Coupling Handbook

Forward

As CEO and owner of Lovejoy, Inc., I am very proud to be leading our company into its second 100

years. After considering the many suggestions for ways to commemorate our 100th anniversary, I

felt the best one was to create something that could have lasting value to our industry and our

customers. From this goal came the idea of our "Coupling Handbook".

During our first 100 years of existence, Lovejoy engineers, product managers and field service

people have accumulated a lot of knowledge about flexible couplings, including practical

experience not found in textbooks. Our Handbook is intended to transfer that knowledge to the

people who can best make use of it -- the designers, machine builders and maintenance people

who work with couplings every day.

Like Lovejoy itself, this handbook will always be a work in progress. There is always more to learn.

To that end, we welcome input from you, the reader, as to how we can improve the contents of this

book in the future. We want it to become a living, changing document that will be updated over the

years to better assist its readers with the selection, installation and maintenance of all types of

flexible couplings. Between successive editions, we will post new and updated material on the

industry-wide informational website we sponsor: couplings.com. Parts of this book also will be

available for training purposes on Lovejoy's website: lovejoy-inc.com.

I hope you will find our efforts informative, helpful and worthwhile, and that you will offer your

comments, knowledge and experience to help us continually make it better.

With thanks to you for making our success possible.

- Mike Hennessy, Chief Executive Officer

Preface

Lovejoy is very proud to be celebrating its 100th anniversary at the start of the new millennium. To

commemorate this occasion, we created a handbook for those people who are involved with

mechanical power transmission, and specifically with the general purpose couplings used in that

field.

The majority of those who leave engineering school are confronted with daunting challenges. For

one, they must bridge the gap between theoretical textbooks and the practical realities of design

engineering in industry today. Engineers spend only a small portion of their time dealing with

flexible couplings.

With the notable exception of gear couplings, industry-wide common designs for flexible couplings

really do not exist. Each coupling designer developed a coupling with a unique geometry and set

the ratings based on that coupling's abilities. This contrasts with other power transmission

components such as chain, v-belts, motors, and bearings where standards exist. Each of the

manufacturers produces these products to standards and in many instances even use the same

nomenclature.

The goal of this handbook is to assist you with the process of sorting out the myriad of coupling

styles that exist to select the one best suited to your application. This handbook is not a textbook.

There are several of those in print which do a great job and are very useful for coupling designers.

What we are attempting to do is to provide down-to-earth useable knowledge. We want to arm you

with information that you need to utilize the variety of styles that exist in flexible couplings to your

best advantage and solve real world problems. Lovejoy has been manufacturing couplings since

1927. More importantly, we have the greatest breadth of coupling types offered by any single

manufacturer in the world. We have the applications experience with couplings to talk about all the

most popular designs out there, even the ones we don't sell. Since flexible couplings are our

strategic focus, we feel you will find this handbook to be a valuable resource.

I. Introduction

A. Why a Flexible Coupling?

A flexible coupling connects two shafts, end-to-end in the same line, for two main purposes. The

first is to transmit power (torque) from one shaft to the other, causing both to rotate in unison, at

the same RPM. The second is to compensate for minor amounts of misalignment and random

movement between the two shafts. Belt, chain, gear and clutch drives also transmit power from

one shaft to another, but not necessarily at the same RPM and not with the shafts in approximately

the same line.

Such compensation is vital because perfect alignment of two shafts is extremely difficult and rarely

attained. The coupling will, to varying degrees, minimize the effect of misaligned shafts. Even with

very good initial shaft alignment there is often a tendency for the coupled equipment to "drift"

from its initial position, thereby causing further misalignment of the shafts. If not properly

compensated, minor shaft misalignment can result in unnecessary wear and premature

replacement of other system components.

In certain cases, flexible couplings are selected for other protective functions as well. One is to

provide a break point between driving and driven shafts that will act as a fuse if a severe torque

overload occurs. This assures that the coupling will fail before something more costly breaks

elsewhere along the drive train. Another is to dampen torsional (rotational) vibration that occurs

naturally in the driving and/or driven equipment.

Each type of coupling has some advantage over another type. There is not one coupling type that

can "do it all". There is a trade-off associated with each, not the least of which can be purchase

costs. Each design has strengths and weaknesses that must be taken into consideration because

they can dramatically impact how well the coupling performs in the application.

This handbook will be a guide to assessing the features and limitations of the many standard types

of couplings on the market. Before we enter into a discussion about all of the evaluation factors to

consider in selecting the right coupling type, let's review some basic terminology that will be used

in this handbook.

B. Basic Terminology

ANGULAR MISALIGNMENT: A measure of the angle between the centerlines of driving and

driven shafts, where those centerlines would intersect approximately halfway between the shaft

ends. Coupling catalogs will show the maximum angular misalignment tolerable in each coupling. A

coupling should not be operated with both angular and parallel misalignment at their maximum

values.

AXIAL: A projection or movement along the line of the axis of rotation. Example: Sliding the hub in

either direction may change the position of a coupling hub, on its shaft. Thus affecting its axial

position on the shaft.

AXIAL DISPLACEMENT: One type of misalignment that must be handled by the coupling. It is the

change in axial position of the shaft and part of the coupling in a direction parallel to the axial

centerline. Can be caused by thermal growth or a floating rotor. Some couplings limit this

displacement and are called limited end float couplings.

AXIAL FORCES: The driver or driven equipment can generate axial forces (thrust) in which case

the coupling will pass those forces to the next available bearing with thrust capability. Because of

the inherent construction of some couplings, forces may be generated in the axial direction when

operating at high speeds or under misalignment. Such forces can place additional loads on the

support bearings.

AXIAL FREEDOM: This characteristic allows for variation in coupling position on the shaft at time

of installation.

BACKLASH: The amount of free movement between two rotating, mating parts. If one half of a

coupling is held rigid and the other half can be rotated a slight amount (with very little force), you

have some amount of backlash. The freedom of movement, or looseness, is the backlash and may

be expressed in degrees. Backlash is not the same as torsional stiffness.

BORE: The central hole that becomes the mounting surface for the coupling on the shaft. Close

tolerances are required. Bores/shafts are not always round, although that is the most common

shape. Other bore types can include hex, square, d-shaped, tapered, and spline. A spline bore is

one with a series of parallel keyways formed internally in the hub and matching corresponding

grooves cut in the shaft. Spline bores and shafts most commonly conform to Society of Automotive

Engineers (SAE) standards.

DAMPING: Some couplings greatly reduce the amount of vibration transmitted between driver and

driven shafts because of the damping capacity of an elastomer in the coupling. It is a hysteresis

effect that will generate heat. The coupling must dissipate this heat or risk losing its strength by

melting down. The stiffness of the elastomer affects the rate at which vibration is damped. All-

metal couplings, for the most part have poor damping capacity.

DISTANCE BETWEEN SHAFTS: The distance between the faces (or ends) of driving and driven

shafts, usually expressed as the "BE" (between ends) dimension or "BSE" (between shaft ends)

dimension.

FACTORS OF SAFETY: The coupling designer applies these factors to compensate for unknown

elements of the product design. The factors can compensate for temperature, material variations,

fatigue strength, dimensional variations, tolerances, and potential stress risers to name a few.

FAIL-SAFE: A fail-safe coupling is one that will continue to operate for a period of time after the

torque-transmitting element has failed. This is characteristic of couplings in which some portion of

both halves operate in the same plane, allowing direct contact between those portions. An example

of this is the jaw coupling, in which driving jaw faces push the driven jaw faces through an

elastomer in compression between them; if the elastomer breaks away, the driving faces simply

advance to push the driven faces directly.

FINITE LIFE VS. INFINITE LIFE IN COUPLINGS:

All couplings fall into one of these two categories:

1.). 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 & bush types. These types 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.

2). Infinite-life couplings (a name given to "non-wear" couplings) transmit torque and compensate

for misalignment by the distorting of flexing elements. The distortion results in fatigue stresses

rather than wear, and the couplings are designed and rated to operate within the fatigue

capabilities of the coupling material. "Infinite life" couplings do not necessarily last forever. This

group includes tire, disc, diaphragm, some donut types, wrapped-spring, flex-link, and most

motion-control types. "Infinite life" couplings remain infinite only as long as the load, including

those caused by misalignment, is kept within the coupling's design capabilities. An overload will fail

an infinite-life coupling (but may only reduce the life of a finite-life coupling). Infinite-life designs

are most often used on maintenance-free systems where maximum torque requirements -

including transient, cyclic and start-up torque – are known.

HORSEPOWER: The unit of power used in the U.S. engineering system. It is the time rate of doing

work. For power transmission it is the torque applied and rotational distance per unit of time.

Applied torque causes a shaft and its connected components to rotate at a certain RPM

(revolutions per minute). Horsepower (HP) is converted to torque as follows:

T = the torque in inch-pounds

Where T = BHP x 63025/RPM

BHP = the motor or other horsepower

RPM = the operating speed in revolutions per minute

63025 = a constant used for inch-pounds; use 5252 for foot-pounds, and 7121 for Newton-meters

The metric system uses kilowatts (kW) for driver ratings. Converting kW to torque:

Where T = BHP x 84452/RPM

T = the torque in inch pounds

kW = the motor or other kilowatts

RPM = the operating speed in revolutions per minute

84518 = a constant used when torque is in inch-pounds. Use 7043 for foot-pounds, and 9550 for

Newton-meters 

KEYWAY: A rectangular opening formed by matching rectangular slots cut axially (lengthwise)

along both the coupling bore and shaft. A square or rectangular metal key is then inserted into the

opening to lock the coupling and shaft in position. Torque is transmitted from shaft to coupling

through the keyway and key.

LENGTH THROUGH BORE: The effective length of the bore in the hub, or that portion of the

length that is useable and may be attached to the shaft.

OUTSIDE DIAMETER: The largest effective diameter of the coupling.

OVERALL LENGTH: The largest effective length of the complete coupling assembly.

PARALLEL MISALIGNMENT: A measure of the offset distance between the centerlines of driving

and driven shafts. Coupling catalogs will show the maximum parallel misalignment tolerable in

each coupling. A coupling should not be operated with both parallel and angular misalignment at

their maximum values.

RADIAL: Any projection outward from the center of a shaft or cylindrically shaped object, or any

motion along that line. The centerline of the projection or motion normally passes through the axial

centerline of the object.

REACTIONARY LOADS: When two shafts are offset (parallel misalignment), the coupling's radial

stiffness will cause a broadside force to be exerted on the shafts. This is called a "reactionary load",

as it causes the shafts to bend slightly in reaction to the broadside force. It may also be called a

"restoring moment", as a force produced by the coupling in an effort to restore, or correct, the

parallel misalignment.

RESTORING MOMENT: see REACTIONARY LOADS

SERVICE FACTORS: Multipliers that are assigned to common applications to compensate for their

typical load characteristics. These are used for the purpose of guiding coupling size selection to a

torque rating that will allow for unforeseen demands those characteristics might make on the

coupling. Such characteristics can include peak torque, start-up torque, transients or cyclic torque,

or any other empirical factor.

Among couplings that have no wear parts (see Finite/Infinite life), service factors are intended to

prevent premature failure due to overload damage. Among couplings that use wear parts to

transmit torque, service factors are intended to prevent premature failure of those parts due to

accelerated wear or degradation.

Caution: Resist the temptation to specify in excess of the published service factors. An oversized

coupling will not perform better or last longer, but will be unnecessarily expensive and force the

system to waste energy. Always base coupling size and service factor on the actual torque

requirements at the point of installation within the drive system.

SET SCREW: A headless screw, with hexagon shaped socket, used over a keyway to keep the key

stock in place and prevent the coupling from moving axially along the shaft. It can also be used for

torque transmission on low torque applications

STIFFNESS

STATIC TORSIONAL STIFFNESS: A resistance to twisting action (rotational displacement)

between driving and driven halves of the coupling. (The opposite - low resistance to twist - is

termed "torsional softness") Stiffness is expressed in lb.-inch/radian and measures the amount of

angular displacement about the coupling's axis of rotation at its static torque rating. Even

seemingly stiff all-metal couplings can have some degree of torsional twist.

TORSIONAL SOFTNESS: Torsional soft or hard is determined by dividing the dynamic torsional

stiffness by the nominal coupling torque rating. Values greater than 30 are hard (very stiff). Values

between 10 and 30 are torsionally flexible. Values less than 10 are considered very soft.

DYNAMIC TORSIONAL STIFFNESS: It is the relationship of the torque to the torsional angle

under the load of actual operation. The dynamic stiffness will be greater than the static. The

dynamic torsional stiffness can be linear, a constant value, or non-linear, an increasing value.

TOLERANCES: The amount of variation permitted on dimensions or surfaces of machined parts. It

is equal to the difference between maximum and minimum limits of any specified dimensions

TORQUE: In rotary motion it is the force multiplied by the radius, to the axis of rotation, at which

the force is applied. Force (F) multiplied by radius (r) = F * r = Torque. In English units (F) is in

pounds and (r) is in inches, expressed as in.-lbs. In metrics (F) is in Newtons and (r) is in meters,

expressed as Newton-meters (Nm).

TORSIONAL VIBRATION: The periodic variation in torque of a rotating system. Some causes of

torsional variation are the geometry of the rotating parts of internal combustion engines, cyclic and

irregular torque demands of the driven equipment, and variations in the output of certain types of

electric motors at startup.

C. Coupling Evaluation Factors 

These are attributes that affect the type of coupling best suited for an application. This is a long list

of evaluation factors. For any one application there may be only three or four attributes which are

extremely important. In fact it would be difficult to satisfy more than a half dozen attributes with

any one coupling. It is important to narrow the requirements for an application down to only the

most critical attributes that come into play.

In the next chapter we summarize the major coupling types discussed in the materials and provide

some ratings of each coupling type against these factors.

Adaptability of Design - Some couplings are available in a variety of configurations (e.g. drop-out

spacers, flywheel mounts, vertical applications, special lengths, brake drums). These alternatives

can be important to users who want to standardize on a particular type of coupling design, but

need to adapt it to suit different application requirements.

Alignment Capabilities - Different couplings have different limitations as to the amount of angular

misalignment, parallel misalignment or axial displacement each can accommodate. First,

determine the amount of misalignment that can reasonably be expected between the two pieces of

equipment to be coupled and let that guide or influence coupling selection.

Axial Freedom - Indicates how much movement can be accommodated by the coupling along the

axis of the two shafts, without compromising the coupling's ability to operate at rated torque and

without imposing reactionary loads on the bearings. This is important in two situations. The first is

when the BE dimension is very small and coupling hubs need to be installed further back from the

shaft ends. The other is when axial float in the shafts is characteristic of system operation. This can

include requirements for slider-type couplings or limited end float couplings.

Backlash - Also defined in the basic terminology section. Backlash is usually not desired in

applications where precise positioning of the shafts is important.

Chemical Resistance - The ability of the coupling components to withstand chemicals in the

environment around it, either mists, baths, dusts, etc.

Damping Capacity - The ability of the coupling to reduce the torsional vibrations transmitted from

one shaft to the other.

Ease of Installation - Some couplings are more complex and take more time to properly install and

align. This might be a concern if large numbers of couplings are to be installed or if they will need

to be replaced or moved frequently.

Fail Safe or Fusible Link - Fail-safe can be important in any application where unexpected stopping

of the driven equipment might jeopardize safety, incur high expense in downtime or scrapping of

material in process. If the equipment can be operated for a while longer, until a more opportune

time for maintenance can be scheduled, fail-safe is extremely valuable. The flip side of this is the

application where the user actually wants the coupling to disengage the drive if the element should

fail. This is sometimes referred to as a "fusible link" function being performed by the coupling.

There are some drives where the possibilities of severe torque or system overloads are high. In

order to protect the driver/driven equipment, a fusible link coupling may be preferred.

Field Repairable - Means that the key components are serviceable on-site so that the entire

coupling does not have to be replaced.

High Speed Capacity - Usually refers to speeds over 3000 RPM. If the coupling fits the application

but its standard off-the-shelf model is not rated for the RPM required, determine whether the

coupling can be economically changed to bring it up to the necessary speed. Sometimes it's a

balance issue and sometimes it's a strength issue due to centrifugal force.

Maintenance Required - Consider not only the frequency of maintenance that a coupling may

require, but also how long it may take to do the work. For instance, lubricated couplings will require

periodic checks of the seals and lubricant. And when the time comes to replace any components

and/or the grease, you usually have to put in new seals.

Number of Component Parts - The more parts a coupling has, the more complex it is, and the more

potential it has for problems. This often means it will take more time to install or disassemble for

repairs or maintenance, will require more spare parts to stock, and will be more costly to balance.

Reactionary Loads Due to Axial Forces - Some coupling designs inherently generate axial forces

during normal operation. Make sure shafts and bearings will be able to withstand the reactionary

loads that these forces will impose.

Reactionary Loads Due to Misalignment - A coupling's ability to accommodate misalignment is

evaluated in the context of the reactionary loads that will result. When misaligned, sometimes

even within their rated levels, each coupling has general propensities for sending reactionary loads

(whether axial or radial) through the system. If shafts are small, or not well supported, or bearings

are not substantial enough, these reactionary loads can cause problems.

Reciprocating Drivers and Loads - Due to torsional pulses generated by reciprocating engines

(most notably diesels) as well as certain kinds of pumps and compressors, coupling selection is

generally limited to a few elastomeric types capable of damping the pulses and providing

reasonable service life.

Temperature Sensitivity - This relates to the highest and/or lowest temperatures within which the

coupling materials can operate and provide normal service life.

Torque Capacity to Diameter (Power Intensity) - Couplings with equivalent torque-transmitting

capacity can vary in diameter. Size alternatives within the same torque range may become

important in applications where space is limited or if weight/inertia is a factor.

Torque Overload Capacity - Some couplings have the capacity to deal with brief torque overloads

many times the running torque, others will fail at only a few times the nominal rating. If you expect

to see high startup torque for instance and the drive starts and stops many times each day, you

would probably want to have a coupling which has good capacities in this area.

Torsional Stiffness - Defined in the basic terminology section, this is an attribute that is neither

good or bad, it just depends on the application and what is needed. You just need to be careful to

select a coupling type that has the proper level of torsional stiffness, in balance with the other

performance features it provides.

II. First Steps in Coupling Selection

Selecting the right coupling is a complex task because operating conditions can vary widely among

applications. Primary factors that will affect the type and size of coupling used for an application

include, but are not limited to: horsepower, torque, speed (RPM), shaft sizes, environment

conditions, type of prime mover, load characteristics of the driven equipment, space limitations

and maintenance and installation requirements. Secondary but possible essential factors can

include starts/stops and reversing requirements, shaft fits, probable misalignment conditions, axial

movement, balancing requirements or conditions peculiar to certain industries.

Because all couplings have a broad band of speed, torque, and shaft size capabilities, those criteria

are not the best place to start. First, determine what attributes beyond those basic criteria will be

required for your application. If none stand out then simply choose the lowest cost that fits those

basics. Almost always, though, there will be other considerations that will narrow your alternatives

down to certain types of couplings.

As we review those other considerations that guide coupling selection, we will omit rigid types and

focus on flexible couplings.

A. Types of Flexible Couplings

Many types of flexible couplings exist because they all serve different purposes. All types, however,

fall into one of two broad categories, Elastomeric and Metallic. The full range of coupling types in

both categories, and the special functions of each, will be discussed thoroughly in later chapters.

The key advantages and limitations of both categories are briefly contrasted here to demonstrate

how they can influence coupling selection.

1. Elastomeric

Couplings in this category include all designs that use a non-metallic element within the coupling,

through which the power is transmitted. The element is to some degree resilient (rubber or plastic).

Elastomeric couplings can be further classified as types with elastomers in compression or shear.

Some may have an elastomer that is in combined compression and shear, or even in tension, but

for simplification they are classified as compression or shear, depending on which is the principle

load on the elastomer. Compression types include jaw, donut, and pin & bushing, while shear types

include tire, sleeve, and molded elements.

There are two basic failure modes for elastomeric couplings. They can break down due to fatigue

from cyclic loading when hysteresis (internal heat buildup in the elastomer) exceeds its limits. That

can occur from either misalignment or torque beyond its capacity. They also can break down from

environmental factors such as high ambient temperatures, ultraviolet light or chemical

contamination. Also keep in mind that all elastomers have a limited shelf life and would require

replacement at some point even if these failure conditions were not present.

Advantages of Elastomeric Type Couplings

• Torsionally soft

• No lubrication or maintenance

• Good vibration damping and shock absorbing qualities

• Field replaceable elastomers

• Usually less expensive than metallic couplings that have the same bore capacity

• Lower reactionary loads on bearings

• More misalignment allowable than most metallic types

Limitations of Elastomeric Type Couplings

• Sensitive to chemicals and high temperatures

• Usually not torsionally stiff enough for positive displacement

• Larger in outside diameter than metallic coupling with same torque capacity (i.e. lower power

density)

• Difficult to balance as an assembly

• Some types do not have good overload torque capacity

2. Metallic

This type has no elastomeric element to transmit the torque. 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 generally are more expensive, need no

lubrication and little maintenance, but their primary cause of failure is fatigue, so they can fail

quickly in a short cycle fatigue if overloaded. If kept within their load ratings, they can be very

long-lived, perhaps outlasting their connected equipment.

Advantages of Metallic Type Couplings

• Torsionally stiff

• Good high temperature capability

• Good chemical resistance with proper materials selection

• High torque in a small package (i.e. high power density)

• High speed and large shaft capability

• Available in stainless steel

• Zero backlash in many types

• Relatively low cost per unit of torque transmitted

Limitations of Metallic Type Couplings

• Fatigue or wear plays a major role in failure

• May need lubrication

• Often many parts to assemble

• Most need very careful alignment

• Usually cannot damp vibration or absorb shock.

• High electrical conductivity, unless modified with insulators

B. Application Considerations

Sometimes selection of coupling type is guided by application, falling into one of five categories;

General-Purpose Industrial, Specific-Purpose Industrial, High-Speed, Motion Control and Torsional.

In each of these application categories there would be elastomeric, metallic membrane flexing, and

mechanical flexing types.

Once the coupling type is selected, there may be variations to consider within that type. For

example, gear couplings offer a wide variety of configurations to combine coupling functions with

other power train requirements, such as shear pin protection or braking. It is always a good idea to

understand as much as possible about the two pieces of equipment to be connected. Let the driven

equipment and the driver dictate the needs of the coupling. For example, is there a shock load or a

cyclic requirement that may lead to an elastomeric coupling? If low speed and high torque are

involved, that means a gear coupling is likely best suited. High-speed machinery will lead to a disc

or diaphragm coupling. Diesel drivers need the benefits of torsional couplings for best results. If the

equipment is susceptible to peaks or transients, the application may want high service factor or a

detailed analysis of the coupling torque capabilities. That brings us to the list of requirements that

will impact the coupling selection.

The charts below will help provide the path among all the couplings for most types of rotating

equipment. The charts are organized into three sections. The first is a list of "Information Required"

for the best possible selection of a coupling. It reflects the selection process used by the OEM

equipment designer, the engineer/contractor, the coupling specifier, or the trouble-shooter. For

other situations, short cuts are sometimes taken towards the conservative side. The second is a

chart of "Coupling Evaluation Characteristics" such as torque, bore and misalignment. The third is

the chart showing "Coupling Functional Capabilities”. They are the attributes of the various

couplings that go beyond the numerical information.

 C. Coupling Evaluation Charts

Information Required

1. Horsepower

2. Operating speed

3. Hub to shaft connection

4. Torque

5. Angular misalignment

6. Offset misalignment

7. Axial travel

8. Ambient temperature

9. Potential excitation or critical frequencies (Torsional, Axial, Lateral)

10. Space limitations

11. Limitation on coupling generated forces (Axial, Moments, Unbalance)

12. Any other unusual condition or requirements or coupling characteristics.

The first seven items of the list above will allow a coupling selection if a service factor is used. The

risk of relying on service factors is the possibility of ending up with an oversized coupling or one

that is missing an essential feature. All the remaining information, where applicable, allows the

coupling to be fine-tuned for the application.

Some types of couplings designed to do a specific job will have a further list of needed information.

For example, a slider coupling has to have the sliding distance and the minimum and maximum

BSE dimension.

Note: Information supplied should include all operating or characteristic values of connected

equipment for minimum, normal, steady-state, transient, and peak levels, plus the frequency of

their occurrence.

Information Required for Cylindrical Bores

1. Size of bore including tolerance or size of shaft and amount of clearance or interference required

2. Length

3. Taper shaft (Amount of taper, Position and size of o-ring grooves if required, Size and location of

oil distribution grooves, Max. pressure available for mounting, Amount of hub draw-up required,

Hub OD requirements, Torque capacity required)

4. Minimum strength of hub material or its hardness

5. If keyways in shaft (How many, Size and tolerance, Radius required in keyway, Location

tolerance of keyway respective to bore and other keyways)

Types of Interface Information Required for Bolted Joints

1. Diameter of bolt circle and true location

2. Number and size of bolt holes

3. Size, grade and types of bolts required

4. Thickness of web and flanges

5. Pilot dimensions

6. Other

Once past the charts that follow, one can go directly to the manufacturers catalog, or can read on

to learn more about specific couplings and the other important coupling issues.

Chart 1: Coupling Evaluation Factors

Chart 2: Functional Capability Chart

III. Popular Elastomeric Coupling Types

General Elastomeric Capabilities and Types

Elastomeric flexible couplings transmit torque between the two shafts by means of an elastomeric material

(rubber, urethane, etc.) positioned between the driving and driven hubs.

The resiliency of the elastomeric material gives these couplings varying degrees of torsional softness not

available in all-metal couplings, and generally greater misalignment capability than all-metal couplings. It also

allows a single flex plane to accommodate both angular and parallel misalignment. Couplings made as metal

flexing element or metal sliding element couplings require two flex planes to achieve parallel misalignment.

Power intensity (torque-carrying capacity vs. coupling size) of elastomeric couplings is lower than that of all-metal

couplings. With no (or little) friction wear between components, however, elastomeric couplings are considered

low maintenance, although elastomer breakdown in some coupling configurations is a maintenance issue.

Elastomeric couplings are quieter than some all-metal types. The softness of the elastomer cushions the

vibration and cyclic torque noises that result from backlash. Noise reduction can be an advantage in certain

applications, such as HVAC systems.

Because the elastomeric element handles misalignment by distorting, that action produces reactionary loads on

the adjacent shaft bearings. The reactionary loads vary in inverse proportion to the softness of the elastomeric

element. In all cases, greater misalignment will mean higher reactionary loads. Combined angular, parallel

(radial) and axial misalignments will result in the greatest reactionary load. Speed is a problem for elastomeric

couplings. The deflection of an elastomeric coupling is large for the load applied. Large centrifugal forces may

cause the element to protrude out of the coupling and hit the coupling guard.

Temperature is a restriction for elastomeric couplings. The material loses its strength as the temperature rises.

Eventually the strength reduces to zero. Temperature limits vary by type of elastomer, but generally 200 to 250

°F (110 °C) is the top end. Some elastomeric couplings may be used to dampen torsional vibration energy.

Hystersis, a characteristic exhibited by rubber with binders, allows the elastomeric material to absorb dynamic

energy. The energy is in turn lost in heat generation. If the material is able to radiate or otherwise conduct the

heat to a sink, damping will occur without damage to the coupling elastomer. If the heat builds up in the

elastomeric element it will fail or melt down. Elastomeric couplings of both the compression type and shear type

are used to control torsional vibration by damping the torsional vibration energy. The amount of hystersis is a

function of the elastomeric material as well as the stress level.

Damping of torsional energy in a power transmission system can also be accomplished by means other than the

flexible coupling. Frictional dampers, viscous dampers and torque converters are all used. The characteristic of

damping exhibited by these couplings is different from torsional tuning of a system. Torsional tuning uses the

dynamic torsional stiffness of the coupling to establish a low torsional critical speed.

Torsional stiffness of a coupling is a mechanical property of the coupling materials, modulus of elasticity, and the

geometry of the coupling element. Metal couplings usually depend on the spacer piece or floating shaft to lower

the resilience or torsional stiffness. Torsional stiffness is described as the torque necessary to deflect a coupling

in the circular direction. When dealing with power transmission couplings, it is usually measured in inch-pounds

per radian (Newton Meters per radian in metric). Rubber in shear and rubber in compression provide the lowest

torsional stiffness. Note that the geometric configuration of the coupling will determine the loading. The unit may

not be acting like a torsional spring just because we are applying a torque load. Other elastomers in the plastic

range are progressively stiffer. Coupling materials like urethane, and Zytel® make for stiff couplings. They have

little resilience, but carry more compressive load. Choice of materials in designing elastomeric couplings is a

balance between resilience and load carrying capability. Resilience is helpful for both cyclic loading and

misalignment capabilities.

Types of Elastomeric Couplings

Elastomeric couplings classify into three main types by the way their elastomeric element transmits torque - i.e.

the element is either "in compression", "in shear", or a combination of the two.

Compression Types. This type of elastomeric coupling is characterized by a design in which the driving and

driven hubs rotate in the same plane, with parts of the driving hub pushing parts of the driven hub through

elastomeric elements positioned as cushions between them, but not attached to either hub. As torque is

transmitted, the elastomeric elements are being compressed. Parallel offset misalignment is accepted via

compressive distortion of the elastomer material. Angular misalignment is accepted via sliding or distortion of the

elastomer material depending on the method of securing to the hubs.

Compression type couplings generally offer two advantages over shear types. First, because elastomers have

higher load capacity in compression than in shear, compression types can transmit higher torque and tolerate

greater overload. Second, they offer a greater degree of torsional stiffness, with some designs approaching the

positive-displacement stiffness of metallic couplings. However, greater torsional stiffness generally produces

higher reactionary shaft loads when the coupling is subject to parallel misalignment.

Shear Types This type of elastomeric coupling is characterized by a design in which all parts of driving and

driven hubs rotate in different planes, with the driving hub pulling the driven hub through an elastomeric element

attached to both hubs by various methods. These can include clamping, intermeshing teeth, or by bonding to

metallic brackets that are bolted to the hubs. As torque is transmitted, the elastomeric element absorbs some of

the torque force by being stretched through twisting. The design accepts misalignment through the deflection and

distortion of the elastomeric member and also through sliding, if the elastomeric member is attached to the hubs

through the use of intermeshing teeth.

Shear type couplings generally offer two advantages over compression types. First, they accommodate more

parallel and angular offset while inducing less reactionary load to the bearing. This makes them especially

appropriate where shafts may be relatively thin and susceptible to bending. Second, they offer a greater degree

of torsional softness, which in some cases provides greater protection against the destructive effects of torsional

vibration. Greater torsional softness generally produces lower reactionary shaft loads when the coupling is

subjected to misalignment.

The in-shear design also allows the coupling to act as a "fuse" to protect the driver and driven equipment from

torque spikes or system overloads which might cause damage elsewhere.

Combination Shear and Compression Type This type of elastomeric coupling transmits torque between hubs

through an elastomeric element in-shear, but transmits torque from hub to element (and back again) by

compression between hub teeth and intermeshing teeth formed into both ends of the element. Misalignment is

accommodated primarily by the sliding of the elastomer against the hub teeth (similar to a gear coupling).

1. Compression Loaded Designs

Jaw Couplings

A classic example of compression-type couplings, first patented in 1927, is the jaw coupling. It is still one of the

most widely used flexible couplings in the world and one of the lowest cost couplings available.

Since elastomeric technology was not what it is today, the spiders were originally made from materials such as

leather. Now a wide array of materials are available. Typical applications include pumps, gearboxes,

compressors, fans/blowers, mixers, conveyors, and generators, usually driven by an electric motor. Jaw

couplings usually are not recommended for engine-driven, frequent stop-start or reciprocating loads because

they are not designed to dampen torsional vibration. However, they might be able to serve such applications if

the proper service factors are used in sizing the coupling. Damping capability depends largely on the geometry,

type and amount of elastomer used.

Its design is simple, usually involving only three parts. Both driving and driven hubs have two to seven jaws

(thick, stubby protrusions) formed around their circumferences, pointing towards the opposing hub. When the

hubs are brought together, jaws from both hubs mesh loosely with each other. Gaps between them, and

sometimes the central inner space between the hubs, are filled with an elastomeric material, usually molded into

a single asterisk-shaped element called a "spider". The legs of the spider protrude radially to become the

cushions between the jaws. Some designs of Jaw couplings use blocks or tubes of rubber that are placed in

between the opposing jaw faces and must be held in place through the use of a retaining collar, or the hubs have

enclosed cavities into which the elastomer is placed.

In general, the greater the surface area (and volume) of the elastomer in compression, the higher the torque

rating of the coupling. Exploded view of Jaw coupling

Torque is transmitted from one shaft to the other through the compression of the elastomer between the driver

hub jaws and the driven hub jaws. Since the jaws between the two hubs rotate intermeshed in the same plane,

this design is called "fail-safe". If the elastomer should fail, the coupling will still transmit the torque, albeit quite

noisily given the metal-to-metal contact. This is still the preferred alternative for some applications, where the

equipment is critical to a production process and cannot be allowed to stop.

Some degree of permanent compressive set is normal as elastomeric elements age in service. This is a helpful

feature for Jaw couplings; when permanent set reduces the element's original thickness by 25% or more, it

provides a visual sign that the element should be replaced.

Another helpful feature unique to Jaw couplings is that compression is applied only to the spider legs or load

cushions forward of the driving jaws - trailing legs or cushions behind the driving jaws remain relaxed.

Accordingly, when compressive set reaches maximum in the driving cushions, the spider's trailing legs or

cushions can be advanced into the driving position. Thus, in most applications, jaw couplings carry a builtin set of

replacement elastomers, which can be used to reduce replacement costs. Note that couplings applied in

reversing drives or those with frequently varying torque usually relinquish this benefit.

Jaw coupling torque ratings are primarily limited by the elastomer material's compression strength, not the

jaw/hub strength. Thus, a jaw coupling can handle brief or infrequent torque spikes above the nominal rating far

better than the elastomer in-shear designs. It would take a torque of 6 or 7 times the nominal rating of rubber

elastomers to break off the hub jaws. If you change the spider from natural rubber to Hytrel® which has much

greater compression strength, the torque rating for the coupling is magnified 2 to 3 times. By contrast, an

elastomer that transmits torque through a shearing action cannot absorb torque any greater than 3 or 4 times its

nominal rating without tearing.

Other features of jaw couplings include; no metal-to-metal contact for quiet operation, resistance to

oil/grease/dirt/moisture in many tough environments, simple to install and align, and low maintenance

requirements. Many variations of jaw couplings are possible, ranging from flywheel designs, spacer couplings,

special hub materials as well as a variety of elastomeric materials to choose from. In addition, jaw couplings are

one of the lowest cost couplings.

There are some limitations to jaw couplings. Their angular and parallel misalignment is more limited than with in-

shear designs. When misaligned they introduce fairly significant reactionary loads on the shafts. Maximum bore

is limited by two factors: the inside diameter of the jaws and the length through bore of the hub. Generally, the

bore (shaft diameter) should be no greater than the length of shaft engagement in the hub. Maximum axial float

accommodated by jaw couplings is limited to about 10% of the axial thickness of the spider. Most designs have

backlash or free play between the fit of the elastomer/jaws and are not suited to motion control applications.

Temperature capacity is usually no greater than 250°F (121°C), that keeps them out of some applications.

Vertical applications are difficult since standard hubs are clearance fit bores with only one set screw, thus hubs

must be modified in order to grip the shaft tightly enough.

Jaw Coupling Types

Several different types of jaw couplings are available to serve different application requirements. Most of them fall

into two general categories: a "straight side" type, in which the jaw side faces are flat and straight; and a "curved

jaw" type, in which the jaw side faces have a cupped shape.

A. Straight-Side Type

Straight-side jaw couplings are available in sizes with bore capacities from 1/8" (4mm) up to 2-7/8" (73mm). This

coupling type is used for light to medium duty applications with a maximum torque capacity of 6,228 in-lbs. (704

Nm).

Small size hubs are made from sintered iron, while larger sizes are cast iron hubs. Neither sintered iron nor cast

iron can be welded to by normal methods.

•   Angular misalignment will vary from ½ to 1° maximum depending on the material used. (Materials discussed

later.)

•   The same goes for parallel misalignment capability, which will vary from .010" to .015" with different spider

materials.

Standard straight-side jaw couplings offer several alternatives in spider constructions in addition to the basic

asterisk-shaped solid spider or open-center spider spiders, both of which are held captive naturally within the

assembled coupling. (Open-center spiders simply allow greater axial freedom for installation on shafts with a

close BE dimension.) Alternatives include collar, ring-in-groove, block, and in-shear. Collar Types are those fitted

with elastomeric elements that are installed and removed externally. Such elements usually take the form of a

linear spider in which the legs are molded into a single strip of elastomeric material that is wrapped around the

assembled coupling so that the legs drop into the spaces between intermeshed jaws. These wrap-around spiders

require a circular collar around the coupling's circumference to prevent the elastomeric strip from being flung off

by centrifugal force. Typically, the collar is a stamped steel ring held in place by three retaining screws to one of

the hubs.

Ring-in-groove types, sometimes called "Snap-Wrap", are similar to collar types except that wrap-around spider

is held in place with a Spiralox retaining ring that snaps into a groove that is molded into the spider's perimeter.

This version is only available in NBR spider material and the maximum speed is 1750 RPM. Standard hubs are

used. The ring is removed easily with needle-nose pliers.

These features are ideal for those situations where the shaft ends must be positioned closely together, yet the

shaft diameters are greater than what can be accommodated in the open center type spider.

The compression block types serve heavy-duty applications that require shaft size and/or torque ratings beyond

the capability of standard Jaw couplings. Usually of larger diameters, these designs transmit torque through

independent blocks of elastomeric material, in cube, oval, or wedge shapes. Sometimes called load cushions or

elastomer cylinders, these blocks are individually inserted into the spaces between the assembled coupling's

intermeshed jaws, and held in place by a steel collar. This design offers the advantage of easily changing its

torsional stiffness by varying the hardness and design of the blocks. The compression block jaw coupling is

available with a maximum bore capacity of 12.0" (300 mm), and torque up to 1,000,000 in-lbs. (113,000 N m).

Common applications include compressors, large fans, blowers, mixers, and municipal or irrigation pumps.

In-shear spiders, the newest improvement in spider design, completely change the way the jaw coupling

functions. These spiders are axially twice as wide as standard spiders for straight sided jaw hubs, so instead of

allowing the jaws of both hubs to intermesh in the same plane, they push the hubs apart so the jaws rotate in

separate planes, and in axial alignment hub-to-hub. This arrangement causes the radially removable elastomer

to transmit torque through a combination of shear and compression method. The spider is held in place with a

floating stainless steel ring, which locks into special grooves in the OD of the spider.

As with the collar and snap-wrap designs, the jaw in-shear allows easy removal and replacement of the spider

without disturbing the hubs. There are no fasteners to worry about either since the retaining ring slides into

grooves in the spider. One available design fits standard straight-sided jaw coupling hubs on the market which

makes it an easy retrofit design. Another version uses special hubs with many shorter, stubbier jaws and a

special elastomer, but achieves the same concept in features/benefits.

The primary benefits are (1) simplified maintenance (2) non-failsafe operation (3) greater angular misalignment

capacity of 2°, and (4) greater torsional softness.

This coupling should only be used for electric motor driven applications, most commonly centrifugal pumps, fans,

mixers, gear boxes, and plastic extruding machines.

Special Hub Materials and Designs

Jaw coupling hubs are typically made of sintered iron or, for larger sizes, cast iron. Neither can be welded to by

normal methods. In some applications customers will desire to weld the connection of the hub to the shaft, or

weld another component such as a shaft collar, sprocket or pulley to the diameter of the hub. Special materials

such as 1018 steel, 303/316 stainless or 660/464 bronze are possible to meet those and other unique application

requirements. The torque and misalignment ratings do not change based upon the hub material. The elastomer

spider determines those ratings.

Light Hubs: This category of the standard jaw coupling uses hubs made from aluminum or other light metals. It

provides for a significantly lighter coupling if lower inertia is important. When an application calls for better

corrosion resistance than sintered or cast iron, but not the expense of stainless steel, aluminum is a good

alternative. Light material hubs use the same spiders as the standard straight sided jaw.

Special modifications such as clamped hubs, bushed hubs, extra long or shorter than standard hubs, and

pinholes are possible. The use of clamped hubs or bushings with couplings is common. Generally these are

advantageous when the application requires a firmer grip on the shaft than is provided with clearance (slip fit)

bores and one or two set screws. These include vertical drives, motion control, or equipment with high levels of

vibration and shock loads.

Elastomer options

When jaw couplings were invented, elastomeric technology was not what it is today, and spiders were originally

made from natural materials such as leather. A wide array of materials are now available, including several non-

rubber-based elastomers that offer light weight, chemical resistance with the ability to be molded into complex

shapes. They are also economical to manufacture and use. Generally, rubber-based spiders are more resilient

and better for cyclic loading and misalignment capabilities, while synthetic spiders make for torsionally stiffer

couplings that can carry a more compressive load.

1. NBR (Nitrile Butadiene Rubber) a.k.a. Buna-N -- is the standard and most economical material for jaw coupling

spiders. It offers the best combination of temperature and chemical resistance, misalignment, and damping

ability. Rubber has the best resiliency in bouncing back from deformations that occur in cyclic or heavy shock

loads. This is the only material suitable for reciprocating engine applications.

Most sizes of NBR spiders are 80A-shore hardness and are black in color. Temperature range is -40°F (-40°C) to

212°F (100°C). Also referred to as "SOX" by some manufacturers. This material will allow the Jaw coupling to

experience a torsional wind-up at full torque load of 4°-10°, depending on the coupling size.

Another attribute of natural rubber products used in compression is that they take a permanent "set" or loss of

volume after just a short time in operation. This does not become a performance problem until the spider

thickness is anything less than 75% of its original size, at which point it should be replaced. This limits their

selection in motion control/precision applications since increased free-play in the coupling results from the "set".

Shelf life of natural rubber elastomers is 5 years.

2. URETHANE has a 1.5 times greater torque capacity than NBR due to its greater compressive strength (either

40D or 55D shore hard ness is used) as well as better abrasion/wear characteristics. It holds up better to

environmental conditions such as ozone, ultravio let, and some oils-chemicals versus the NBR. It is limited to -

30°F (-34°C) to 160°F (71°C) temperatures however, and should not be used in heavy cyclic or start/stop

applications since the damping ability is limited. The in-shear spider is a slightly different type of urethane and is

rated for -30°F (-34°C) to 200°F (93°C). Urethane spiders typically are blue color and offer a shelf life of 5 years.

3. HYTREL® increases the torque capacity of the jaw coupling approximately 2½ times versus the NBR with its

higher compressiveload carrying ability. These spiders are a tan or cream color with a 55D shore hardness. This

material provides the best chemical resistance as well as a temperature range of -60°F (-51°C) to 250°F (121°C).

However, as with Urethane, it should not be used in appli cations where cyclic loads, frequent starts/stops, or

regular shocks and vibrations occur. The shelf life is 10 years. Angular misalign ment is only 1/2° versus the NBR

and Urethane that are both 1°.

4. BRONZE is not an elastomer, but is one of the options available for those high temperature requirements (up

to 450°F) which most other materials are not capable of. Most commonly, bronze is selected in salt water/marine

applications. It is only to be used for slow speeds, less than 250 RPM, since the coupling will prematurely wear

from metal-to-metal contact otherwise.

5. NYLON is a good electrical insulator, holds up well under heavy continuous loading, and may be substituted

where bronze is too noisy. The torque rating is the same as for Hytrel®.

6. VITON® is a synthetic rubber that has a temperature range of -65°F to 450°F with a durometer of 75-85A

scale. It provides the high temperature capability of bronze with excellent chemical resistance. The torque rating

is the same as for NBR and may be slightly derated depending on the application conditions.

7. ZYTEL® is a fiberglass reinforced compound with excellent resistance to most chemicals and corrosion. It is

three times more torsionally stiff than Hytrel® and can operate in temperatures ranging from -40°F (-40°C) to

300°F (149°C).

Curved Jaw Couplings

While the straight jaw coupling is known around the world, there is also another design that has wide acceptance,

primarily in Europe and Asia. It is generically referred to as the curved jaw coupling. This jaw coupling product is

available in sizes covering bores from 5/32" up through 5-11/16" (145mm) and torque from 35 in-lbs. up to

66,375 in-lbs. (7,500 Nm).

While the coupling still consists of two hubs and a spider in the center that is under compression, the main

difference is in the geometry of the jaws and the corresponding spider legs. The intermeshing faces of a radial

curvature, giving them a concave or cupped shape. This provides a built-in encapsulation of the spider legs by

the hubs. The corresponding spider legs are crowned, or curved both axially and radially to follow the jaw face

shape, making them similar to a gear tooth in geometry.

The jaw and spider curvature has two important benefits. First, by encapsulating the spider legs, it permits higher

speed ratings compared with similar size straight-sided jaw couplings. It also extends angular misalignment

capacity to 1.3° for some sizes.

Most curved jaw hubs have four jaws vs. three for similar sizes of straight-sided, with the jaws pushed farther out

toward the perimeter of the hub. This enables the spiders to have large open centers. The design characteristic's

combine to allow larger maximum bores in most cases and to accommodate close "BSE" dimensions.

The curved jaw design also results in some special limitations. Due to encapsulation, radially removable spiders

(wrap around, block) cannot be used. The damping capacity of the design is lessened under greater loads. The

overall length of the coupling is usually greater than the similar straight-sided jaw coupling. The type of sintered

iron commonly used in the smaller sizes is much denser, translating into heavier couplings. And finally, spacer

couplings can only be achieved by using extended hub lengths, this adds a lot of weight and still does not allow

for a true drop-out section.

Spider types

The standard material is urethane for all curved jaw spiders. There are simply three different shore hardness

which yield differing levels of torque capacity. Each of the shore hardness numbers are color-coded for easy

identification, blue for 80-shore, white/yellow for 92-shore, and red for the 98/95-shore. All of the spiders are an

Open Center Type (OCT). The urethane composition allows for a maximum temperature rating of 212°F (100°C)

versus the 160°F upper limit for the L-type urethane. Some manufacturers also offer Hytrel® as an alternate

material as well.

Also available for curvedjaw applications is the No Backlash (NBL) spider. This is simply a special, thicker spider

that can be used with the standard hubs to provide a snugger fit for those low backlash requirements. It only

provides a true zero backlash up to 10% of the rated torque of the spider. It is available in two-shore hardness

(92-yellow and 98-red). Some manufacturers also offer a special hub, often referred to as the "GS" style, for use

with the NBL spiders. The GS hubs are of similar geometry to the standard curved jaw hub except the jaws are

slightly oversized to make the intermeshing of the three components a true interference fit. This style can either

be pre-assembled at the factory or assembled by the user with the aid of a lubricant since the components are so

tightly fitted. It provides full zero backlash performance for motion control applications up to 10-25% of their rated

torque, depending on the size.

Special Considerations for Selection

Because this coupling was designed in Europe, it uses the DIN 740 rating methodology, which gives you a

Nominal (Tkn) as well as Maximum (Tkmax) rating. Nominal torque Tkn is the steady state design torque for the

coupling. Maximum torque Tkmax is a cyclic torque capability for 100,000 cycles or 50,000 reversing cycles.

In terms of the selection process, it means that the Service Factors are unique for the curved jaw coupling. There

are independent factors which must be multiplied by the nominal torque of the application to arrive at the design

torque. Only when the coupling/spider Tkn and Tkmax ratings are both greater than the respective nominal and

design torque (calculated for the application) do you have the proper size coupling. The urethane spiders, while

rated for a maximum temperature of 212°F, have a de-rating factor that must be applied to their misalignment

capability. This takes effect at any condition above 86°F.

Donut Shaped Elastomeric Couplings

This style of coupling was developed in 1970 for use with diesel engines. The donut shaped elastomeric coupling

consists of a rubber donut fastened with cap screws to hubs. The hubs provide the shaft connection. The

elastomer mounts in between the hubs to transmit the torque and allow misalignment. Metal inserts (either

aluminum or steel) are bonded into the elastomer and provide a durable material through which the fasteners

attach to the hubs. The elastomer donut is precompressed between the fasteners to make certain that the torque

is always transferred in a compression mode. The elastomer is stronger in compression than in tension. By

preloading the donut any tensile forces merely relieve the compression and do not put the unit into a tensile load-

carrying situation. Donuts can have a square, rectangular, octagonal or other cross-section design. They do not

have to be round.

Donut couplings can have one hub that is smaller than the other to fit inside the donut. It is called the cylindrical

hub. The donut is fastened to the inner or cylindrical hub by radial fasteners. The other hub is a flanged hub to

which the donut is attached by axial fasteners. The elastomer uses metal inserts that transfer torque by friction

between the metal inserts and the metal hubs then through the elastomer to the next set of fasteners attached to

the other hub. The torque path alternates from one leg of the donut to the next. The fasteners are tightened to

make a high friction joint and avoid loading the bolts in shear. Donut couplings that use the cylinder and flange

hub system have bore limits on the cylindrical hubs compared to other couplings of similar torque capabilities.

One way to eliminate the cylindrical hub limitations is to use a wraparound type of elastomer. The inserts are all

devised to use radial cap screws to fasten the elastomeric element to alternating hubs. This one does not have

the flywheel plate option or material options for the element.

Another design has spider shaped hubs with arms that are at the same diameter as the bolt circle within the

donut. Once again the attachment alternates from one hub to the other, but the fasteners are all axial. Torque is

carried by an elastomer in compression and is transferred to the hub via metal tabs inserted in the rubber.

Torque is also transmitted by the friction between the bolt sleeve inserts and the hubs. The torque path is from

hub to insert to elastomer to insert to hub. The elastomer carries the load in compression on alternate legs. This

design is not available with flywheel plates or stiff elastomer materials.

Donut type couplings can handle a load in either direction as the load shifts to alternate legs still in compression.

Even more importantly the donut can accommodate alternating loads and cyclic loads without backlash. There is

windup in elastomeric couplings. These couplings when constructed of rubber exhibit a quality of hystersis. That

quality enables the coupling to dampen the vibration energy that passes through the coupling.

Elastomers for Donut Type Couplings

The base elastomer is a natural rubber with binders. It is suitable to about 190°F temperature before it loses

strength. When the temperature increases the coupling must be derated. The formulations of this elastomer are

identified by the shore hardness. Each successively harder rubber carries more torque, but is torsionally less

resilient.

Alternate elastomers include Hytrel® and Zytel®. Each is considerably more stiff than rubber. The change in

materials will mean an increase in normal torque capability. The change in material may require the coupling

design to change in order to accommodate the fastening of the Donut coupling with all bolts axial elastomer to

the metal hub.

Pin & Bushing Type Couplings

Pin & Bushing couplings transmit torque through cylindrical or barrel shaped metal pins that are enclosed in

elastomeric bushings. The elastomeric bushing covers one half the pin while the other half has stepped

diameters with a threaded end. The shaft connections are flanged hubs drilled to hold the bushing or the

threaded end. It can be done with the bushings all in one flanged hub or they can be alternated from side to side.

The bushing is inserted into a cylindrical hole while the threaded and stepped end is inserted into a stepped hole

with counter bores on either side. A nut is attached to the threads to hold the bushing in place during operation.

The elastomers are compressed into the holes and may have a shape that permits easy installation. Elastomers

can be rubber, the original bushing type, Viton®, or urethane type materials. Hubs are cast iron, steel or stainless

steel. Pins are steel or stainless steel. The elastomer cushions shock loads and compensates for misalignment.

Pin and bushing couplings are inherently fail-safe with the pins continuing to transmit torque when the bushing is

worn. The bushing can both wear and fatigue from usage. Pin and bushing couplings are non-lubricated.

The pin and bushing coupling are high capacity vs. their size. The capacity or torque capability is directly related

to the bolt circle diameter, number of pins, and the type of elastomer. They are designed to make it easy to

replace the pins and bushings which are the wearing components.

2. Shear Loaded Designs

Shear-Type Donut (Sleeve)

The original patent on this design was issued in the late 1950's. The shear type donut coupling (sometimes called

a sleeve type) was marketed heavily to the pump industry, in particular the ANSI chemical process pump

segment. A strong following was built up which continues to this day. Much like a jaw coupling, the shear type

donut is simple in design. A standard coupling is composed of three components, 2 flanges and 1 sleeve. The

sleeve (a short, spool-shaped, tubular element) has serrations molded around the perimeter of each end, which

mate with corresponding serrations molded into both hub flanges. This puts the element in-shear between the

two flanges, so the torque is transmitted through the twisting of the elastomeric sleeve. There are several

features to this coupling which translate into tangible benefits to the user:

•   Because this design is double-engagement, it is radially very soft and produces very little reactionary load on

bearings and shafts when misaligned. However, misalignment will shorten sleeve life.

•   The torsionally soft design of an in-shear elastomer helps to dampen out most peak overloads and prevent

vibratory torque from going back to the driver.

•   The sleeve has a large open center, which allows close positioning of the shafts.

•   The torque overload capacity of this coupling is only 3 or 4 times the rated torque (the point at which the

sleeve will tear, round-off the teeth, or "pop out"), versus the 6 or 7 times for a jaw coupling. Thus it provides the

"fusible link" protection characteristic of in-shear couplings.

The shear type donut style coupling is best suited in the following applications:

•   Where system alignment may be hard to maintain over a period of time, and the coupling needs to tolerate the

drift.

•   Where the motor and pump are on a common base plate but there is no pump mounting bracket involved, i.e.

a "non-piloted" pump application.

•   Where shafts are closely coupled (i.e. minimal BE dimension).

•   Where shafts are relatively small for the torque loads, or the bearings are light duty.

In general, the shear type donut coupling will work well on electric motor driven applications with uniform loads

such as; centrifugal pumps, blowers and fans, screw compressors, some conveyors, line shafts, and vacuum

pumps. Care should be taken however, that shear type donut couplings are not used under the following

conditions:

•   Where loads have high-inertia, especially if they produce variable torque loads, or where overloads/spikes are

expected to be greater than 2X nominal ratings.

•   Where reciprocating engines, compressors or pumps are involved. Shear type donut couplings do not respond

well to torsional vibrations. •   Where the coupling will operate regularly at less than 25% of its rated torque. The

sleeve teeth will wear prematurely due to the rubbing action against the flange if too lightly loaded. This can be a

concern particularly with the Hytrel® sleeves since they have such high ratings.

There are five manufacturers of this design. All produce their product to be fully interchangeable. However,

serrations in sleeve ends and hub flanges must mate, so components from different manufacturers may not

always fit together properly. This is due to the tolerance that is built into each company's initial design criterion

(i.e. how tight or loose they want the fit between components to be), and the state of wear of the tooling that

produces the sleeves and flanges. Mixing of components from different manufacturers must be avoided if at all

possible.

Flange designs

A. J-type

A basic, economical flange, the J-type is available only in four smaller sizes 3 through 6, with smaller bore

models cast in zinc alloy and larger bore models in cast iron, all limited to the lower torque sleeve materials

(discussed later).

B. S-type

Provides a greater variety of sizes, from 5S to 16S, with all flanges made of cast iron. Characterized by extra

cast-in thickness projecting from the inner face of the hub, which allows greater through-the-bore shaft

engagement, S-type flanges allow larger bores than available with J-type flanges, and can be used with all

sleeve materials.

C. B-type

This flange is modified to accept an industry-standard bushing. Offered in sizes 6B through 16B. The use of a

bushing limits the bore capacity of the coupling, but provides a better grip on the shaft. It can also simplify the

stock room of many users, if they use bushings on other P.T. components. Due to the torque limits of the

bushing, Btype flanges cannot be used with higher torque sleeves.

D. T-type

Similar to the B-type for on industry standard bushing, the T-type is a standard flange modified to accept another

industry style of bushing. There are two ways to mount the bushing to the flange. The first way is from the

serration side (rear) or from the same side of the flange as the shaft is inserted initially (front). As with the B-

flanges, the T-type cannot be used with high torque sleeves due to the limits of the bushing ratings.

E. SC-type

Intended primarily for pump applications, these flanges are separable from their shaft-mounted hubs by removal

of four hex-head cap screws axially installed through each hub. This enables the flange-and-sleeve assembly to

drop out so routine pump maintenance can be performed without disturbing pump or motor mounting and

alignment. Various sizes for Spacer Flanges and Spacer Hubs can be mixed/matched to provide ANSI standard

separations of 3-1/2", 5" and 7", and dozens of other non-standard shaft separations as well, in coupling sizes 5

through 14. SC type flanges and hubs can also be used in combination with other flanged hubs to create a half-

spacer coupling. Any of the available sleeve materials can be used.

Elastomer (Sleeve) Types

A. Materials

EPDM is the standard material used. It is a rubber-like compound that allows the sleeve to twist as much as 15°

at full torque. It has the highest temperature rating (275°F/135°C) of the sleeves available. It provides good

resistance to most commonly found chemicals and is not affected by dirt or moisture. This sleeve is a dull black

color. Sleeves have angular misalignment capability of 1° and parallel misalignment ranging from .010" (size 3

coupling) to .062" (size 16). Neoprene® sleeves, which also can twist as much as 15° at full torque, offer better

chemical resistance than EPDM, especially to oil, but is rated only for a max. temperature of 200°F(93°C). The

color of the sleeve is black with a shiny finish and a green dot for easy identification. As with EPDM, Neoprene®

sleeves have angular misalignment of 1° with parallel misalignment ranging from .010" up to .062" .

HYTREL® is a polyester elastomer designed for high torque and excellent chemical resistance. It carries four

times the torque of the EPDM/Neoprene® materials but is limited to ¼° angular misalignment and parallel

misalignment from .010" (size 6) up to .035" for size 14 couplings. It only twists to about 7° at full torque. The

Hytrel® material is orange in color.

B. Sleeve Designs

One-Piece Solid sleeves are identified by material as JE (EPDM), JN (Neoprene®), and H (Hytrel®) types. They

are the least expensive of the rubber sleeves, available in sizes 3 to 10 for JE, and 3 to 8 for JN. For Hytrel®,

they are available in sizes 6 to 12.

One-Piece Split sleeves are identified by material as the JES (EPDM) and JNS (Neoprene®) types. They are

used for applications where the shafts are positioned closely together and the sleeve must be "peeled away" for

replacement. They are available in the same sizes as the JE and JN sleeves.

Two-Piece Split sleeve are made up of two completely separated halves. For the E (EPDM) and N (Neoprene®)

styles, a retaining ring is used to prevent the sleeve from bowing outward or being flung off under speed. The HS

(Hytrel®) is such a rigid material that the ring is not necessary. The E sleeve is available for size 5 - 16 couplings,

the N sleeve for sizes 5 - 14, HS for sizes 6 - 14. This design provides the greatest ease of installation and

replacement.

Clamped Elastomer in Shear Couplings

Corded tire types

This design came about in the late 1950's as a solution for dealing with transient torque peaks and shock loads in

diesel-driven pumps. Named for their resemblance to an auto tire, this design consists of two flanged hubs

equipped with clamping plates, which grip the coupling's hollow, ring-shaped element, by its inner rims.

Furthering the similarity, tire coupling elements usually are rubber derivative elastomers with layers of cord, such

as nylon, vulcanized into the tire shape. The coupling transmits torque through the friction of the clamp applied to

the inner rims of the tire and a shearing of the element. Slippage of the coupling may be expected to occur at

about four times the rated torque.

The two significant limitations to the corded tire type coupling are speed and space constraints. As speed

increases, the coupling exerts axial forces on the shafts due to the centrifugal forces working on the elastomer.

And the geometry of the tire itself makes for a large outside diameter for its torque capability. A design variation

includes an inverted tire coupling in which the tire element arcs inward toward the axis, thus overcoming the

centrifugal forces at speed. This affords 10-30% higher RPM service, depending on its size.

The corded tire coupling is torsionally soft and can dampen vibration. High radial softness accommodates

angular misalignment up to 4° and parallel offset up to 1/8". Rare among elastomeric couplings is its capability to

allow a certain amount of axial shaft movement. These properties give corded tire designs a wide variety of

applications including those driven by internal combustion engines. This coupling is offered in spacer designs as

well as with hubs which can accept bushings.

Bonded Urethane Tire

This design was first marketed in the 1970's and has found success primarily in the process pump industry

because of several features that the corded tire lacks. The design utilizes a urethane material that is bonded to

two half-circle metal rings (a.k.a. "shoes") which are then bolted to the two hubs. Torque is transmitted from the

hubs through the shoes/bonded joint and then the shear-plane of the split urethane tire.

The design offers advantages such as radial removal of the element halves, high angular misalignment capability

(4°), and shock load cushioning. In its standard close coupled configuration, it can span greater BSE lengths than

most in-compression couplings, and it also has the large opening in the center of the tire to allow complete

flexibility in positioning shaft ends. The outside diameter (OD) of this design is also smaller than the Corded Tire

type for similar shaft and torque capacities.

Spacer couplings are achieved by using the same shaft hubs and simply extending the lengths of the steel shoes

onto which the elastomer is bonded. Hubs can also be reversed in their mounting orientation to further add to the

BSE permutations possible. Bushings are also commonly used on this style of coupling. A heavy duty elastomer

option (25% more torque) is available, but it reduces the misalignment capacity by 50%.

It has proven to be ideal in applications such as pumps, screw compressors, blowers, mixers, crushers, and

general power transmission drives.

Limitations of the urethane tire type include the large number of fasteners required for installation and removal of

the elastomer, and the fatigue of the element and the bond between steel and elastomer under torsional

vibrations.

3. Combination Shear & Compression Loaded Designs

Jaw with Elastomer In-Shear

Another design of elastomeric jaw coupling completely changes the way the jaw coupling functions. Instead of

the jaws of the hubs interlocking, the use of an in-shear spider pushes the hubs apart and aligns the jaws of each

hub along the same axial plane. Thein-shear spider then is twice the axial width of a standard spider, and it is

loaded in shear rather than in compression. This spider provides certain features different from common jaw

couplings.

•   Radial removable spider

•   In-Shear design for non-failsafe operation

•   No metal-to-metal contact should the elastomer fail

•   No need for tools to install or replace the elastomer

•   Non-lubrication benefit of an elastomeric coupling

•   2° angular misalignment

A floating ring encases the outside of the spider and locks into special grooves on its OD. There are several

designs on the market, with only one manufacturer offering the benefit that this special in-shear spider is used

with regular jaw coupling hubs, same as for the in compression design. Urethane is the most common elastomer

material available. It has a combination of durability, chemical resistance, and torque/load carrying strengths.

This coupling should only be selected for electric motor driven applications. The most common ones include

centrifugal pumps, fans, mixers, gearboxes, and plastic extruding machines.

Torque ratings and service factors are unique for this version of a jaw coupling.

Gear with Elastomer

This is a gear coupling like the continuous sleeve gear coupling, except the sleeve is made from a slippery

elastomeric material. The advantage of this type of sleeve material is the no lubrication feature. They are limited

in torque, speed and size. The limits are imposed because the hub tooth to sleeve tooth friction eventually

exceeds the elastomers inherent lubrication capability.

The coupling consists of a molded nylon continuous sleeve with internal gear teeth that match gear teeth on the

periphery of a metal hub. The metal hubs are made from steel bar stock or powdered metal. The combination of

molded sleeves and powdered metal pressed hubs are a very economical coupling combination. The hubs are

held in the sleeve by spiral rings. The hubs are mounted on the shaft with the traditional clearance fit key, set

screw or a clamped type split hub. The nylon sleeve has high torsional stiffness and is resistant to chemical

attack. The hub tooth is crowned to obtain the misalignment capability. Backlash is designed to be at a minimum

in this style of coupling. Misalignment capability varies from 1° to as much as 5° depending on the manufacturer.

Nylon sleeve couplings are used for motor/generator sets, pump sets and other light to medium duty industrial

applications. Often they are used on the front power take-off of internal combustion engines because they are

small and lightweight. The coupling configuration permits vertical and blind assembly when needed. Speed

capabilities under light loads can reach 5000 RPM, misalignment to 5° and ambient temperature to 150 °F. The

bore capability is usually less than 2 inches for the popular sizes, however like all clearance fit couplings some

versions are available to 4 inches bore. For torque capabilities refer to the manufacturers catalogs as the torque

is tied to the speed rating as much as to the physical properties of the coupling.

4. Torsional Couplings

Several designs of couplings were developed to solve the problem of damping torsional vibration. The primary

source of those vibrations are diesel engines, but there could be other sources. The torsional vibration travels

through the coupling to the connected equipment. The vibrations can damage both the connected equipment and

the coupling itself. A discussion of torsional systems and vibrations are included in the Applications section of this

handbook.

Couplings

The primary torsional coupling uses a resilient elastomer as the flexing medium. All of the couplings described in

the elastomeric section of the handbook have been used on torsional service with varying degrees of success.

The elastomeric types described in the following section are the couplings with the best attributes for torsional

service.

The elastomer shape used in the coupling is very important for damping and damping is an important attribute for

the torsional coupling. The most successful shapes are radially loaded cylinders, toruses, and spheres. In

addition, the thickness may be much larger than found in conventional elastomeric couplings. Sharp corners are

usually avoided in the torsional designs to reduce stress concentrations.

There are also non-resilient elastomers used for torsional couplings. Non resilient couplings or stiff torsional

couplings are used for low inertia, diesel driven equipment.

Some metallic couplings were designed with the diesel in mind. One notable one is the grid coupling described in

its own section of this handbook.

The resiliency and the torsional softness of the coupling are used to judge the coupling's ability in torsional

systems. Torsionally soft couplings are those units that have a ratio of dynamic torsional stiffness to nominal

torque of less than 30.

Donut Shaped Elastomeric Couplings

The donut shaped elastomeric coupling consists of a rubber donut fastened with cap screws to hubs. The hubs

provide the shaft connection, the elastomer mounts in between the hubs to transmit the torque and allow

misalignment. Metal inserts (either aluminum or steel) are bonded into the elastomer and provide a durable

material through which the fasteners attach to the hubs. The elastomer donut is precompressed between the

fasteners to make certain that the torque is always transferred in a compression mode. The elastomer is stronger

in compression than in tension. By preloading the donut, any tensile forces merely relieves the compression and

does not put the unit into a tensile load-carrying situation. Donuts can have a square, rectangular, octagonal or

other cross-section design. They do not have to be round.

Donut couplings have one hub that is smaller than the other to fit inside the donut. It is called the cylindrical hub.

The donut is fastened to the inner or cylindrical hub by radial fasteners. The other hub is a flanged hub to which

the donut is attached by axial fasteners. The elastomer uses metal inserts that transfer torque by friction between

the metal inserts and the metal hubs then through the elastomer to the next set of fasteners attached to the other

hub. The torque path alternates from one leg of the donut to the next. The fasteners are tightened so make a

high friction joint to avoid loading the bolts in shear. Donut couplings that use the cylinder and flange hub system

have bore limits on the cylindrical hubs compared to other couplings of similar torque capabilities.

The donut style elastomeric coupling is primarily used on torsional damping and tuning systems associated with

Diesel drivers. In such a case a flywheel plate replaces the flanged hub. The flywheel plate is drilled to match

various SAE designated or DIN designated flywheel dimensions. The coupling is configured to dissipate heat that

is generated by hysteresis. It is also rated for a maximum torque, a nominal torque, and a vibratory torque. Each

of the values are different, the maximum torque is limited to a specific number of cycles.

Donut type couplings can handle a load in either direction as the load shifts to alternate legs still in compression.

The donut can accommodate alternating loads and cyclic loads without backlash. There is windup in elastomeric

couplings. These couplings when constructed of rubber exhibit a quality of hystersis. That quality enables the

coupling to dampen the vibration energy that passes through the coupling.

Elastomers for Donut Type couplings

The base elastomer is a natural rubber with binders. It is suitable to about 190°F temperature before it loses

strength. When the temperature increases the coupling must be derated. The formulations of this elastomer are

identified by the shore hardness. Each successively harder rubber carries more torque, but is torsionally less

resilient. The variations allow the application engineer to tune the system for critical speed as well as torsional

vibration damping. Hystersis, a characteristic exhibited by rubber with binders, allows the elastomeric material to

adsorb dynamic energy. The energy is in turn is lost in heat generation.

If the material is able to radiate or otherwise conduct the heat to a sink, damping will occur without damage to the

coupling elastomer. If the heat builds up in the elastomeric element it will fail or melt down.

Alternate elastomers include Hytrel® and Zytel®. Each is considerably more stiff than rubber. The change in

material may require the coupling design to change to accommodate the fastening of the elastomer to the metal

hub. The increase in stiffness changes the unit from torsionally soft to torsionally stiff, and as a result the tuned

critical moves from a value below operating speed to one above operating speed. The change in materials will

mean an increase in normal torque capability. Refer to the chapter on torsional applications for more information

on critical speeds and damping requirements.

Elastomer Block Compression Couplings

This type of coupling is similar to the jaw coupling in that torque is transferred from one hub to the other by

compressing captured rubber blocks. In this case the hubs consist of an external claw hub matched to an internal

pocket hub that contains the elastomer. There are several varieties of these couplings.

Variations include the shape of the elastomer blocks and the type of elastomer. The ideal shape for the

elastomer is a cylinder, loaded radially. Alternatives use rounded-off rectangular shapes. The coupling is used for

both shaft-to-shaft connections as well as shaft-to-flywheel connections. A popular application for this coupling is

the diesel driven generator. Another common application is the synchronous motor driven compressor. Both of

these applications are very high horsepower units. An example of a lower horsepower application is the electric

motor driven reciprocating compressor.

This style of torsional coupling is manufactured with OD's of 3 inches to several feet. Obviously the large size

carries a very high torque that is associated with the large generator sets and ship propulsion. The hubs can be a

casting of iron or bar stock or a forging of steel. Torsional stiffness ratio for these units is in the medium range,

which is consistent with the application requirements. A very soft unit would either not have the torque capability

or would have to be dimensionally too big to get the torque capability.

Bonded Elastomer in Shear Couplings

This coupling was developed in 1980 for diesel flywheel applications. There are two basic types and both use an

elastomer element in-shear. These couplings have a very low torsional stiffness ratio, in the range of 1.5 to 12.

The normal torque capability of these couplings range from 900 inch-pounds to more than 1,000,000 inch-pounds

depending on the size and type. They are used with the largest of diesel engines and small ones when extremely

low torsional stiffness is needed. The couplings can be configured with elements in series to reduce the torsional

stiffness even more, or in parallel to increase the torque capabilities. As with other rubber couplings there are

several elastomer variations that are identified by the shore hardness. The higher the "shore" number ithe stiffer

the torsionaly coupling is.

The first type of bonded elastomer in-shear coupling is a rubber disk bonded to an inner (or driven) metal ring.

The inner ring can be combined with various hub types for fastening to a shaft. The types include tapered OD

split hubs, bolted straight bore cylindrical hubs, and special hubs for connecting to U-Joint shafting. The outer

diameter of the rubber disk is an external toothed form that slides into a circular metal ring with internal teeth. The

circular metal ring is cast aluminum to keep the weight, and therefore the inertia, low. The outer ring OD is

configured to bolt to a diesel engine flywheel.

The rubber disk has a designed shape to ensure that equal stress occurs over most of its section, thus providing

a large torsional angle and avoiding high stress in these areas. Loading at the inner ring and outer teeth is

reduced below normally accepted levels by the design of those two areas. The load is carried in shear from its

periphery to its center. This style has a non-linear torsional stiffness. The element could have some backlash in

the tooth form at the periphery, although normallyit it is a tight fit. The tooth area becomes a wear point when the

coupling is misaligned. Coupling life therefore, is dependant on the wear as well as the torsional loading cycles.

The second type of bonder elastomer in-shear is a four-sided closed ring of elastomer with a special cross

sectional shape. The OD, the ID, and one side are flat and perpendicular to each other, the fourth side is tapered

from OD to ID in a conical shape. The torque load is carried from one side to the other via shear forces. Both the

OD and the thickness from side to side determine torque capability. The elastomer is bonded to metal plates on

each side. The plate on the flat side is configured to attach to a flywheel adapter or a shaft hub disk at the OD.

The plate on the other side matches the conical shape and is configured to bolt to coupling hubs or half couplings

at the ID. The ID may include plain bearings to carry minor radial and axial loads.

There is a wide variety of secondary couplings that are bolted to the side opposite the flywheel. They include

gear coupling halves, link couplings, and disc plates. The secondary couplings provide misalignment capabilities

not available from the primary torsional coupling. Cardan shaft adapters and clutches have also been attached to

the coupling.

The rubber element again has a designed shape to provide for equal stresses across the element. The element

has a linear torsional stiffness. There is no backlash in this style of element, however there is torsional windup.

This coupling is a non-wear configuration and coupling life is dependent on the torsional damping and maximum

load cycles.

Torsionally Stiff Couplings (Flywheel)

While torsional softness can be a benefit for elastomeric couplings, there are some applications that require stiff

elastomers. Most of the elastomeric coupling types have an alternative stiff elastomeric material. Jaw coupling,

donut shaped compression loaded, and unclamped donut in shear are sometimes supplied with Hytrel® or other

stiff elastomers such as Zytel® or urethane. The stiff elastomer is used for greater torque capability without going

to a larger size. Stiff elastomers have less resilience and may restrict the angular misalignment capability to

much lower values.

Many times the switch to a torsionally stiffer elastomer is to tune the torsional system to a higher natural

frequency. This is done on some diesel driven systems with light inertia loading. One example is a diesel driven

hydraulic pump for off-highway equipment. Torsionally stiff couplings for these applications are a significant

coupling need. The couplings are designed for attachment from a flywheel to a driven shaft. The couplings can

be of the compression type as described under the "Donut Shaped Elastomeric Coupling" section or can be a stiff

elastomeric disc loaded in shear.

The stiff elastomeric disk, loaded in shear, has a torque capability up to 21,240 inch pounds. The torsional

stiffness ratio is above 100. Normally this is a high volume molded disk to make an economical coupling for small

diesel production engines.

Shear loaded disks are molded of Zytel® or nylon with strengthening fibers. The disk is designed with boltholes

on the periphery to match a flywheel-drilling pattern. The ID is designed to mate with a coupling hub in a sliding

fit. The coupling hub can have typical gear coupling teeth with crowning or can have four to six crowned dogs.

The crowning accommodates a limited amount of angular misalignment while transferring torque from the

element to the hub. Hubs are made from steel bar stock or from powdered metal. The hub bore is usually a

spline to match a standard hydraulic pump shaft, but could be a straight bore with key.

The torsionally stiff coupling for flywheel applications is designed for blind assembly. Having the shaft hub slide

into the flywheel attached elastomer does this. This coupling type is often supplied with pump mounting plates

and flywheel enclosures.

Metallic Couplings

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

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 teeth.

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.

Coupling Configuration

Sleeve Alternatives

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 bolted

together.

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" coupling.

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.

Axial Displacement

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 act concurrently.

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 failed teeth.

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.

6. Sliders

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

Vertical Couplings

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.

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

Interchangeability

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.

Alignment

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.

Indexing Couplings

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

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.

Other Considerations

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.

Speed

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.

Misalignment

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

Introduction

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 system.

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 clutches.

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 installation

time.

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.

Cover Styles

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 spacer couplings.

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 hubs. 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.

Spacer Type

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.

Grid Spring

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.

Routine Maintenance

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 lubrication.

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 replacement.

Special Designs

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.

Chain Couplings

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 shafts.

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 wear.

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.

Non-Lubricated Coupling Types

Standard Couplings and Special Purpose Couplings

All metal couplings such as the gear, disc or diaphragm types are available in standard cataloged

units and special or made to order (MTO) units. The catalog standard units are designed and

manufactured to meet competitive market requirements. MTO couplings can be totally new

designs, but most often for economy, simply add modifications to standard units. Catalog data

usually provides a guide to available capabilities. In addition there are special or made to order

units for which the catalog data is a guide to available capabilities.

In the case of gear couplings there is an AGMA standard for mating dimensions from size 1½ to 9

that is followed by coupling suppliers all over the world. Disc couplings are not quite as standard as

the gear units, but market forces have driven the various manufacturers to similar designs for

catalog standard units. Diaphragm couplings have the least standardization of the three, and in

fact more often fit into the special purpose category.

Special purpose couplings are used to meet unusual OEM requirements, process industry

requirements, military requirements, or extreme horsepower/speed applications. The special

purpose units are found in the gear, disc, and diaphragm design of couplings. Each manufacturer

makes claims for their design under competing circumstances. The competing designs could be

two disc couplings, disc vs. diaphragm or gear vs. disc. There is not necessarily a complete overlap

of the designs so the coupling user or system designer must pick the coupling design that has the

most favorable attributes.

Metallic Flexible Element Couplings

This group of power transmission couplings, consisting of disc, diaphragm, bellows, link and spiral

wound springs, uses the flexibility of thin metal elements to handle misalignment requirements

while the strength of metal is used to transmit torque. Each type will be discussed later in this

chapter, but first we offer an overview to put them in perspective.

All five types are designated as infinite-life, low-maintenance and nonlubricated. Infinite life

indicates they are designed to operate without breaking or wearing out over time. This usually

means the flexing and load carrying methods do not cause wear and are within the endurance limit

of the flexing materials. These couplings require a sophisticated stress analysis to determine the

loads on the flexing element under varying combinations of load from torque, misalignment and

displacement, plus temperature and the loads associated with rotation over a wide range of

speeds. In addition, the method of attaching the flexing element to a shaft hub requires a close

analysis, as that junction could be the weakest point under some conditions.

The disc and the diaphragm coupling have enjoyed increased reliability from the advancements

made in metals technology in analyzing the stresses in flex elements. One analysis tool that is

most often used is the Finite Element Analysis (FEA). While this method has been available for

many years, the advent of cheap fast personal computers has made it much more practical as a

tool for the design engineer.

Three of these types are found in general purpose applications; they are the disc, the link and the

spiral wound spring. The disc and the diaphragm are used for high-energy applications and the

bellows and the disc are used for motion control applications. Although disc couplings have the

broadest popularity, users should be aware that each type of coupling has its limits. Sometimes the

limit is an economic one, where the coupling can be suitable but is economically at a big

disadvantage to another coupling type that works equally well.

This group of couplings has been around for many years, in the case of disc couplings, since 1914.

The advent of high-speed machinery made them more popular. They offer the solution to many of

the problems associated with high speed, such as balance and the need for lubrication. To make

sure the couplings were suitable for high speed machines it was necessary for the engineering

community to develop computerized stress analysis methods, and new materials to make the

couplings more practical and economical. Even with the advances of modern engineering, these

couplings remain more expensive than many of the alternative gear and elastomeric units. These

couplings are more sensitive to environmental conditions and rough handling than other types of

couplings. While they have a theoretical infinite life, their life can be shortened to a finite one if

conditions of the application change, or if they are not carefully accounted for at the time of

coupling selection.

With careful selection, the metallic flexible element coupling can be the best choice for high speed,

high power and high accuracy applications. In some cases, such as those that require no-backlash

or torsionally very stiff couplings, no other couplings are suitable.

Disc Couplings

The disc coupling became popular as high-speed machinery created demand for non-lubricated,

low maintenance, long life couplings. That machinery includes boiler feed pumps, gas and steam

turbines, compressors, high-speed test stands and marine propulsion systems. Now industrial disc

couplings are used by a wide range of rotating machinery applications.

With careful design and analysis the disc coupling can provide a near infinite life unit. It consists of

two flanged hubs and one or more disc packs and a center member. There is also a version that

uses two hubs and one disc pack. The disc pack is a flat ring with boltholes on a fixed pitch circle.

The disc pack is attached to flanges on either side via alternate bolts. The leg of the disc is loaded

in tension, and transmits torque in a tangential direction to the disc. Disc packs are often unitized

by the coupling manufacturer to make assembly easier.

Single disc pack units (one flex point) will accommodate angular misalignment and axial

displacement only. The design configuration that uses two disc packs separated by a center

member can accommodate parallel misalignment, axial displacement and angular misalignment.

The amount of radial misalignment depends on the combination of disc pack angular capability and

the distance between the two flexing points. Sometimes, a long floating shaft separates two single

disc pack couplings. These are used to span long gaps as well as provide radial misalignment.

The discs can be circular in shape, flatsided on the OD, or can be scalloped on the OD. As we go

from circular to flat to scalloped ODs, the cost of the disc is theoretically higher as the shape is

more complex to

form.

The disc performance improves with the same progression in shape. Circular discs act like a curved

beam, which means the stresses can be very high on the extreme fibers. The flat-sided disc

eliminates the curve beam problems and puts the leg in tension. It is more flexible than the circular

one since it is slimmer across the width of the leg. The scalloped disc is the most flexible as it is the

slimmest. In the cases of the flat- sided and scalloped disc, the metal retains an ample cross-

section at the area of the greatest bending stress, which is the area surrounding the bolthole. The

slimmest cross-section requires the least force to flex the disc pack when misaligned. That means

lower bending forces are reflected back as lower reactionary loads on the nearby bearings.

Disc couplings can have up to 12 bolts alternately fastening the disc pack from one side to the

other at the bolt circle. The motion control couplings and light industrial couplings have 4 bolts,

industrial couplings 6 or 8, with 10 and 12 reserved for extremely high torque applications. The

diameter of the bolt circle, thickness of discs, the number of discs, and the number of bolts are

indications of the coupling's torque capability.

The thickness of the discs and the thickness of the disc pack indicate the coupling's flexibility.

Short legs with many bolts reduce the lateral (angular) flexibility of the leg.

Thin discs chosen for flexibility can be stacked to an optimum height. A stack of thin discs provides

parallel paths for the torque transfer, offering more flexibility than the same thickness of a single

disc.

Stacked discs are clamped together using the disc attachment bolts and nuts plus bushings and

washers. The clamping is critical in the coupling design and manufacture. Force must be

transferred via friction from the shaft hub flange to the disc pack and from the disc pack to the

next flange. Otherwise, the bolt would be loaded in shear, and the cross-sectional area around the

hole could be overly stressed. It is normal to have a slight bending moment on the bolt from the

torque transmission. The washers serve to spread the clamping pressure over a specified area and

are beveled on the disc side to prevent stress risers when the disc is deflected. The bolts may be

"body fit" to the bolthole or may be bushed. Bushings and "body fits" serve to equalize the loads

and prevent high stress areas. A second type of bushing called an overload bushing can be

inserted between the nut and the washer, in the flange counterbore. It serves to prevent overloads

of torque from stretching the disc, or if the disc were to fail, it would prevent flailing. Bolts and nuts

must be torqued to a specified value to obtain the proper clamping force. It is accomplished by

holding the bolt while applying torque to the nut. The bolts are high strength, special pieces,

machined to be used on disc couplings. Standard bolts or commercial bolts should not be

substituted.

The allowable number and thickness of the individual discs in the pack is a function of how well the

clamping force can be transmitted through the pack. The center members of very thick disc packs

would not have sufficient clamping. Also important are the true positions of the boltholes in the

discs, the dimensions of the discs and the true position of the boltholes in the flanges. Those

details serve to make sure loads are equalized on each leg of the disc pack.

The loads on the disc pack of the coupling consist of two types. There are steady state loads

associated with the transmitted torque, the rotational

load (centrifugal force), thermal load, and the axial displacement. Then there are reversing loads

associated with the angular misalignment. The force that is bending the links when the coupling is

misaligned is a reversing fatigue loading on the flex link. 

Designers use a fatigue analysis to determine the stress levels of the disc elements. They can use

a Modified Goodman Diagram or Constant Life Curve to establish coupling ratings. The objective is

to limit the stresses so that the disc pack has a theoretical infinite life at the "catalog" rated values

of torque and misalignment. In the case of special purpose couplings, the rated values are

determined by the needs of the application and the job specifications. Misalignment and torque

capabilities are interrelated on the disc coupling.

The reactionary force generated by a disc coupling is equal and opposite to the force bending the

link when the coupling is misaligned or displaced. Forces for axial and radial misalignment are

predictable and available from the coupling manufacturer as a chart of load vs. deflection. Axial

forces are usually low and the coupling tends to be self-centering. Friction between the discs

creates critical damping and a non-linear stiffness prevents an axial natural frequency from being a

problem for this style of flexible element coupling. It has a built in limited end float capability.

The application engineer or user of disc couplings must account for the misalignment and the

torque when selecting the coupling. An unaccounted for extra amount of misalignment is

detrimental to coupling life.

The coupling is an infinite life unit as long as the discs do not receive an overload in excess of the

fatigue capability. The construction of the disc coupling, with thin metal disc stacked very closely

together makes it subject to varying types of corrosion. Fretting corrosion between discs is

common, but can be prevented by coating the discs or by using corrosion resistant metals. Another

problem is atmospheric corrosion such as salt in the air or chemicals in the operating environment.

The thin discs would corrode quickly in the stressed areas. Once the metal is reduced in cross-

section its life is reduced to a finite value.

The disc coupling is backlash free and torsionally very stiff. Because it is all metal and has no

sliding or rubbing parts, the coupling is very suitable for high-speed applications.

Because the disc coupling is often used on high-speed applications, it is designed and built to be

balanced. Unbalance is minimized through concentricity of the parts and by piloted fits. That would

include piloted fits of all the connected parts, and tight tolerances on individual parts.

Types of Disc Couplings

Disc couplings can be made as a close-coupled design, or built as spacer type couplings. Close-

coupled units that are designed to drop in the place of standard gear couplings have split

centerpieces for replacing the disc pack. Spacer couplings, the most popular configuration, are

either the marine type or reduced moment type. The marine type, with the flex element positioned

outside the shaft hub, is easier to disassemble, and the ID of the disc pack is not controlled by the

shaft hub diameter. Reduced moment couplings place the flex point and therefore flexing element

directly over the shaft hubs nearest to a support bearing. The moment arm is reduced, reducing

bearing loads and changing the lateral critical speed. (See Balancing Chapter for information on

lateral critical speeds.) Reduced moment couplings can increase the diameter of the flexing

element or limit the hub bore.

Disc couplings have almost as many configurations as are described in the gear coupling section of

this handbook. This often allows disc couplings to be used in place of gear couplings, especially on

high-speed applications where the higher cost of the disc coupling is justified by the reduction in

maintenance. Disc couplings are capable of carrying loads as great as the gear coupling while

using just a little more space. In fact, a variation of the disc coupling will match up to the flange

drilling of a rigid gear coupling. That variation meets AGMA 516 Flange Dimensions (future AGMA

9008-AXX) and can be built as a reduced moment or floating shaft type. It is often used in paper

mills or other markets where a disc coupling is used to replace a floating shaft gear coupling.

The basic industrial disc coupling consists of two hubs, two disc packs, and a center or spacer

piece. The disc packs are bolted to the spacer and then to the shaft hub. Torque capabilities range

from 1.0 inch-lb. in the 4-bolt miniature coupling sizes, to 2.0 million inch-lbs. or greater in the

large 8-or 10-bolt industrial units. The lowest torque units may have only one disc at each flex

point.

It is possible to build disc couplings as vertical units, electrically insulating units, shear pin units or

to include a floating (torsion) shaft. The spacer units can be built with lightweight composite fiber

spacers between two single flex element couplings, a design often-used on cooling tower

applications. Couplings with axial movement limitations are available, as are split spacer close-

coupled units. The industrial units can have flange adapters rather than hubs or hubs can be

adapted to use either internal or external clamping devices (shrink discs).

The petroleum industry is a major user of disc couplings because of their high-speed rotating

equipment. Many variations of the disc coupling are built to meet specific requirements of API 610

(Pumps) or API 671 (Unspared and Critical Equipment). One variation is the "drop out" type. It is a

spacer-type disc coupling that utilizes a prepackaged center section that drops into or out of place

and fastens to the shaft hubs.This factory-assembled center section includes two unitized disc

packs, a center section (spacer) and two "guard rings" all bolted together with the bolts properly

torqued.

The "drop-out" style of disc coupling is available with safety devices either built into the guard

rings or as separate pieces. The safety device keeps the disc and spacer from flying out if a failure

occurs. A shim pack is used to properly place the center section on high performance drop-out

special purpose disc couplings. The shim pack allows for a variation in distance between shaft ends

without prestretching the discs in an axial direction. Prestretching the discs is reserved for

couplings that have thermal movement when in operation. Oversize hubs are available for the drop

out type of disc couplings, but not available for the industrial type. The oversize hub allows the

coupling to have torque and bore capabilities that are comparable with each other.

All varieties of the disc coupling can be inspected for damage without disassembly as long as you

can see the discs. It is also possible to check the coupling with a strobe light while in operation.

Under a strobe light the discs can be seen flexing and moving and a damaged disc would show up.

Disc couplings should not be used when rugged environmental conditions exist or when axial

movement is required and must be carefully applied when transient loads are encountered. It is

more expensive than a gear coupling.

Diaphragm Couplings

The diaphragm coupling was developed in the mid-fifties as an infinite life, non-lubricated, low

maintenance design for aircraft applications. It was introduced into industrial applications in the

late sixties, initially to

meet the demands of high-speed high horsepower service in petrochemical processing.

In diaphragm couplings the flexing element is a metal plate that is loaded in shear by introducing

the torque at the OD and transferring it to the ID. The process is reversed at the opposite flex

point. Each flex point of the coupling can utilize one plate or a stack of plates operating in parallel.

This coupling is used for the extreme applications of torque and speed and in those applications

where reliability is paramount. Applications range from helicopter drives to high horsepower high-

speed gas turbine drivers for generators and compressors.

Diaphragm couplings will have the largest OD of the three; disc, gear or diaphragm. The large OD

reduces the forces at the periphery where the coupling is attached to the flange. Attachment is via

clamping bolts at the OD and welding or bolting or major diameter fit splines at the ID. Where low

moment versions are needed the diaphragm couplings can be welded at the OD of the flexible

element, machined integral at the ID and bolted at the removal joint.

There are probably more types and styles of flexing elements available in diaphragm couplings

than in any other style of coupling. Each style has its adherents and all work to some level of

satisfaction. The user or specifier of the diaphragm coupling must have confidence in their chosen

supplier of diaphragm couplings. This is more true of this style than the other types of couplings

because no common designs exist. The couplings are used in a more sophisticated application than

the gear or the disc.

Single diaphragms are contoured or shaped to have a constant stress across the diaphragm. The

stress combines shear loading for torque transmittal with a bending moment for angular

misalignment and stretching for axial displacement. As we found on disc couplings, ratings are

established by a stress analysis using a Modified Goodman Diagram or a Constant Life Curve and

augmented with sufficient factors of safety.

Diaphragm materials are high strength alloy steels perhaps in corrosion resistant versions. Single

element diaphragm couplings when made from corrosion resistant materials can be considered the

true infinite life device since fretting corrosion between a stack of fretting elements is eliminated.

Or, multiple flex elements can be assembled to have space between the plates to eliminate the

fretting corrosion that otherwise would reduce coupling life to a finite value.

Multiple flex element construction reduces the reactionary forces from misalignment. Although the

flex elements must have uniform cross-section in the flexing plane, they offer a variety of styles,

including convoluted (wavy), perforated, or spoke designs. The reason for the various shapes is to

allow flexing with the least amount of force while still transmitting the torque.

Diaphragm couplings have more flexibility than disc couplings, therefore reducing cyclic loads

associated with angular misalignment. This results in a stronger coupling with less susceptibility to

fatigue failure, greater reliability and higher torque capability in a general sense. 

Diaphragm couplings with one flex point have only angular and axial flexing capability. To attain

radial (parallel) misalignment capability, they must have two flex points. In this configuration, the

coupling can be described as a spacer unit. As with the gear and disc couplings, diaphragm

constructions can be "marine type", "reduced moment type" or a combination of the two. As on the

other couplings, the reduced moment coupling can restrict the hub diameter and the shaft

capability, or can make the coupling larger in overall diameter.

Comparing the flexing element couplings illustrates that the disc and the diaphragm are close in

capabilities. The disc unit can have a lower cost, but the diaphragm can be considered more

reliable and suitable for higher torque and speed.

Here are some comparative comments applicable to this group of couplings.

• The diaphragm coupling is considered to be stronger than the disc coupling and can be made in

larger diameters than the disc coupling. (Note diameter is an indicator of torque capability.)

• The diaphragm coupling is considered to have more axial displacement capability at lower forces

than the disc coupling.

• Single straight contoured diaphragm couplings eliminate the fretting corrosion associated with

multiple elements. Single wavy types have more axial capability.

• Diaphragm couplings disconnect upon failure, which can be beneficial to the connected

machinery. It becomes a fusible link.

• In contrast with gear couplings, both the disc and the diaphragm are non-lubricated, low

maintenance, infinite life couplings.

• They both produce lower reactionary loads than gear couplings when misaligned.

• They both accept greater misalignment at higher speeds than gear couplings.

• Both the disc and the diaphragm coupling are easier to balance than the gear coupling.

• The gear coupling has the lowest first cost and the least sensitivity to rough circumstances.

• The gear coupling has the most axial displacement capability.

• The gear coupling has the highest power intensity.

• Gear coupling life is determined by wear, and overload generally results in reduced wear life

rather than abrupt failure. This is subject to existing wear and the size of the overload.

• The disc coupling is usually smaller in diameter than the diaphragm coupling.

• The disc coupling is usually less expensive in first cost than the diaphragm coupling.

• The disc coupling has a built in overload capability. Disc failure, while it may be abrupt, will not

disconnect the coupling if the bolts are capable of carrying the load in shear. This feature often can

allow the coupling to continue operating through an orderly shut down.

Triple- Wound Spring Couplings

The spring type coupling uses a tension spring fitted with rounded caplike hubs on both ends to

facilitate shaft mounting. Normally the spring should be loaded so as to tighten the wrap of the

spring because it can carry more torque that way. In this coupling, the triple wound construction

has the middle coil wound in a direction opposite the inner and outer coils, so it can transmit

torque effectively in either direction.

Flexing the spring as a tension spring accommodates angular misalignment. However, each end of

the spring can flex independently, creating two flex planes that accommodate radial (parallel)

misalignment. The amount of parallel offset allowed is primarily a function of the coupling (spring)

length.

Generally these are small couplings with torque capabilities up to 1800 inch-pounds (213 Nm) in

the largest size. Bores range up to 1.5 inches (38 mm). Angular misalignment can reach 3° for the

triple wound spring. Speeds can run up to 30,000 RPM in the smaller sizes. They are capable of

some endplay.

The triple wound spring coupling has both backlash and windup. The backlash amounts to 1/3 of

the windup at full torque. Windup, depending on size, can range from .85° to 1.8° at maximum

torque. Because the windup and backlash are known values these couplings are sometimes used

for indexing, positioning or robotic applications.

Flange ends could be substituted for the round hubs or a combination of hub on one end and

flange on the other. A bolted-in combination of flange and hub on both ends can be used to make a

drop out version.

When a spring coupling operates in the misaligned system it is subject to both a nominal torque

load and a cyclic flexing load. Overload on infinite life couplings will result in fatigue failure. The

overload can be continuous torque, cyclic misalignment forces or a combination of both. Spring

couplings must not be operated in excess of the maximum capability or the failure is likely to be

sudden and dramatic.

Three Link Couplings

Link-type couplings are a variation of the disc design. One variation uses three flat strip springs

called "flexlinks" in place of a laminated disc pack. The single thickness corrosion resistant spring

steel flat strips are

proportioned to transmit torque in compressive or tensile beam end loading. These flexlinks are

attached to a high strength steel triangular plate at the inner diameter and a circular flange at the

outer side with rivets.

The riveting of the flexlinks to their connected parts is a factory assembly. The triangular plate

bolts to a delta shaped hub or a round hub that can be mounted on the shaft. The round flange is

formed with three axial arms that bolt to another round flange that has identical flexlinks,

triangular plates, and hubs. The two pieces form a double flex coupling with a torque tube between

the flex planes. The bolting is connected when the unit is assembled to the equipment shafts.

The links function the same as a disc leg. The difference is that the links are thick enough to be

loaded in compression on one side and tension on the other. The link thickness is sufficient to

prevent buckling under maximum rated loading, yet not too thick so as to bend near the rivet and

accommodate angular misalignment. The flexlinks are sized to keep stresses well within the

infinite-life fatigue limit of the steel. Spacer washers position the flexlinks so that each can flex in a

prescribed path without interfering with adjacent flexlinks. Angular misalignment is in the order of

5°, and because there are two flex planes (one at each flange), their combined angular

misalignment capability provides radial (parallel) misalignment as well. However, axial

displacement adds an additional tensile load to the link.

When the coupling is operating under misalignment the fatigue points are located at the rivets and

at the bolts in the center of the torque tube. Additional tensile loads on the flexlinks that come

from axial movement

will detract from the nominal torque capability. One could identify this via the Modified Goodman

Diagram. One variation of the flex link coupling eliminates the center bolting. It is the spacer

version or the floating shaft version. In that case a hollow tube with triangular plates at each end is

inserted between the flanges. The flanges are modified to eliminate the axial arms. The flange on

the hollow tube bolts directly to the circular flange and the misalignment forces that result in

fatigue loading are taken at the link to rivet connection.

The normal configuration of the flexlink coupling is that of a reduced moment coupling. The delta

shaped hubs will fit inside the flange torque tube for reduced moment or they can be mounted

externally for marine style. The reduced moment coupling is shorter, but is not available as a

spacer or a floating shaft version. Round shaped hubs are always mounted externally as they won't

fit into the inside of the torque tube. Round shaped hubs are the only types that are available in

stainless steel.

The "flexlink" coupling was developed for highly misaligned, but lightly loaded applications. It is

most suitable for smooth torque loads without transients. It works best when the load is not cyclic.

The reason for the load capability is that most of the fatigue or cyclic capability is used up in the

misalignment cyclic loading.

Motion Control Couplings

Typical operational requirements for Motion Control couplings include torsional rigidity, low inertia,

constant velocity, low radial stiffness, zero backlash, corrosion resistance and the capability of

cyclic (repeated start/stop/reverse) activity. Motion Control couplings can be used on 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.

Although there are many alternatives for Motion Control Couplings, the most popular include the

bellows, beam, disc and curved jaw. Covered in this section are the bellows and beam couplings.

The disc type coupling has been covered separately in the earlier portion of this section, and

curved jaw couplings (including the low backlash version for motion control) are covered in the

Elastomeric Coupling section.

Bellows Couplings

The bellows coupling uses a thin cross section tubular metal bellows formed with annular

corrugations as the flexible element. The bellows material is usually steel or stainless steel and can

be single thickness or two layers. Each end of the bellows is attached to a hub. Torque is

transmitted in shear or from twisting across the bellows. Since the bellows is thin, it is capable of

flexing for misalignment. The metal is subject to fatigue failure because the load from

misalignment is cyclic and imposed on top of the nominal torque. In addition, the motion control

application often calls for start-stop and reversal of the applied torque. The bellows coupling is one

of the most torsionally stiff types of couplings, which give it an advantage in positioning drives and

encoders. Currently its main limits are bore sizes to about 2 inches ( 50 mm) and nominal torque to

1750 inch-pounds (200 Nm). The majority of the applications are for shafts under 1 inch.

While the bellows itself is very stiff and strong, the connection of the hubs to the shaft and the

method of attaching the hubs to the bellows can be weak points. The hub must attach to the

bellows without becoming a fatigue point or a slip point.

The hub must be tightly attached to the shaft to prevent slip and backlash. Often the hubs are

split, which allows a clamping action. Other types use tapered bushings attached with keys or a

setscrew to the shaft. Shaft sizes are small and the hubs are often made from aluminum to reduce

inertia, so the shrink fit option is not used.

There are many similarities between the bellows and beam couplings described in the next section.

Although they are used for the same type of applications, the bellows is consistently more stiff than

the beam with a few exceptions.

Beam Couplings

The beam coupling is an all-metal, non-lubricated coupling made from bar stock. It is manufactured

by cutting a helix in a hollow bar forming a curved beam spring. The variations in OD, ID, number

of helix starts, width of the beams (related to thickness of coils) and number of coils within a

specific coupling length all serve to define the characteristics of the coupling.

The coupling is formed from a single piece of aluminum or stainless steel and can be cut with

single or double flex planes. Some beam models can far exceed the bellows in angular

misalignment, while others have configurations that make them act like a different device than a

coupling. The devices include U-joints and springs. The flexure allowed by the curved beam portion

of the coupling is capable of accommodating angular, parallel and axial misalignment while

continuing to transmit torque between the attached shafts.

The beam coupling has a torsional stiffness that ranges from medium high to very high, which

makes it a "constant velocity" coupling i.e. at every point during rotation, the driven half turns

exactly the same amount and at the same rate as the driving half. It is highly appropriate for power

transmission and motion control applications where extremely accurate positioning, frequent

start/stop or reversing, and zero backlash are essential. It operates either clockwise or counter

clockwise without sacrificing windup or torque capabilities. It has a very high tolerance to heat,

chemicals, and corrosion.

The beam coupling and the bellows coupling have some overlap of attributes that allows them to

be used on similar applications. In general the beam coupling, while torsionally stiff, is not as stiff

as the bellows coupling. The cross-section of the beam coupling is sufficient to allow it to function,

as a single thickness spring whereas the torsional spring coupling mentioned earlier in this section

needs to be triple wound.

Beam couplings are offered with either one setscrew or with a splitclamping hub for attachment to

shafts. The clamping hub is highly recommended for instrumentation type applications due to the

reliability of the attachment. It inhibits any movement or slip between the hub and the shaft.

Beam couplings operate with 3°, 4°, or 5° of angular misalignment, depending on the size and type

selected. Some special variations of flexure will handle 90°, and can serve as U-Joints. They will

also accommodate lateral translation. The maximum angular and parallel offsets cannot occur at

the same time. The amount of parallel offset depends on size and configuration of the coupling, but

it ranges from 0.004 to 0.035 inches. Beam couplings will allow axial motion up to .010 inches

(.254 mm) in the small units and twice that in large units. However, care must be taken at the time

of installation not to compress or expand the coupling axially, as the beam should be relaxed when

installed.

The beam coupling can be used up to 10,000 RPM while misaligned and produces very low

reactionary loads. Consequently the support bearings will be lightly loaded even when the coupling

is misaligned within its specified limits. Reactionary loads are proportional to the spring flex

constant.

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 because

elastomers 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-carrying

capabilities. 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.

Application Considerations

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.

System Torque

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 Torque

The torque needed when the system starts its operation. This torque can be greater or less than

the normal operating torque.

Peak Torque

The 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 Torque

It 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 Torque

This is a cyclic torque that passes through zero and becomes negative or "reverses" to the

opposite direction.

Transient Torque

A 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 Torque

It 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 Torque

In 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 Torque

This 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 Torque

The 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 Decelerate

The 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 Torque

The 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 Torque

This is the maximum torque available from the driver, it may not be able to operate for extended

time periods at this torque.

Stall Torque

This 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 more

torque 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 Torque

The 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 Factor

The 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 Factors

Sometimes "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 Factor

A 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 Catalogs

Coupling 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 / RPM

Where

T = the torque in inch-pounds

BHP = the motor or other horsepower

RPM = the operating speed in revolutions per minute

63025 = a constant used for inch-pounds, use 5252 for foot-pounds, and 7121 for Newton-meters

Converting kW to torque

Where

T = BHP x 84518 / RPM

T = the torque in inch-pounds

kW = the motor or other kilowatts

RPM = the operating speed in revolutions per minute

84518 = 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-pounds

Wk2 = the inertia of the piece to be accelerated or decelerated in pound

feet squared

N = the absolute change of speed in RPM

t = the time for the speed change in seconds

307 = 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 Control

While 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.

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 Frequency

All 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 driver

JL is the polar inertia of the driven

Ctdyn 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.5000

Campbell 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 Frequency

In 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 Systems

Torsional vibration problems appear primarily in four types of applications briefly discussed here.

High Speed Machines

High-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 and

the 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 Drives

The 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 Motors

Synchronous 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 Drives

There 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 Systems

In 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 Torque

This 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 Torque

It 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.

Conclusions

Torsional 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 Size

Shaft 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 Transmission

Several 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 a

shrink 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 Disks

Shaft 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 Attachments

Couplings 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.

Splines

Splines 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.

Coupling Alignment

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 Flexibility

We 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 Misalignment

Parallel (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 Misalignment

Angular 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 Displacement

Axial 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 Type

The 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 Misalignment

A 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 Misalignment

Shaft 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 Installation

Alignment 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 Alignment

Dial 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 Alignment

Laser 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.

Conclusions

Proper 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) Balance

Balance 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.

Vibration

Vibration 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 pounds

RPM = 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 Speed

All 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 Balance

The 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 Coupling

Good 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. In

effect 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 Balancing

While 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 Standards

The 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

References

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.