Shaft Couplings for Special Purpose Rotary Machines

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Engineering Design Guide:
GBHE-EDG-MAC-1101
Shaft Couplings for Special
Purpose Rotary Machines
Process Disclaimer
Information contained in this publication or as otherwise supplied to Users is
believed to be accurate and correct at time of going to press, and is given in
good faith, but it is for the User to satisfy itself of the suitability of the information
for its own particular purpose. GBHE gives no warranty as to the fitness of this
information for any particular purpose and any implied warranty or condition
(statutory or otherwise) is excluded except to the extent that exclusion is
prevented by law. GBHE accepts no liability resulting from reliance on this
information. Freedom under Patent, Copyright and Designs cannot be assumed.

Engineering Design Guide: Shaft Couplings for Special
Purpose Rotary Machines
CONTENTS SECTION
0 INTRODUCTION
1 SCOPE
2 PRELIMINARY COUPLING SELECTION
2.1 Establish Data
2.2 Type Selection
SECTION 1 - COUPLING FEATURES BY TYPE
3 GEAR TOOTH COUPLINGS
3.1 Crowning
3.2 Application Envelope
3.3 Oil Lubrication
3.4 Grease Lubrication
3.5 Forces on Machine Bearings
3.6 Balancing
4 DIAPHRAGM COUPLINGS
4.2 Forces on Machine
4.3 Natural Axial Frequency of the Spacer
4.4 Stability of Paired Diaphragms
4.5 Balancing
4.6 Contoured Single Diaphragm
4.7 Multiple Laminate Diaphragms

5 ELASTOMERIC ELEMENT COUPLINGS
5.2 Elastomer Behavior
5.3 Design Factors in Selection
5.4 Balance
5.5 Alignment
5.6 Materials
6 QUILL SHAFT COUPLINGS
6.1 Stresses
7 CRITICAL SPEEDS
7.1 Intrinsic Natural Lateral Frequency of Spacer
7.2 Effect on Rotor Dynamic Response
8 LIMITED END FLOAT
9 COUPLING HUB ATTACHMENT
9.1 Attachment by Interference Fit
9.2 Friction Drive Elements
10 ACCESSORIES
10.1 Alignment Measurement
10.2 Torque Measurement
10.3 Guards
SECTION THREE - PURCHASE PROCEDURES

11 SPECIFICATION
11.1 Supplementary Clauses to API 671, 1st Edition 1979
11.2 Clauses Requiring Purchaser Decisions
12 VENDOR CO-ORDINATION MEETING AGENDA
TABLES
1 SERVICE FACTORS
2 COUPLING TYPES AND ATTRIBUTES
3 OPTIMUM RUNNING MISALIGNMENT
4 LUBRICATING COUPLING LIMITS
5 MACHINE CO-ORDINATOR

FIGURES
1 PRELIMINARY SELECTION SEQUENCE
2 COUPLING EXPERIENCE
3 COUPLINGS AVAILABLE
4 GEAR COUPLING SELECTION SEQUENCE
5 DEFINITION OF CROWNING
6 APPLICATION ENVELOPE FOR GEAR TYPE COUPLINGS
7 SELECTION OF MATERIALS FOR GEAR COUPLINGS
8 TYPICAL GEAR COUPLING AND PITCH CIRCLE DIAMETER
9 GUIDE TO SPACER MASS
10 GENERAL LIMITS FOR LUBRICATION
11 DIAPHRAGM COUPLING SELECTION SEQUENCE
12 DIAPHRAGM COUPLINGS
13 ELASTOMER ELEMENT COUPLING SELECTION SEQUENCE
14 APPLICATION OF ELASTOMERIC-ELEMENT COUPLINGS
15 CHANGE IN ELASTOMER COMPLIANCE WITH FREQUENCY AT
VARIOUS IMPRESSED STEADY LOADS
16 HYDRAULIC FITTING AND REMOVAL OF HUBS
17 PERMISSIBLE TORQUES FOR TAPER RING DRIVERS
BIBLIOGRAPHY
DOCUMENTS REFERRED TO IN THIS GBHE ENGINEERING DESIGN GUIDE

0 INTRODUCTION
Couplings between shafts may perform more than one function, thus
hydrodynamic couplings can also act as variable-speed devices.
Rigid couplings permit no relative movement of the shaft ends. They range
from flanges integral with the shaft, to sleeve couplings which permit easy
dismantling (exemplified by the oil-injected OK-HB device made by SKF
Steel) and for small shafts, by the shrink disc.
However, in this GBHE Engineering Design Guide we consider only
flexible couplings which permit relative misalignment of the rotating shafts
of machines and drivers typically employed in petrochemical plant service.
1 SCOPE
This GBHE Engineering Design Guide covers the choice and specification
of flexible shaft couplings for transmitting powers above 75 kW between
special-purpose rotary machines.
The Guide does not cover allied devices exemplified by:
(a) Variable-speed couplings.
(b) Freewheel couplings.
(c) Clutches.
(d) Very low speed couplings used for mechanical handling
equipment.
(e) Torque limiting couplings for reducing the starting load
on electric motors.
(f) Line shaft couplings for deep well vertical pumps.
Additional considerations beyond the scope of this Guide are needed to
cover couplings for reciprocating machines or for other machines where
large regular cyclic torque fluctuations are encountered.
Supplementary clauses to API 671 purchase specification are included.

FIGURE 1: PRELIMINARY SELECTION SEQUENCE

2 PRELIMINARY COUPLING SELECTION
2.1 Establish Data
(a) Maximum power transmitted.
(b) Maximum speed of operation (not necessarily at maximum power).
(c) Service factor from Table 1.
TABLE 1: SERVICE FACTORS (abstract from API.613. Ed.3)
Enter the Experience/Availability Charts (Figures 2 and 3) to obtain the
preliminary choice of coupling.

2.2 Type Selection
Consider the following topics for type selection:
FIGURE 2: COUPLING EXPERIENCE

FIGURE 3: COUPLINGS AVAILABLE

(a) High Power
On large machine trains, a flexible coupling duty may lie outside the
operational envelope of Figures 2 and 3, thus leading to re-arrangement of
the train to split the power flow. Alternatively the machine shaft ends may
be coupled rigidly or by a quill shaft.
(b) High Transient Torque
Electric motor drivers are inherently capable of high transient torques.
When the drive is through a gearbox, it is current practice to insert a
torsionally compliant coupling, typically an elastomer coupling, between an
electric motor and the adjacent gearbox where:
(1) The driver is an induction motor with a switched start device such
as the Korndorfer System.
(2) The driver is a synchronous motor with salient poles.
(3) The speed-increasing gearbox is an epiyclic type.
Note: Elastomer-element couplings need not be the first choice for:
(i) Direct-coupled motors.
(ii) Parallel-shaft gearboxes used in conjunction with induction motors.
(c) High Speed
Gear tooth couplings have been the first choice for high speed
machines defined by:

(d) Low Forces or Moments applied to Shafts
Where a minimal forces on thrust and radial bearings are required
then diaphragm couplings are the first choice.
As an example, the following loads are representative for a
coupling transmitting 15 MW at 108 r/s with parallel offset of 1.0
mm per meter of spacer length:
Moments:
Gear coupling tooth pitch dia 230 mm 7270 N m
Contoured diaphragm dia 420 mm 124 N m
Continuous rig diaphragm dia 350 mm 28 N m
Axial Forces:
Gear 30 kN
Diaphragm 1 kN
Note: If the gear coupling is worn, the value will significantly
increase.
(e) Drive Continuity
Coupling element failure may disconnect the drive. Such
disconnection may cause an immediate hazard. Review
instrumentation required to alert the operators on failure.
Auxiliary back-up coupling mechanisms may be required.
(f) Regular Lubrication Unacceptable
Select a diaphragm coupling or an elastomer element
coupling.
(g) Zero Backlash Drives
Where torque reversals may occur and zero deadband
transmission is required, for governors or other controllers,
then select a diaphragm coupling.

(h) High Offset
Specify a spacer coupling. Note that long spacers are
sometimes termed ‘Cardan Shafts‘.
The spacer length for a gear coupling is determined by the
maximum angular misalignment, which is given by the limit
on tooth sliding speed (see Clause 3.2(b)).
As a first estimate take this maximum misalignment as:

SECTION 1 - COUPLING FEATURES BY TYPE
3 GEAR TOOTH COUPLINGS
The sequence of selection is shown in Figure 4.
3.1 Crowning
Crowning the teeth is necessary to avoid edge loading due to
misalignment [6]; it is illustrated in Figure 5. Do not confuse crowning with
edge relief; the latter applies over only a small part of the face width to
eliminate tooth end effects on ordinary gears. Note that crowning is
sometimes called barreling.
FIGURE 5: DEFINITION OF CROWNING
A crowned gear tooth coupling has an optimum running misalignment,
from the point of view of lubrication and coefficient of friction, as shown in
the following table [12].

TABLE 3: OPTIMUM RUNNING MISALIGNMENT

FIGURE 6: APPLICATION ENVELOPE FOR GEAR TYPE COUPLINGS

The limits shown in Figure 6 are derived from the following empirical
design criteria:
(a) Resistance to Fretting
Dudley estimated the fretting wear limits for involutes splines [1].
Following Woodley [7], his results have been re-presented in Figure
7.
FIGURE 7: SELECTION OF MATERIALS FOR GEAR COUPLINGS

(b) Boylan limit on tooth sliding speed [2].
For typical turbine and compressor couplings, V should be less
than 0.12 m/s for continuous running. Above this value severe wear
of a form known as 'worm tracking' occurs [6] unless the lubricant
viscosity is high enough to ensure an oil film thickness above 0.5
μm. Low speed (< 15 r/s) couplings with high crowning and high
viscosity lubricant have operated with sliding speeds up to 0.40
m/s.
(c) Contact Stress Limit
The contact stress is indirectly assessed by the tooth contact load
factor (analogous to Lloyds K factor for gears), obtained from:

(d) Conti-Barbaran Limit
Conti-Barbaran [3] developed a criterion defined as:
Figures 8 and 9 give some guidance on spacer mass (hub with male
teeth) [6].
Elastic deformation of the coupling muff by the combination of speed and
residual unbalance leads to increased load on the teeth.
The criterion was developed for couplings typical of marine practice for
which B should exceed 10. For higher hardness teeth, lower minimum
values are permissible; given by:

FIGURE 8: TYPICAL GEAR COUPLING AND PITCH CIRCLE DIAMETER

FIGURE 9: GUIDE TO SPACER MASS
3.3 Oil Lubrication
Typical oil flow needs at each end of a coupling [4] are shown in Figure
10.
With short male teeth driving, tooth inclination is favorable for generating
hydrodynamic oil pressure. At the other end of a symmetrical coupling the
opposite applies. If the misalignment at one end of the coupling is very
small, local fretting can occur. Oil viscosity is an important parameter.

Coupling wear reduces with increasing oil viscosity up to a 40 cSt at
100°C (Viscosity Grade VG 680) [5]. The minimum operating viscosity is
50 cP, equivalent to VG 46 at 40°C or VG 32 at 30°C. Hot high-viscosity
grade oil is better than cool low-viscosity grade oil that has the same
operating viscosity.
Coupling lock-up can be caused by fine solids (< 15 μm) centrifuged out of
the oil. Inbreathing at bearing labyrinths is one source of contamination.
The return of bearing housing vents to an oil drain header reduces this.
3.4 Grease Lubrication
A widely quoted limit is a pitch-line velocity of 40 m/s but the centrifugal
acceleration at the pitch line is a more relevant parameter than velocity.

The type of oscillating motion that occurs at the tooth contacts is best
lubricated by a fluid, but most manufacturers recommend grease as it is
easier to retain in the coupling than oil. In the centrifugal field grease tends
to separate into oil, which escapes from the enclosure by creep, and a
hard soap-rich phase which is ineffective as a lubricant. Experience with
lubricated couplings gives the following limits:
Re-lubrication implies that the coupling has to be split to allow the old
grease to be removed completely. These re-lubrication periods make
grease-packed couplings unsuitable for machines in Reliability Class 1
and 2 as defined in GBHE-EDG-MAC-1101.
An alternative to grease lubrication is to use a semi-fluid soap thickened
polyglycol, such as Shell Tivela Compound A. The soap content of Tivela
Compound A is lower than that of a grease so that the material remains a
fluid. The polyglycols are more effective as lubricants than mineral oils.
Experience with Shell Tivela Compound A is that it provides satisfactory
lubrication without changing for two years, thus making lubricant-filled
couplings suitable for Reliability Class 2 operation. Lubricant
retention is ESSENTIAL.

A misalignment of up to 2 mm/m is common [7].
The direction of the load is at an angle to the direction of offset, typically
trailing 40° in the direction of rotation.

3.6 Balancing
Dynamic balance of the complete coupling requires tip centering or,
(where the tooth flanks define the position of the spacer under load)
locked in torque. Tip centering is likely to be required [7] if:
If the complete coupling cannot be balanced with locked in torque or tip
centering, the spacer should be balanced when centered on its outside
diameter.
Tip clearance in operation depends on growth of the female coupling half,
distortion due to shaft/hub interference and the radial component of
contact force.
Tip centered components may not be interchangeable.
4 DIAPHRAGM COUPLING
The selection sequence is shown in Figure 11.
This type of coupling sub-divides into two classes:
(a) Single contoured diaphragm (Type I in Figure 12).
(b) Multiple diaphragm (laminate) either spoked or continuous ring,
transmitting torque by shear forces or tension in alternate sections
respectively.
Continuous ring couplings (Type III in Figure 12) are smaller in
diameter than the others for the same torque, they also continue to
drive on diaphragm failure by interference between the diaphragm
fixings. They permit higher misalignments.

FIGURE 11: DIAPHRAGM COUPLING SELECTION SEQUENCE
See Figures 2 and 3.

4.2 Forces on Machine
(a) All diaphragm couplings should be installed so that at operating
conditions they are not axially offset. The axial load is then
insignificant.
(b) Radial forces due to parallel misalignment are then:
4.3 Natural Axial Frequency of the Spacer
Any axial natural frequency band (due to a non-linear axial spring rate)
shall be avoided. The low end of the band is given by:

4.4 Stability of Paired Diaphragms
Couplings are produced commercially with an enhanced misalignment
tolerance by using paired diaphragms. Paired diaphragms are forbidden
for use at either end of a spacer coupling, unless an unlubricated
centering bearing is provided, as they provide no radial location and the
spacer mass centre will move to an eccentric position, over-stressing the
diaphragms at high speeds.
4.5 Balancing
The hubs are balanced individually. The diaphragm unit is balanced while
mounted between both hubs.
4.6 Contoured Single Diaphragm
This is illustrated as Type I in Figure 12.
The diaphragm is contoured to decrease in thickness from hub to rim. Key
parameters are:-
(a) Stresses
Stresses should be less than 50% of the fatigue limit at 107 cycles.
Coarse surface finish leads to scatter in fatigue test results,
resulting in an undefined endurance limit.
Axial deflection leads to high stresses at the hub and centrifugal
forces are highest at the hub. The torque load gives only small
incremental stresses.
Angular misalignment leads to similar stress levels for a tapered or
parallel diaphragm provided that:
For constant shear stress, the thickness is inversely proportional to
the radius squared.
Radial corrugation can allow more axial deflection.

(b) Cooling
The disc in this type of coupling is of necessity a large
diameter in order to gain the required flexibility. High
windage losses due to the large diameter can result in high
temperatures:
Some installations have been reported to need an oil jet or
air blast for cooling.
(c) Materials
The material selection is typically AISI 4340 alloy steel,
vacuum melted and cross-rolled to obtain isotropic
properties. Heat treatment is applied to obtain a tensile
strength of about 1200 MN/m2. Resin coating may be
required to prevent corrosion. This coating also permits rapid
checking for localized mechanical damage.

FIG 12: DIAPHRAGM COUPLINGS

4.7 Multiple Laminate Diaphragms
Type II (Figure 12), typified by the Metastream M series, transmits power
by subjecting the laminate to shear forces due to the differing diameters of
the driving and driven bolt circles. In this type, the flexible element is built
up from a number of thin steel laminates in a single pack, clamped by
bolts. The laminates are 0.5 mm to 0.25 mm thick.
Experience has shown that AISI 310 material is satisfactory, being
resistant to corrosion and suitable for stamping in volume manufacture.
The individual laminates are machined and finished ground together in a
pack of the size intended in the application. After this stage the packs are
never broken down or shuffled. The packs and precision-machined
clamping washers are held together by bolting carefully torqued to about
75% of the clamp bolt proof load. This procedure drastically reduces the
probability of fretting failure in the pack.
Type III (Figure 12), is typified by the Metastream T series or Thomas SN
and 51, 52 series. Torque is transmitted only by tensile loads in half of the
ring. The section of the stressed element may be shaped to obtain
continuity of strain energy, producing a coupling better able to withstand
shock loads. Four bolt ring type couplings do not transmit steady torque
and velocity when angularly misaligned; ideally 5 to 8 bolts should be
used.
Type IV (Figure 12), is a variation of Type III, having separate links. It is
extremely difficult to match the links during assembly; thus load sharing is
imperfect and the life expectancy erratic. This type should be avoided.
(a) Application Envelope
The use of these couplings is unrestricted up to the boundary given
by:
Above this boundary, a vibration analysis should always be done.

(b) Natural Frequencies
For the ring type (Figure 12, Type III), use as a rough guide:
For spoked and ring types, when the slack is taken up, the axial spring
rate is non-linear. Thus the natural frequency depends on the amplitude,
increasing as the excitation increases.
Current practice is to avoid operation at a running speed above the
coupling natural frequency unless the margin exceeds 30%. Operation
below the speed corresponding to the natural frequency for infinitesimal
amplitudes is acceptable when the margin exceeds 6%.
The axial natural frequencies of Metastream spacer and coupling
elements are typically within the range 65 to 95 Hz; consequently such
couplings should not be used above 60 r/s without specific review of
possible excitation. Note that this only refers to resonance of the spacer.

FIGURE 13: ELASTOMER ELEMENT COUPLING SELECTION SEQUENCE

5 ELASTOMERIC ELEMENT COUPLINGS
For the selection sequence of the Elastomer Element Coupling, see
Figure 13.
The limits applicable to commercially available couplings are shown in
Figure 14.
It is possible to operate up to 80 r/s in the small sizes, the limit being set
by the precision of manufacture and the dynamic response of the
elastomer element.
Above the 'normal' limit line couplings may be made individually.
Consequently accuracy depends on operator skill. It has been necessary
to specify on the following manufacturing accuracy for drives up to 25 r/s.
(a) All cylindrical surfaces to be concentric to ±0.1 mm radially and 1 in
10,000 maximum angular deviation relative to the axis.
(b) All radial surfaces to be flat to ±0.1 mm and equi-spaced to 1 in
10,000.
(c) Surface finish, where there is contact with elastomeric blocks, to be
better than 3.2 μm Ra.
5.2 Elastomer Behavior
The non-linear elastic modulus and the intrinsic hysteresis of elastomers
tend to suppress torsional resonance. The fundamental torsional mode of
the machine set should occur at less than 20% of normal running speed
under normal load conditions and should not exceed 60% under maximum
acceleration.
The properties of elastomers are not well defined nor are they time-stable.
The elastomer dynamic compliance decreases with increasing frequency
[14], see Figure 15. Hardness has a major effect. The design of inserts
should be based on a minimum Shore Hardness of 60, though for on-site
modifications a hardness value of 45 is permissible.

The shape of the inserts relative to the shape of the cavity in the body of
the coupling has an effect on compliance and the movement under
centrifugal load affects balance. When operating close to the speed limit
line a wedge shape insert is needed to minimize the average stress by
ensuring a close fit in the coupling body at the periphery.
For unprotected machines in cold climates check the 'glass-point'
temperature of the rubbers. The minimum operating temperature is then
48°C above this temperature.

FIGURE 14: APPLICATION OF ELASTOMERIC-ELEMENT COUPLINGS

5.3 Design Factors in Selection
Couplings should be selected on the basis of being capable of continuous
operation at the unfactored nominal electric motor peak torque.
The coupling design should permit the renewal of elastomer inserts
without disturbing the driver or driven shafts or the coupling hubs.
Cardan shaft type couplings should use non-lubricated centering bearings.
Coupling types which rely on placing the rubber in shear should be
avoided since progressive failure is possible with resulting loss of drive
continuity.
5.4 Balance
Insert sets are supplied in match weighed sets, normally within 0.5%
band. It is essential that the batch is clearly identified and is not mixed in
any way with a different batch. Inserts should be changed only as
complete sets. Silicone oil is used to ease assembly.
5.5 Alignment
Experience on Holset WB couplings shows that residual misalignment at
start-up will be locked-in under the heavy load peculiar to electric motor
drivers. It is necessary to specify to the machine manufacturers that:-
(a) Cold parallel misalignment shall not exceed ± 0.02 mm.
(b) Differential displacement from cold to hot operating conditions shall
be limited to 1/500 of the effective spacer length or 0.20 mm
whichever is smaller.
These requirements constrain the machine maker's choice of support; it
may be necessary to use centre-line mounting.

5.6 Materials
Coupling bodies should be made from carbon steel or nodular cast iron.
Grey cast iron should be prohibited.
Resin laminate material used for limit stops should mate with 13% chrome
steel hardened to 380-420 HV.

6 QUILL SHAFT COUPLINGS
6.1 Stresses
The resultant stresses, constant and cyclic, in a quill shaft should not
exceed 50% of those that permit a life in excess of 107 cycles for the
particular material.
For solid uniform section circular steel shafts, the maximum stresses are:

SECTION TWO - EFFECTS ON COUPLED MACHINES
7 CRITICAL SPEEDS
7.1 Intrinsic Natural Lateral Frequency of Spacer
A long coupling spacer should have a natural frequency in bending
at least 50% above running speed.
Neglecting friction at the points of articulation, i.e. assuming pin jointed
ends [4]:
7.2 Effect on Rotor Dynamic Response
Where the ratio of central mass (including shaft between bearings) to shaft
end mass (including shaft overhang) is less than 0.25 and the ratio of
overhang to span is less than 0.3 the reduction in the first critical speed
will be less than 10%.
The second critical speed can be dramatically affected by overhung mass.
This can be assessed by mathematically modeling the rotor (e.g. Current
computer program are ROMAX and BIRD).

8 LIMITED END FLOAT
Under driving torque, both gear and elastomer block couplings require a
significant force to cause relative axial displacement of the shaft ends.
This is likely to be greater than the axial magnetic centering forces of a
motor thus, as the shafts expand thermally, the motor rotor will be
displaced.
On run down the torque is reduced but the magnetic force is absent so the
running position is held until the coupling comes to rest. On cooling in the
absence of driving torque the coupling is likely to permit axial movement,
thus increasing the displacement for the next start. Current practice is to
specify a limited end-float feature such that the motor rotor is held off its
bump stops.
Diaphragm couplings normally provide sufficient axial restraint for sleeve
bearing motors during run up and run down and while reaching operating
alignment. Where the axial movement is large or the master state is
misplaced significant axial forces can be generated while the motor rotor
is off its magnetic centre.
9 COUPLING HUB ATTACHMENT
Coupling hubs integral with the shaft are preferred but frequently
cannot economically be provided.
9.1 Attachment by Interference Fit
The preferred method of attaching coupling hubs to the shaft is by
interference fit, without keys. See Figure 16.
Installation can be by hydraulic expansion of the hub bore together with
hydraulic control of axial position [13]. Expansion pressures of 1500-4000
bar and pushing pressures of 500-1000 bar are typical. Robust stops to
prevent ejection of the hub during hydraulic removal are necessary [15].
Use of threaded components for assembly in poor conditions can lead to
shaft end damage or failure to fit the hub.
Thermal expansion demands that there is no delay or pause in fitting;
temperatures in excess of 250°C should not be used.

Precise final axial location of the hub and specified pull up need
to be achieved where a taper surface is employed.
9.2 Friction Drive Elements
For smaller sizes of coupling (<80 mm dia), split tapered bushes are
conveniently employed, e.g. Fenner 'Taperlock' bush. Slip torques are
shown in Figure 17.
Systems based on wedging action but not needing taper machining of
shaft or coupling are the Ringfeder System (suitable for shaft diameters
from 100-500 mm dia) or the SKF 'SH' expansion sleeve. Such
attachment methods should NOT be used where the coupled machines
operate with continual impulsive torque, e.g. a Pelton type hydraulic
turbine.

FIGURE 16: HYDRAULIC FITTING AND REMOVAL OF HUBS

Fig 17: PERMISSIBLE TORQUES FOR TAPER RING DRIVERS

10 ACCESSORIES
10.1 Alignment Measurement
Machines in Reliability Class 1 (see GBHE-EDG-MAC-1101) may need
provision for shaft alignment measurement whilst in operation.
Methods of achieving this are:-
(a) External displacement probes on coupling surfaces.
(b) Measurement of coupling articulation angle by displacement probes
and equipment mounted in the coupling spacer, the power and
signals being transmitted to and from the spacer without contact.
(Indikon system, used on the T8 Air compressor). This equipment
complicates the balancing procedure.
(c) Measurement of movement of bearing housings/bearings by
proximity probes registering on extension rods fitted to the bearing
housing/bearing (Dynalign system).
Generally once the pattern of machine relative movement is established
further measurement is less necessary. The equipment may be removed
for re-use.
10.2 Torque Measurement
Torque measurement devices are available (e.g. from Vibrometer
Stockport, Torquemeter Northampton) built into coupling spacers and may
have application during investigations (e.g. proof of test facilities). NEL
East Kilbride have a selection of these and provide an on-site measuring
service. NEL have units of torque capacities: 13, 8, 5, 2.8, 1.7 kN.m. A unit
of 20 kN.m is produced. Speed limitations are 50 r/s for the largest unit,
200 r/s for the smallest. The largest unit is 900 mm long by 350 mm
diameter and weights 70 kg. This coupling has an outer stationary
component which requires support.

10.3 Guards
Coupling guard design should allow the guard to follow machine
movement. Without this, the guard may transmit loads from one machine
to the other, or be distorted and permit oil leakage.
Consider the effect of compressor seal gas leakage into the coupling. The
guard design should be suitable for the maximum pressure achievable
from either purge gas or the leakage unless a bursting disc is provided.
Pressure testing will be necessary.
Diaphragm and elastomer-element couplings need ventilation to dissipate
the heat arising from windage and mechanical hysteresis losses.
SECTION THREE - PURCHASE PROCEDURES
11 SPECIFICATION
Shaft couplings for special-purpose rotary machines should be purchased
in accordance with API 671 with supplementary clauses as follows:
11.1 Supplementary Clauses to API 671, 1st Edition 1979
Clauses are numbered as the relevant clause of API 671 and are followed
by an indication of the nature of the supplement as:
Add : an additional requirement to API clause.
Mod : a change to the API clause.
New : a clause on a topic not covered by API.
Opt : selected option from those stated.
Note: For Machines in Reliability Classes 1, 2 and 3 refer to GBHE-EDG-
MAC-1101.
2.1.7.5
(Add) Any sealing components for hydraulically fitted hubs shall be
replaceable with the coupling hub in position on the shaft.
2.1.8.3
(Add) A matched set of plug and ring gauges shall be supplied by the
machine vendor for each taper shaft end.

2.1.8.5
(Add) Where there is more than one connection in the hub for pressurized
oil, each connection shall be permanently identified to show its
function.
2.1.9.1
(Opt) Keyed hubs shall have two keys diametrically opposed.
2.1.9.2
(Add) or BS 4235 close fit.
2.1.11.4
(Add) The required bolt extension shall also be stated.
2.1.14
(Add) The type and size of coupling shall be marked close to the Serial
Number.
2.2.3
(Opt) Limited end float feature shall be provided if one of the coupled
machines does not have an axial location bearing.
2.2.9
(Add) Flooded mesh designs with oil retention shall not be employed. (I.e.
there shall be no stagnant areas in the vicinity of the gear teeth.)
2.3.3
(Add) Dismantling of factory assembled membrane parts shall be made
impracticable.
2.3.5
(Add) The Purchaser shall furnish any cooling requirements specified by
the coupling Manufacturer.
2.3.8
(Add) Radial movement of the spacer assembly following diaphragm
failure shall be positively limited.
2.5.3.1
(Opt) Component balance and assembly check balance shall be carried
out.

3.6
(Add) Fastenings to BS 3692 Strength grade 8.8 or better are acceptable.
4.2.4.1
(Opt) Axial natural frequencies shall be measured and recorded where
such results from a coupling made to the same dimensions are not
available and where the results are not load/speed dependent.
4.3.1.2
(Opt/Add) Couplings shall be protected:
(a) By waxy film preservative (where all parts are readily
accessible for cleaning).
or
(b) Vapor phase inhibitor.
or
(c) A dry atmosphere, maintained by desiccant, with a visible
dryness indicator.
In all cases a substantial outer wooden crate shall be provided. In
cases (b) and (c), a transparent vapor barrier shall surround the
coupling.
4.3.1.2
(Opt) Couplings purchased as spares shall be assumed to have a
storage life of 8 years in a dry store.
4.3.1.3
(Opt/Add) The outer container shall be marked with order number, GBHE
project number and GBHE equipment number and, where there is
more than one coupling in a drive train, the location within the drive
train.
4.3.2.4
(Opt) A stop ring, or equivalent means, shall be provided to permit axial
positioning of a taper fitted coupling without measurement being
required.

4.2.3.4
(Opt/Add) A puller for keyed hubs shall be provided and permanently
labeled with equipment number, coupling location (if more than one
coupling in drive train), size and type.
4.3.2.5
(Add) Means shall be provided for fixing any trapped coupling
components to permit the driver to be run independently of the
driven machine without disturbing either machine or
removing a coupling hub.
6.2.2
(Add) Item 3; data shall be provided for each assembly bounded by an
elastic element e.g. hubs, spacer, hub and guard disc, centre spool.
11.2 Clauses Requiring Purchaser Decisions
2.1.5 If integral hubs required (preferred but only if no components need
to be removed over shaft end).
2.1.6 The coupling vendor should provide drilling jig for coupling flanges
integral with shaft.
2.1.7.1 Interference fit of hub. Where not hydraulically fitted the hub shall
be assumed to be heated by 200°C for fitting.
2.1.7.1 Hydraulically fitted hubs without keys should be used wherever
possible to avoid shaft end damage and facilitate maintenance.
Procedures should ensure follow up of hub as it is expanded to
avoid seal extrusion.
2.1.8.1 Taper for hydraulic fitting of 1:30 or finer is required and can be
provided by a thin tapered sleeve thus permitting a parallel shaft
end.
2.1.8.2 Keyed hub taper required if not 1:16 on diameter.
2.1.9.1 Where keys are used they shall be displaced at 180° to each other.

2.1.12 Stray circulating currents can cause bearing damage if complete
circuits exist through the bearings. Static charges can build up on
rotors of non-electric machines for example due to wet steam
impingement. Rotor earthing in these circumstances is required
when the rotor is supported in plain bearings.
2.2.2 If external teeth on spacers or hub are preferred.
2.2.8 Gear couplings shall be continuously lubricated unless lubrication is
by semi-fluid Polyglycol grease in a sealed coupling.
2.2.10 For Class 1 machines, oil flow to each mesh shall be via two
nozzles. The preferred distribution is via drillings in the tooth roots.
2.5.1.5 For Reliability Classes 1, 2 and 3 (see GBHE-EDG-MAC-1101)
balancing machine sensitivity shall be recorded and reported.
2.5.3.1 For Class 1 machines the allowable mass eccentricity on check
balance is the lower of 1 - (n/nc)2
times the eccentricity’s permitted
for the rotors carrying the coupling, applied to each coupled rotor in
turn [11].
For Class 2 and 3 machines, check balance shall be to BS 5265,
G.6.3.
2.5.4 For Classes 1, 2 and 3, trim balance holes shall be provided.
3.5 Specify environmental contaminants to be allowed for.
4.2.4.3 Oil-fed couplings shall have the oil flow to each nozzle measured
and demonstrated.
12 VENDOR CO-ORDINATION MEETING AGENDA

TABLE 5: MACHINE CO-ORDINATOR
Machine co-ordinator has to specify:

BIBLIOGRAPHY
1 Dudley, D W ”How to Design Involute Splines” and ”When Splines need
Stress Control”. Product Eng. New York Oct 57 pp 75-80, Dec 57 pp 56-
62
2 Boylan, W ”Marine Application of Dental Couplings” Naval Architect and
Nav Engrs, Marine Power Plant Symposium Paper 26, May 1966.
3 Conti-Barbaran, D ”Some Remarks on Tooth Type Flexible Couplings”.
Marine Engr and Naval Architect, November 1963, pp 544-546.
4 Staedli, 0 ”Tooth couplings”. Maag Gear Co October 1973.
5 Calistrat, M M ”Wear and Lubrication of Gear Couplings” Mech Eng,
October 1975. International Conference on Flexible Couplings.
University of Sussex, June 1977.
6 A B Crease Design Principles and Lubrication of Gear Couplings
Paper B.1.
7 B J Woodley Materials for Gear Couplings Paper B.2.
8 A B Crease Forces Generated by Gear Couplings Paper B.3.
9 A B Crease Design Principles of Flexible Element Couplings Paper C.1.
10 E W Goody Laminated Membrane Couplings for High Power and Speed
Paper C.2.
11 J S Woodcock Effects of Couplings on Vibrations of Rotating Machinery
Paper E.1.
12 Pahl, G 'Operating Characteristics of gear-type couplings' Proceedings of
7th Turbo machinery Symposium. Texas A & M University. December
1978.
13 Oil Injection Method for Shaft Fixing SKF Publication 3064E (1977)
Reprint Ball Bearing Journal 192

14 F Stearman Tests on HOLSET Coupling Rubbers Report
15 M M Calistrat Hydraulically Fitted Hubs - Theory and Practice Koppers
Research Publication 2477.
16 BS 3170 Specification for Flexible Couplings for Power Transmission.
DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE
This Engineering Design Guide makes reference to the following documents:
AMERICAN STANDARDS
API 613 Special-Purpose Gear Units for Refinery Services
(referred to in Table 1).
API 671 Special-Purpose Couplings for Refinery Services (referred to
in Clause 1, 11 and 11.1).
BRITISH STANDARDS
BS 3692 ISO Metric Precision Hexagon Bolts, Screws and Nuts
(referred to in Clause 11.1).
BS 4235 Metric Keys and Keyways (referred to in Clause 11.1).
BS 5265 Mechanical Balancing of Rotating Bodies (referred to in
Clause 11.2).
OTHER DOCUMENTS
ENGINEERING DESIGN GUIDE
GBHE-EDG-MAC-5100 Reliability Analysis - the Weibull Method
(referred to in Clause 3.4, 10.1, and 11.1).

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