "reduced moment" couplings, because they reduce the reactionary load imposed on the bearing. After all it is the machinery bearings that ultimately carry the extra loading.
Alignment Occurs at InstallationAlignment is accomplished at the final installation point of the equipment. Alignment is done when the coupling is installed. Equipment can be aligned when the driver and the driven equipment is assembled at the manufacturer, however the alignment still must be checked and likely adjusted at the final installation.
When aligning the coupling, the installer must take into account the conditions affecting the rotating machinery. Not only is it important to understand where the equipment is at standstill, but also it is important to know where the equipment will move when in operation. For example, hot operating equipment grows when it is brought up to operating temperature. The coupling is aligned cold at one location with the expectation that movement will bring it to closer alignment. The initial cold alignment should be within the coupling and rotating equipment capabilities. Sometimes the determining the position of the rotating equipment at operating temperature is the most difficult part of alignment and installation.
A motor rotor will seek its magnetic center causing a thrust that must be transmitted through the coupling to the next thrust bearing. That axial displacement is a misalignment problem for the equipment.
In addition to temperature considerations the rotating equipment alignment can be affected by tolerance stackup, pipe loading, the foundation and conditions such as bent shafts or soft foot. Before aligning the equipment it is best to check for those occurrences.
Alignment can be measured by use of a straight edge and feeler gage (or calipers or taper gage). That would be the simplest method, but the least accurate. Dial indicators are used with the reverse indicator and the face and rim method. The most modern method uses a laser alignment system. Each method has its strengths and weakness and all can be satisfactory depending on the skill of the installer.
There are many books and papers written on the "how to" of alignment. Some are listed in the bibliography. The reader should see those for details of alignment. The following paragraphs are for the purpose of general descriptions of the processes.
Methods to Check Alignment
Straight Edge and Taper Gage Alignment
A straight edge is used to determine the shaft offset by eye. It is used for both vertical and horizontal planes. The taper gage (or calipers or feeler gage) is used for angular misalignment. The shaft separation or "BSE" dimension is measured with a ruler. It is a trial and error process.
Dial Indicator AlignmentDial indicators are used with the reverse indicator or the face and rim method. The dial indicators are mounted on the shaft opposite to the reading to be taken. These indicators are accurate to ± 1 mil.
In the reverse indicator method, readings are taken from coupling hub on shaft "A" to the rim of the coupling hub on shaft "B". A second set of readings are taken from the coupling hub on shaft "B" to the rim of the coupling hub on shaft "A". Both sets of readings are plotted on graph paper or become the input to a personal computer program. With the proper calculations in plane geometry, the misalignment of both parallel offset, and angularity of the shafts can be determined.
The face and rim method uses the dial indicator mounted on one coupling hub to take readings on the face and the rim of the second coupling hub. Again with graphical plotting or a computer and plane geometry the misalignment of both types can be determined.
This method can be very accurate if done with graphical assistance or computer assistance. Other commercial mechanical and electrical devices can obtain the results by measuring the positions of two shafts.
Laser Beam AlignmentLaser beam alignment uses the laser to replace the dial indicator. It is a little more accurate, but is much more costly. Included with the laser package is the means of direct input to a computer program that calculates the moves necessary to align the equipment. Lasers are accurate to ± 3 micron or better.
The laser is a light beam that is very narrow and focused. The beam generating equipment is mounted on the equipment shaft and aimed at a device on the opposite shaft. The device can be a reflector or can be the photodiode target cell that will generate a voltage. The amount of voltage that is generated will depend on the position of the light beam as it hits the cell. A reflector will cause the beam to return to a target cell that is mounted with the laser generator. The generated voltage becomes the input to a system that calculates the misalignment and needed corrections.
ConclusionsProper installation and alignment procedures are included with the installation instructions of rotating equipment and couplings. Often the coupling manufacturer can provide further guidelines for installing the coupling and aligning the rotating equipment. There are many published papers and pamphlets on the subject. High speed high powered equipment often is ordered with the services of installation start up supervision from the manufacturer. That service is well worth its extra cost. Also there are many companies within specific industries that provide alignment service.
There are many combinations of angles and spacing which can be calculated by plane geometry to obtain the ideal situation for an application. Always keep in mind that equipment should be aligned to the rotating equipment manufacturers' standards and requirements, not the coupling manufacturers. When operating misaligned, the coupling can transmit reactionary loads and vibration that are within the coupling capabilities, but not the equipment capabilities.
Coupling (Un) BalanceBalance in any rotating system is important because without it the system vibrates. Vibration in rotating systems invariably results in problems ranging from premature wear to severe damage and failure in all parts of the system. But perfect balance is not achievable, so instead we talk about how much unbalance is acceptable. The rotor designer and rotating equipment specialist determines the amount of acceptable unbalance. There are standards to help them. The standards have been developed through empirical data generated over years of service conditions as well as design and testing.
Vibration can occur in lateral, torsional or axial directions, but only lateral vibration involves coupling unbalance so the others will not be discussed here. Lateral vibration refers to sideways movement (radial, or perpendicular to the axis of rotation). It occurs regardless of whether the axis is vertical or horizontal. Reducing the rotor unbalance is an important method of reducing vibration.
Because coupling unbalance forces contribute to vibration, reducing coupling unbalance can be an important means of preventing these kinds of problems. But we cannot discuss coupling unbalance without first understanding a little bit about the relationship between vibration, unbalance, and critical speeds.
VibrationVibration is cyclic force acting on the rotor. One of the causes of lateral vibration is an unbalanced mass within the rotor, the coupling or both. Other forces that are vibratory in nature include vane passing frequencies, misalignment, gears meshing, external system variations, pressure pulsation, and internal machine functions. Coupling unbalance can be a serious contributor to vibratory forces for two reasons. First, the coupling is overhung outside the bearings where the shaft might be more vulnerable to unbalance forces. Second, the coupling might be added later in the application without the benefit of being balanced to the same standards as the rotor. Coupling and rotor unbalance can be corrected by manufacturing process or by utilizing a balancing machine. Other vibratory forces should be minimized or damped by the equipment system.
Unbalance, where the center of mass of the rotor is not the same as the center of rotation, causes the shaft to whirl as it rotates. The force from the unbalance mass increases with increasing speed.The unbalance force is equal to:
F = 1.77 x (RPM/1000)^2 x U
F = the unbalance force in poundsRPM = operating speed in revolutions/minute.U = the unbalance in ounce-inches
A 50-pound coupling with center of mass .002 inches from its center of rotation has 1.6 oz-ins of unbalance. 1.6 oz-inches of unbalance at 1800 RPM are 9 lbs. of untamed force. That same unbalance at 3600 RPM becomes 36 lbs. of force. It shows as vibration and wear on the rotor bearings. A .004 TIR on the bore concentricity would displace the mass center by .002 inch.
As the system rotates, the combined unbalance of the rotor and coupling causes the shaft to deflect or whirl. The system tries to rotate about its mass center rather than its shaft centerline. The magnitude of the unbalance forces increase as shaft RPM increases, and at some point, these forces become strong enough to seriously damage the bearings and/or fatigue the shaft. Still higher magnitudes can cause the machinery to jump around (or try to), and that will lead to fatigue failures in machine housings and moorings. A severe unbalance will impact the rotor and bearings at virtually any speed and must be avoided. The cyclic forces are always present, and will cause a vibratory response. That vibratory response will be reflective of the ratio of running speed to critical speed. At a ratio of 1:1 the response is infinite and resonance occurs.
Critical SpeedAll systems, including the rotating shaft with one half of a coupling have at least one lateral natural frequency. When this frequency (in cycles per minute) matches RPM, it may also be called the critical speed. The critical frequency or speed is defined as the point where the kinetic energy of the rotating masses equals the potential energy stored in the shaft acting as a spring.
The natural frequency becomes a critical frequency, which is resonating, when an external force is applied cyclically at the same frequency. Since the system is rotating at this frequency or RPM, there are many available forces to trigger resonance. Since the system must be triggered or forced, it is called a forcing frequency. At resonance the system could fly apart.
The lateral natural frequency is a function of the deflection of the shaft and the masses attached. The amount of deflection (amplitude) at any point along the shaft will depend upon the location and weight of rotor attachments, length of shaft span, shaft overhang, diameter of shaft, and material stiffness. Shafts with relatively large diameters and short spans don't deflect much, and so are termed "rigid rotors". While we would like all rotors to be rigid, it may not be possible for economic or geometric reasons. Shafts with smaller diameters and larger spans deflect more, and so are termed "limber rotors".Generally the more rigid the rotor, the higher the lateral natural (critical) frequency. It will take a greater force to cause deflection. The more limber the rotor, the lower the natural (critical) frequency will be. Because of all the variables involved, lateral vibration can become complex. There may be more than one point of deflection along the shaft, both between the bearings and in the overhung portion, and thus more than one natural (critical) frequency. So, the terms rigid and limber have roots in flexibility just like they do outside of engineering.
Another type of critical speed is the self-induced critical speed that is caused by unbalance. The deflection from the unbalance forces increases with increasing speed until the critical frequency is reached. This frequency is the same value as the natural frequency of the shaft system. At this point the shaft will whirl like a child's jump rope. This critical and the response is self-induced by the unbalanced mass, as opposed to the lateral frequency that needs to be triggered by external forces. Unbalance does not change the natural frequency or the critical frequency of the system. It does add another force to the system. The added force can be large in comparison to other forces. The unbalance force can be reduced by efforts to balance the rotor and the coupling as opposed to trying to dampen the forces.
When the lateral critical frequency and the self-induced critical frequency are equal to the operating or forcing frequency (operating speed) the vibratory response is infinite and the system resonates (vibrates out of control). Note two different effects are happening at the same time. One is the natural frequency being forced by an external trigger, and the other a induced force from the out of center mass. The combined effect is disastrous to the system. There are two solutions to the problem. The first and obvious one is to always operate the system below the critical speeds. The second would be to operate above the first critical and below the second critical. The driver must have enough torque to accelerate through the critical very quickly, the unbalance must be very low and the forcing or trigger forces must be damped in order to pass through without incident.
Rigid rotors, with deflection so slight as to approach zero, have natural frequencies so high that normal operating speeds typically don't come anywhere close to critical speeds. These are usually rugged, slowspeed systems such as found in mill applications. They are not affected by adding a normal amount of coupling unbalance. Limber rotors and rotors with overhung loads are more likely to have natural frequencies uncomfortably close to their normal operating speeds. These are often found in multistage pump compressors and other equipment that operates at speeds of 3000 RPM or higher.
The term "sensitive" is used to identify systems in which adding coupling unbalance can adversely affect rotor unbalance. For example a relatively light coupling mounted on a rigid (stiff) rotor is not a sensitive situation. On the other hand a relatively heavy coupling mounted with a long overhang to the first bearing, with an overhung machine load and operating at a high speed (close to critical most likely) is a sensitive situation. Everything else is in between.
Typically the lateral critical speed calculation for each piece of rotating equipment is calculated by its manufacturer. This means that critical speeds for driving and driven equipment are calculated separately. For this reason it is common practice for the rotor manufacturer to consider one-half of the coupling as a part of the rotor for purposes of critical speed calculation. The coupling manufacturer could calculate the critical speed of the spacer piece or floating shaft in the center of that type of coupling. To determine the critical speed of the complete equipment train is a complicated issue, well beyond the scope of the coupling supplier.
Issues for Coupling BalanceThe coupling is an integral part of the rotating system and, as a part of the system, it must be balanced to the same criteria as the other elements of the system. The coupling cannot be the balancing device to solve problems in these two systems, but neither should it be the unbalancing device.
Balancing is an expensive option for the coupling and possibly the system. It usually does not pay to put severe unbalance restrictions on a coupling, unless it is a coupling for a high-speed sensitive rotor. In those cases, reducing the couplings vibratory forces due to unbalance is usually worth the expense, because those forces increase by the square of the speed. A doubling of the speed increases the force by four times. High-speed equipment therefore has more reasons to include strict balancing criteria. Long floating shaft couplings should also be considered a candidate for a critical speed check, and if necessary, a more strict balancing criteria.
Sometimes vibration can be damped with bearings, bushings or fluids (in the case of pumps). However, even with some damping and good balance criteria, the rotor should never be operated near a lateral critical speed.
In equipment with sensitive rotors, it's the rotating equipment designer's responsibility to keep the bigger picture in view. The couplings should be selected not only for size and torque/misalignment capabilities, but also for the more subtle criteria of weight, balance, moment arm, spring stiffness and reactionary loads.
Coupling balance should be accounted for at the onset of the equipment selection and design. If a machine begins vibrating after operating smoothly for a period of time, it is not because an unbalanced coupling suddenly came to light. The first suspect in that case would be the alignment which should always be checked first. Second, find out what changes occurred between smooth operation and vibration. It is important that all changes be accounted for in the system. There is also a good chance that something has become worn or has broken. Discovering the problem could be so costly in time and resources that commissioning a vibration analysis could be cheaper in the long run. Don't assume that reducing a coupling's unbalance will solve the problem by itself.
Balancing the CouplingGood coupling balance originates in the design and manufacturing stages. As a practical matter, however, all standard couplings are manufactured with a certain amount of unbalance called "inherent unbalance". The final exact unbalance of the coupling will be related to this, but cannot be determined until each half of the coupling is ultimately mounted on a rotor shaft that, in effect, becomes part of that half of the coupling.
The amount of unbalance permitted in new couplings is a matter of tolerances applied at the factory and balance categories selected by the coupling manufacturer or the coupling specifications provided by the system designer or user. AGMA, ISO and API have developed standards for coupling balance.
Each coupling has a unique potential unbalance resulting from the displacement of the center of mass with respect to the center of rotation. Displacement of mass can occur by non-uniform density of the material, by non-concentric shapes, non-symmetric geometry such as keyways or by machining tolerance stack up. Elastomeric couplings have the problem of parts that change shape under load and at speed, which has the effect of mass displacement
To understand the causes of mass displacement, envision the coupling as a perfectly shaped cylinder on its side. Each end is a perfect circle, the top and bottom sides are straight lines of equal length, and the cylinder is solid with uniform density throughout. This cylinder would have a center of mass in the horizontal direction that is a line lengthwise through the geometric center of the cylinder. If we attach a perfect axle to each end of the cylinder exactly in the middle of the end circle, we could then rotate the cylinder on that axis with no unbalance forces. The center of rotation would be coincident with the center of mass. All centrifugal forces resulting from rotation of the mass particles would be equal and opposite in a radial line from the center.
The perfect cylinder described above cannot be practically achieved. The material, if a casting, would very likely have internal voids or spots of high density. Our axles, as represented by the coupling's shaft bore holes, would not be placed exactly in the geometric center of each end and coincident with each other. Instead, each would be slightly off center by a tolerance normally allowed for manufacturing capabilities. With these bores not exactly concentric with the outer wall of the coupling cylinder, there is now more coupling mass on one side of the shaft than on the other. Because the shaft is the center of rotation, the side with more mass will generate a higher centrifugal force than the other side, resulting in an unbalance force, or vibration.
In our perfect cylinder the end circles were perpendicular to the side walls. Again, manufacturing tolerances would allow the ends to be somewhat skewed. That means one side of the cylinder is longer than the other side. The long side has more mass so our mass center is not at the geometric center anymore and possibly even further from the center of rotation as defined by the axles. That means more out-of-balance centrifugal forces.
If our cylinder were made up of many pieces bolted together, as many couplings are, every bolted joint would have the potential to displace the center of mass with respect to the center of rotation. That is because the bolt holes have a clearance with respect to the body of the bolt, the bolt holes are not perfectly placed on the bolt circle and the bolt circle is not perfectly concentric with the center of the part. More mass displacement will occur.
Every piece of the bolted-together cylinder would have its own manufacturing tolerance that would stack up and cause the center of mass to be different from the center of rotation. The unbalance forces they create would be spread all around the center of rotation, some canceling each other out and others reinforcing each other. With good luck all forces would be equal and opposite and cancel each other out. No unbalance forces would exist. Of course the situation with the greatest chance is that there will be a net residual non-canceled force called unbalance.
When the coupling is assembled, our cylinder will also become longer along the axis compared to the diameter perpendicular to the axis. That complicates the unbalance, and is of special concern with spacer type couplings. Our opposing forces are no longer in line with each other so they become a force couple rather than canceling each other out. Ineffect we have two unbalance forces acting on the coupling one at each end of a long span. The vibration becomes complicated.
These special concerns regarding multi-piece bolted together couplings emphasize the importance of match marking the coupling. Match marking allows the parts to be put back exactly where they came from when disassembling and reassembling the coupling for maintenance. Thus eliminating the chances of introducing more unbalance into the system.
The best way to resolve the balance issue calls for design of the coupling parts with low manufacturing tolerance consistent with cost and with the application requirements. The design and manufacturing features for good balance include bores and cylindrical surfaces that are concentric with the center of the coupling, hub faces that are perpendicular to the center line, and bolt circles that are true positioned. The design can also include pilot fits to place the components in the proper plane. Bolts and bolt holes can be tight fits too. Some applications allow for high-cost couplings that have tighter tolerances and fits. On those high cost special couplings everything including the weight of individual bolts, the keys, and the keyways must be accounted for. Remember that the unbalance force is a function of speed, so that the need for tighter tolerances to minimize unbalance is more important for high speed rotors.
Balance Machine BalancingWhile close attention to design and manufacturing can result in very low unbalance forces, it is sometimes necessary to fine-tune or further reduce the coupling unbalance. This is done by adding or removing weights (mass) equal and opposite to the unbalance, as directed by the use of balancing machines. The balance machine is not a substitute for good design practice and manufacturing tolerances. It can provide about 5% to 20% adjustment on the balance built into the coupling through design and manufacturing. Some couplings also include devices that can adjust the unbalance once the coupling is mounted on the rotor. This is called trim balancing, and is used on couplings for very high speed or sensitive rotors.
The coupling manufacturer talks about single plane and two plane balancing, static and dynamic balancing, or spin balancing. Let's examine these terms.
Static balancing consists of placing the part with its shaft on knife-edges so that the heavy side rotates to the bottom. By trial and error one can place weight or remove weight to the point where the piece no longer rotates to the "low side". This is a crude method, and seeks to compensate for unbalance as a single point in a single plane of rotation.
A more sophisticated method is to spin the coupling on a turntable or between two bearing supports, with either device spring-loaded so as to be displaced as the heavy side passes, and instrumented to report the amount and location of the unbalance mass. Turntables are single plane units used for testing full round coupling components whose diameter is equal to or larger than the parts axial length, such as hubs, flanges or collars. A dual bearing balancing machine is used for testing coupling assemblies in which axial length may exceed diameter. The two-plane approach is needed to account for the unbalanced-mass "couples" that are formed on long parts as mentioned earlier, which do not cancel each other out but act independently on whichever end of the coupling is closest.
Mounting the coupling or the coupling part on the balance machine is an art in itself. The arbor, mandrel or shaft substitute could be "unbalanced" and must be accounted for in the balance work. The coupling has to be rigidified to allow the balance machine to work properly. Gear couplings must be made with a tight fit on the major diameter for the balancing process and then relieved for actual installation in a system. Disc couplings use a clamping screw to hold the discs tight. Some couplings cannot be balanced as an assembly, only as components.
After components are balanced, the high sides are marked so that the coupling can be assembled with high sides properly located versus other high sides. The couplings can also be assembly check-balanced once the components are assembled. Component balanced couplings should only utilize substitute parts that are balanced to the same criteria.
When a coupling has been balanced as a complete assembly, its components are "match marked" to guide future assembly. Assembly balanced couplings cannot have substitute parts installed unless the whole assembly is re-balanced. Otherwise the balance will be compromised.
When a coupling has been machine balanced, and is still mounted in the machine, the result is described in terms of how much residual unbalance remains. Once removed from the machine the unbalance is called potential unbalance.
Balance StandardsThe question for the coupling manufacturer, coupling specifier or coupling user becomes one of how much unbalance is acceptable for their application. One way to determine that is to consult one of the several standards that have been developed.
ISO Standard 1940/1-1986 (E) Mechanical Vibration-Balance Quality Requirements of Rigid Rotors gives recommendations for rotor unbalance. Representatives of the coupling industry wrote AGMA Standard 9000-C90 Flexible Couplings-Potential Unbalance Classification with input by users and other machinery designers. It is now recognized as the definitive standard for most industrial coupling applications Several arbitrary standards exist such as the balance criteria in the API Standard 671 Special-Purpose Couplings for Petroleum, Chemical, and Gas Industry Services. The standards, whether written for rotors or couplings are based on empirical data and good design practice developed over many years.
The ISO standard sets out to define an acceptable residual unbalance criterion for rotors. That criterion includes a balance quality grade or "G" value. "G" values range from .4 to 4000 in steps that are factored by a 2.5 multiplier. "G" values are designated by rotor types based on long term practical experience. The "G" value is the product of specific unbalance value and the angular velocity of the rotor. Similar rotors would have the same "G" value, but as the speed is increased the specific unbalance value would decrease directly with the speed increase. Residual unbalance would be the product of the specific unbalance and the rotor mass. Extremely low values of "G" such as G.4 or G1 may only be reached by special procedures.
The standard also proposes a second and third means of establishing the balance quality requirement. It could be defined by experimental determination (measurement) in cases of mass production applications, or on permissible bearing forces.
The ISO standard provides information on defining the balance problem and allocation of the residual unbalance to the correction planes. Although it is a standard designed to deal with rotors, it is commonly applied to couplings as well.
The AGMA standard approaches the unbalance from a coupling perspective recognizing that the coupling will end up suspended between two rotors. AGMA also establishes a balance criterion or classification system. The classes ranging from 4 to 11 identify the maximum potential displacement of the center of mass with respect to the axis of rotation. The balance class is selected by a series of charts that take in to account the coupling half weight and speed and then the system sensitivity to coupling unbalance. Coupling half weight is used, acknowledging that half of the coupling is attached to each rotor.
The AGMA standard calculates the potential unbalance starting with the uncorrected coupling, the component balanced coupling, or the assembly balanced coupling. The method used is to calculate the unbalance contributions from various sources and combine them by taking the square root of the sum of the squares. Included in the unbalance contribution is the residual unbalance.
This method accounts for the unbalance contribution from all the items that cannot be taken care of on the balance machine. It is a potential unbalance that is greater than the residual unbalance and different from the unbalance when mounted in the system.
The AGMA standard makes it possible for the coupling manufacturer to calculate the inherent unbalance of its coupling line by using the design and manufacturing allowances. That could also be done by a statistical analysis of unbalance data for the couplings as manufactured.
A discussion of the types of unbalance is included in the AGMA standard. They include static unbalance, couple unbalance, dynamic unbalance and quasi-static unbalance. Appendixes of the AGMA standard provide calculation examples.
Some coupling manufacturers will publish an ISO or AGMA balance class in their catalogs and offer upgrades for applications that require a better class. The balance standards are designed to be a starting point or an ending point. Used as a starting point, they allow the manufacturer to identify the balance class or quality of a certain coupling line. Used as an ending point, they enable the user to specify the grade that the manufacturer must meet.
Some user groups have established arbitrary standards. An example is the API Standard 671. It calls for a low-speed coupling to be component balanced and the assembly of the components to not exceed AGMA grade 9 residual mass center displacement by calculation. A high-speed coupling is to be component-balanced to the larger of 4W/N, 0.0008W or 0.01 oz-ins and assembly-check-balanced to 10 times that value. An alternate method is for an assembly balance to the larger of 4W/N, 0.0008W or 0.01 oz-ins. This arbitrary specification also calls out repeatability checks and allows for trim balance. The unbalance is separately stated per each of the two balance planes. "W" is the coupling weight and "N" is the speed. Refer to the standard for units of measure either metric or imperial.
The API 671 standard was developed for high-speed, high-power couplings used in refinery applications. Those couplings are usually very sophisticated types such as disc or diaphragm couplings.
For more information and discussion on vibration and balance or critical speed, the reader should consult one of the many fine articles or books on balancing. Some are listed in the bibliography.
Often used for reference in this handbook
American Gear Manufacturers Association
AGMA 9002-A86 (1986) - Bore and keyways for flexible couplings (inch series)
AGMA 9004-A99 (1991) - Flexible couplings - Mass Elastic Properties and Other Characterisitics
AGMA 9003-A91 (1991) - Flexible couplings - Keyless Fits
AGMA 9001-B97 (1997) - Flexible couplings - Lubrication
AGMA 9000-C90 (1990) - Flexible couplings - Potential Unbalance Classification
AGMA 922-A96 (1996) - Load classification and service factors for flexible couplings
AGMA 510 - Nomenclature for flexible couplings
American Petroleum Institute
Pumps for refinery service API 610 Edition 1990 & 1998
Special-purpose couplings for refinery service API 671 2nd & 3rd Edition 1990 & 1998
Standards for Couplings ANSI
ANSI B4.2 (1984) - Preferred Metric Limits and Fits
ANSI S2 19 (1989) - Mechanical vibration balance quality requirements of rigid rotors
DIN 740 1&2 (1986) - Power Transmission Engineering Flexible Shaft Couplings Part 1 & 2
ISO 1940/1 - Balance quality of rotating rigid bodies
ISO DIS 10441 Draft - Flexible couplings for mechanical power transmission - special purpose applications
ISO DIS 14691 Draft - Flexible couplings for mechanical power transmission in general purpose applications
Boylan, W. Marine application of dental couplings. Society of Naval Architects and Marine Engineers (SNAME) (1966)
Calistrat, M. Design of Coupling Enclosures. Turbomachinery Symposium (1985)
Calistrat, M. Extended gear coupling life (part 1 & 2). Hydrocarbon Processing (1978)
Calistrat, M. Flexible Couplings, their Design Selection and Use. Caroline Publishing (1994)
Calistrat, M. Hydraulically Fitted Hubs, Theory and Practice. Turbomachinery Symposium (1980)
Carter-Garvey-Corcoran. Baffling and temperatures predictions of coupling enclosures. Turbomachinery Symposium (1994)
Corbo, Malanoski. Practical Design Against Torsional Vibration. Proceedings 25th Turbomachinery Symposium (1996)
Gibbons, C.B. Diaphragm couplings in Turbomachinery. Machinery Vibration Institute (1979)
Gibbons, Mancuso, Munyon. Tutorial on couplings. Turbomachinery Symposium (1989)
Goody, R.E. Some coupling Questions Answered (1996)
Jackson, Moore. Alignment is Plain Geometry. Turbomachinery Symposium (1996)
Jones, J.E. Torsional Vibration in Diesel Engine Drives. Draughtsman's and Allied Technicians' Association (1962)
Mancuso et al., Short Course on Couplings. Turbomachinery Symposium (1998-1999)
Mancuso, J. Couplings and Joints, Design, Selection and Application. Marcel Dekker Inc. (1986)
Mancuso, J. General Purpose vs. Special Purpose Couplings. Turbomachinery Symposium (1994)
Mancuso, J. Lets Try to Really Understand Coupling Balance. ASME (1996)
Mancuso, J. Manufacturers World of Coupling Potential Unbalance. Turbomachinery Symposium (1984)
Mancuso, J. Retrofitting Gear Couplings with Diaphragm Couplings. Hyrdocarbon Processing (1988)
Mancuso, Ziberman, Corcoran, D'Ercole. Flexible-element couplings; how safe is safe? Hydrocarbon Processing (1994)
Piotrowski, John. Understanding and Using Shaft-to-Shaft Alignment Measurement systems. Pumps and Systems (1999)
Piotrowski, John. Shaft Alignment Handbook (2nd Edition) Marcel Dekker (1995)
Rivin, E.I. Design and Application Criteria for Connecting Couplings Transactions of the ASME Journal of Mechanisms, Transmissions and Automation in Design (1986)
South-Mancuso. Mechanical Power Transmission Components. Marcel Dekker (1994)
Wolford, C. Retrofitting Turbomachinery with High Performance Flexible Dry Couplings. Turbomachinery Symposium (1990)
Wright, John. Transient Torsional Vibration in Synchronous Motor Drives. ASME paper 75-DE-15 (1975)
Hytrel(R), Viton(R), and Zytel(R) are all trademaarks of E.I. du Pont de Numours & Co.
Thanks to John Peters for his suggestions and assistance in editing the handbook.
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