Section 7

7.3 WINCH DRIVES

Winch drums are driven through reduction gears by a motor which may be powered by hydraulic
pressure, electricity or steam. Hydraulic motors may be powered by a central power pack (which
may also be used to power other motors such as those of cranes or cargo winches), or may be
part of a self-contained winch incorporating an electrically driven hydraulic pump in each winch.
Another version of a unitised hydraulic winch drive employs a torque converter for power
transmission between an electric motor and the winch gearing.

The winch drive should ideally allow for continuous variation in line speed, permitting heaving
and rendering at high speed when the load is small, and developing high pull when the speed is
low. The pull at stalled heave should not exceed 50% of the mooring line's MBL to prevent
routine overstressing of the mooring line.

The considerations guiding the selection of drive systems at a ship's design stage will not be
discussed in detail, since many are beyond the scope of this guide: these include initial cost,
maintenance cost, energy cost, reliability, company experience, climatic conditions, availability of
steam and interconnection with other systems. There are also many variations within each
category (especially within hydraulic and electric systems) that may significantly alter the
performance, reliability and cost. Any of these are acceptable aboard tankers if they are properly
designed and executed.

Onboard arrangements should address the need to ensure that power supplies to mooring
equipment, including steam, hydraulic or electric, are sufficient and adequately protected. Where
the power source is a single hydraulic motor, alternatives in the form of a spare motor or cross-
connection capability, should be available.

The following short discussion on each basic type places emphasis on the speed/pull
characteristics:

7.3.1 Hydraulic Drives

The speed/pull characteristic varies with the type of motor:

• The speed of a low pressure vane motor can be infinitely varied up to the rated speed by

a special throttle valve. High slack rope speeds can be obtained only by suitable drive
modifications to provide a high speed range at reduced torque.

• A high pressure variable displacement hydraulic motor has an ideal speed/pull

characteristic within certain limits. To achieve a high slack line speed, a speed change
provision is required.

• A low speed high pressure radial piston motor has high torque and can be infinitely

varied up to the rated speed by the manoeuvring valve. To achieve high slack line
speed, the motor can have a dual speed step enabling high speed without any gear
arrangement.

• A torque converter drive has a similar characteristic as a variable displacement high

pressure motor and it also requires a second gear step to provide for a high slack line
speed.

7.3.2 Self-Contained Electro-Hydraulic Drives

Self-contained electro-hydraulic winches have their own hydraulic pump unit which alleviates the
need for any external hydraulic pipework. They demonstrate the same operational benefits as
other hydraulic winches.

7.3.3 Electric Drives

Electric winches have been fitted to several tankers in recent years. The advantages of electric
drives include reliable operation, reduced maintenance and the low risk of oil pollution when
compared to hydraulic drives. In general, each electric motor is equipped with a fail-safe disc
brake and each winch can be operated either individually or simultaneously with all other
winches.Electric winches are of two different types:

Section 7

• frequency controlled.

• pole changing multispeed motors.

Frequency control.

A frequency controlled electric winch is driven by a single speed electric motor controlled by a
frequency converter which performs infinitely varied speed and torque control allied with smooth
starting and stopping equivalent to hydraulic winches. The frequency control also results in
acceptable slack line speeds.

Pole changing multispeed motors.

Electric winches with pole changing motors are operated in fixed speed steps. Most common
units have three speed steps, where the first two are for normal operation and the third is for high
slack line speeds. Speed control for these winches is via on-off switching for each of the fixed
speed steps.

7.3.4 Steam

In the past, steam winches have been popular on large tankers. They are simple, safe, robust
and have an ideal speed/pull characteristic. A minor drawback of the typical twin cylinder double
acting steam engine is the non-uniform torque (varying by up to 50% for each 45% of crankshaft
rotation). This is most noticeable at very low speeds.

Because of the widespread change from steam main propulsion systems to diesel systems, now
also apparent on LNG carriers, alternative winch drive systems utilising hydraulic or electric
power have gained in popularity.

7.4 WINCH BRAKES

The brake is the heart of the mooring system, since the brake secures the drum and thus the
mooring line at the shipboard end. A further important function of the brake is to act as a safety
device in case the line load becomes excessive, by rendering and allowing the line to shed its
load before it breaks.

Ideally, a brake should hold and render within a very small range, and once it renders, should
shed only enough load to bring the line tension back to a safe level. Unfortunately, the widely
used band brake with screw application is only marginally satisfactory in fulfilling these
requirements and its operation requires special care.

The brake holding power is based on the final mooring calculations which are a function of the
ship's size and hull form. The following diagram depicts the process used to assess mooring line
characteristics and winch design parameters.

Section 7

FIGURE 7.3: CALCULATION OF MOORING LINE MBL AND RELATIONSHIP TO WINCH

PARAMETERS

7.4.1 Layers of Mooring Line on Drum

Split-drum Winch

The rated brake holding capacity is only achieved with one layer of line on the tension section of
the drum. Operation with additional layers will decrease the brake holding capacity.

As an example, the theoretical reduction in holding load for more than one layer on a 610mm
(24") diameter drum is as follows and assumes a rated brake holding capacity of 55 tonnes:

Layer of Line

Theoretical

Holding Capacity

(tonnes) (kN)

% Rated

Holding

Capacity

1 55

539

100

2 49

481

3 45

441

4 41

402

5 38

373

Undivided Drum Winch

Brake holding capacity of the undivided drum winch, as with the split-drum winch, is affected by
the number of layers on the drum. It is therefore, essential that the operator of an undivided drum
winch knows the number of layers of mooring line on the drum that the manufacturer states will
develop the design brake capacity. More layers on the drum will increase the moment applied to
the drum because of the increased moment arm. This means that a smaller mooring line load will
provide the moment which will cause the brake to slip. Conversely with less layers on the drum,
mooring line loads must be larger to cause the brake to slip. It is very possible that mooring line
loads could become excessive, thereby increasing the possibility of mooring line failure. Thus,
the operators, as in the case of split-drums, should be advised and trained in the use of the
undivided drum winch to ensure the integrity of moorings and the safety of the ship.

Ship Size and Hull Form

(input data from Shipyard/Ship Designer

Mooring Force Calculation

Ref Section 2.3

Mooring Re straint Requirements

Ref Section 2.4

Mooring Re straint Requirement/Number of

Mooring Lines in Same Group =

MBL of Mooring Line

MBL gives ‘De sign Rope’

Table 7.1 and ISO 3730

‘Design Rope’ leads to following Winch
parameters:

Brake Design Load = 80% of line MBL

Brake Holding Load = 60% of line MBL

Winch Pull: not to exceed 30% of line MBL

Drum Diameter = 16 x line diameter

Width of Tension Part = 10 x line diameter

Ship Size and Hull Form

(input data from Shipyard/Ship Designer

Ship Size and Hull Form

(input data from Shipyard/Ship Designer

Mooring Force Calculation

Ref Section 2.3

Mooring Force Calculation

Ref Section 2.3

Mooring Re straint Requirements

Ref Section 2.4

Mooring Re straint Requirements

Ref Section 2.4

Mooring Re straint Requirement/Number of

Mooring Lines in Same Group =

MBL of Mooring Line

Mooring Re straint Requirement/Number of

Mooring Lines in Same Group =

MBL of Mooring Line

MBL gives ‘De sign Rope’

Table 7.1 and ISO 3730

MBL gives ‘De sign Rope’

Table 7.1 and ISO 3730

‘Design Rope’ leads to following Winch
parameters:

Brake Design Load = 80% of line MBL

Brake Holding Load = 60% of line MBL

Winch Pull: not to exceed 30% of line MBL

Drum Diameter = 16 x line diameter

Width of Tension Part = 10 x line diameter

‘Design Rope’ leads to following Winch
parameters:

Brake Design Load = 80% of line MBL

Brake Holding Load = 60% of line MBL

Winch Pull: not to exceed 30% of line MBL

Drum Diameter = 16 x line diameter

Width of Tension Part = 10 x line diameter

Section 7

When applying the requirement for winch brakes to hold to a minimum holding load of 60% of
MBL to undivided winch drums, due consideration should be given to the number of layers of
mooring line which will be wound on a drum during normal operations. For this purpose, it should
be assumed that a minimum of 30-50 metres of mooring line will be outboard of the chock.

7.4.2 Band Brakes

Band brakes follow the same principle as wrapping a rope around a bitt or warping head to hold a
line's force. Relatively little force is required to hold a high load. This principle provides for easy
brake application, but has disadvantages such as sensitivity to changes in friction, dependency of
setting force on rope force and sensitivity to reeling direction.

The main factors which affect the actual holding capacity of the band brake are discussed
below:

7.4.2.1 Torque Applied

Figure 7.4 shows the results of a series of brake holding load tests on a VLCC when the torque
applied is varied. It can be seen that the holding load will drop appreciably when the torque
applied is lower than the recommended value. It is, therefore, essential that the operator applies
the brake properly.

FIGURE 7.4: EFFECT OF APPLIED TORQUE ON BRAKE HOLDING POWER

Shipboard inspections have shown that winch brakes are very often not torqued to their design
level and, in some tests, two men were required to apply the required torque. Hence, several
VLCC operators are now installing hydraulically actuated brakes on new VLCCs and are
retrofitting them on older VLCCs to ensure proper brake holding load. As an alternative, to
ensure proper brake holding load, many operators are installing spring-applied brakes with
hydraulic or manual release on new ships.

Whichever brake arrangement is installed, it is essential that it is operated properly to achieve
desired holding loads.

Also, it has been found that some winch actuators and controls are poorly positioned. As a result,
the operator is unable to see what is happening on deck. Proper design is essential to ensure
proper operation.

Section 7

7.4.2.2 Condition of the Winch

The physical condition of the winch gearing and brake shoe linings have a significant effect on
brake holding load capacity. Oil, moisture or heavy rust on the brake linings or brake drum can
reduce brake holding load capacity by up to 75%. Many operators run the winch with the brake
set slightly to burn off or wear off the oil or moisture. Care, however, must be taken to ensure
that excess wear is not caused by this practice when using composite brake linings. Excessive
winch speed can also reduce brake holding capacity by the build-up of heat in the composite
brake lining.

7.4.2.3 Winch in Gear

Winches should never be left in gear with the mooring drum brake on. Hydraulic or electric drives
can suffer severe damage should the brake render.

Mooring drums should always be left disconnected from the winch drive whenever the mooring
line is tensioned and the drum brake is fully applied.

7.4.2.4 Friction Coefficient

A small change in the friction coefficient will cause a large change in the holding capacity. As a
rule of thumb for typical mooring winch brakes, the change in holding capacity is twice the
change in the friction coefficient. For example, a 10% change in friction will cause a 20% change
in holding capacity. The friction change can be due to wear of the rim or brake lining, or to
changes in weather conditions, among others. This drawback can be partially offset by stainless
steel lining of brake rims and by frequent testing and recalibration.

Asbestos-free brake linings generally have a higher friction coefficient than asbestos linings and
their performance remains relatively constant in wet and dry conditions.

When changing from an asbestos to asbestos-free brake lining, it should be confirmed that the
new lining has a friction coefficient that corresponds to the original design. The brake should be
tested before the winch is returned to service and its holding power reconfirmed.

7.4.2.5 Load Dependency of Holding Capacity

Once a line load is applied to the drum, the brake band will stretch, reducing the load on the
brake controls. For this reason, a conventional screw brake can be easily re-tightened when the
mooring line is under high load, even if it was set hard originally. This means that there is no way
to determine the proper handwheel torque required once the winch is subjected to a high line
load. The danger exists that under worsening environmental conditions the brakes can be re-
tightened to the point where the line may part before the brake slips. The problem can be solved
by using spring applied brakes. The spring automatically compensates for the elongation of the
brake band, thus assuring a constant holding capacity of the brake. It also prevents the brake
from being over-tightened.

Section 7

FIGURE 7.5: SPRING-APPLIED BRAKE WITH HYDRAULIC RELEASE

A schematic of a spring-applied band brake is shown in Fig. 7.5. In the case shown, the brake is
released by a hydraulic hand pump. To apply the brake, a valve is opened to release the
hydraulic pressure. Winches with hydraulic drives can utilise the main hydraulic pressure to
release the brake. The handwheel is not used for routine operation: it serves to adjust the spring
compression during calibration, and to release the brake in case of a hydraulic malfunction. The
handwheel should be secured with a seal after each calibration to prevent tampering.
Alternatively, the brake screw can be equipped with a brake setting indicator to provide an easy
visual check of the correct adjustment of the brake setting.

A manually operated spring-applied brake is shown in Fig. 7.6. When tightening the brake screw
the spring is compressed and the compression is proportional to the holding load of the brake.
When the brake is adjusted to the correct rendering force, an indicator is adjusted so the brake
can be reset repeatedly to its correct value. This type of spring-applied brake setting indicator
compensates fully for any wear of the brake lining and elongation in the brake mechanism.

FIGURE 7.6 SPRING-APPLIED BRAKE WITH MANUAL SETTING AND RELEASE

Section 7

7.4.2.6 Sensitivity in Reeling Direction

A band brake is designed to work in one direction only. Therefore the line must always be reeled
correctly onto the drum. Each arrangement should be assessed on a case-by-case basis with
reference to manufacturer's guidance. With lines correctly reeled, tension on the line should be in
a direction that causes the free end of the band to be forced towards the fixed end, thereby
forcing the two halves of the band to close together.

Disc brakes work equally well in either direction.

Some winches are equipped with hydraulic-assist brakes where a hydraulic cylinder is used to
set the brake. A pressure gauge connected to the cylinder allows the brake to be applied with
the proper force, predetermined during brake testing. Once the brake is applied with the
hydraulic assist, the load is transferred to the mechanical linkage by tightening the handwheel
until the pressure in the assist cylinder starts to drop. Although this feature eliminates the need
for a torque wrench, the brake has the same drawbacks as the mechanical screw brake
mentioned above. This type of brake is therefore not recommended for new ships.

Since hydraulic-assist brakes and some spring-applied brakes with hand hydraulic release look
similar, it is very important that operators study the instruction book for their particular ship's
installation.

7.4.3 Disc Brakes

Disc brakes are in wide use for input brakes (see Section 7.4.4), but few, if any, winch
manufacturers offer them for drum brakes.

7.4.4 Input Brakes

Some hydraulic winches and most electric winches are provided with spring-applied brakes at
the drive motor. They are automatically applied by springs when the control lever is in neutral
and automatically released when the control lever is in the heave or rendering position (when the
motor is powered).

Some high pressure hydraulic winches are equipped with a counterbalance valve acting as a
motor brake. In these cases, it is important to apply the drum brake immediately after completion
of the mooring operation as the motor will creep in the pay-out direction due to internal leakage
in the hydraulic motor.

ISO Standard 3730 requires all electric winches to be provided with automatic brakes having a
holding capacity of 1.5 times the rated load.

Multiple-drum winches always require a brake for each drum.

Most input brakes are not rated to serve as a primary brake due to strength limitations of the
gears. If this is the case, once the ship is moored the drum must be set on the drum brake and
disengaged from the drive by means of the dog clutch. If the drum is left engaged and the drum
brake is set, both brakes will work in unison (rated holding capacity is directly additive). The
combined holding capacity will exceed 100% of the line's MBL, an undesirable result for reasons
explained earlier. (See also 7.4.2.3).

7.4.5 Winch Brake Testing

Regardless of the brake type, periodic testing is essential to assure a safe mooring. The following
provides a guide for testing of mooring winch brakes:

7.4.5.1 General

Before testing a winch brake, it should be ensured that the brake and the brake drum is in a
satisfactory condition. Any damage or failure should be rectified before any testing takes place. If
the brake has not been in use for some time, it should be applied slightly with the winch running
in the pay out direction, thereby cleaning the brake drum surface and the brake lining.

Section 7

The main purpose of brake testing is to verify that the brake will render at a load less than the
design rope's MBL. New ships are normally supplied with a brake test kit of the simplified type.
Each winch manufacturer will have their own test equipment and procedures which should be
followed by the operator, details of which will be contained in the instruction manual.

7.4.5.2 Frequency

Each winch brake is to be tested individually and tests are to be carried out prior to the ship's
delivery and every year thereafter in line with recommendations in the International Safety Guide
for Oil Tankers and Terminals (ISGOTT), Reference 4. In addition, individual winches are to be
tested after completion of any modification or repair involving the winch brakes, or upon any
evidence of premature brake slippage or related malfunctions. Brakes should be tested to prove
that they render at a load that is equivalent to 60% of the line's MBL.

The provision of a stopper arrangement on the tightening screw, such as placing locking nuts on
the threaded end, is NOT acceptable due to safety reasons as it will impede the brake setting
and reduce the brake holding load. See Figure 7.7 below.

FIGURE 7.7: IMPROPER FITTING OF LOCKING NUTS TO BRAKE TIGHTENING SCREW

7.4.5.3 Test Specification

For each ship a winch test specification is prepared incorporating specific instructions for setting
up the test gear, preparation of the winch for testing, setting of the winch brakes, application of
the test load, revision of torque wrench or hydraulic pressure readings if required, and recording
of test results. It should be noted that the coefficient of friction of the asbestos brake lining is
considerably affected by moisture. Newer types of asbestos-free brake linings are more stable
and are not so sensitive to moisture. To assure constant results the winch should be operated for
a short period with the brake set slightly on to dry the brake surface.

7.4.5.4 Supervision of Testing

All winch testing is to be carried out under the supervision or in the presence of a senior officer
designated by the Master or Chief Engineer or repair superintendent familiar with the test
procedure and the operation of the winches.

7.4.5.5 Test Equipment

Typical equipment for testing the brakes is shown in Figure 7.8 , and includes the following items:

• Lever consisting usually of two pieces of bar as shown on the sketch. The lever is

secured to the drum of the winch by means of bolts furnished with the test kit and fitted
through holes provided in the drum flange;

• Hydraulic jack with pressure gauge; and

• Foundation to be placed under the hydraulic jack for the purpose of distributing the load

into the deck structure.

Section 7

It is recommended that a complete set of test equipment is placed on board each ship, properly
stowed in an appropriate location. Alternatively, the Owner may elect to procure one or two sets
of testing equipment for each type and size of winch and retain this equipment in a convenient
central location for shipment to repair facilities as required.

FIGURE 7.8 : TYPICAL WINCH BRAKE TEST EQUIPMENT

7.4.5.6 Method of Testing

The testing arms are bolted to the flange of the winch drum with the hydraulic jack pressed under
the end of the arms at the designated location and resting on supports. The flange brake is set
as recommended in the test specification. If the winches are set manually, a torque wrench
should be used. If they are set hydraulically, the pressure gauge should first be calibrated.

Before testing, the detailed instructions for testing included in the test specification should be
reviewed and the equipment prepared accordingly. The instructions will include:

• The values for torque wrench or pressure gauge fitted for setting up the brakes;

• A curve or table relating hydraulic jack test pressure to line pull; and

• Hydraulic jack pressure at which the brake is designed to render.

With the winch prepared for testing, the testing gear securely in place and winch brakes set in
accordance with the recommendations, pressure is applied to the hydraulic jack. The winch drum
is to be carefully observed. At the first sign of movement, the hydraulic pressure applied to the
jack is recorded and the following action taken:

• If slippage occurs at a pressure less than designed, the brake should be tightened or

Section 7

repaired and jack pressure reapplied;

• If the recorded pressure corresponds to the design pressure the jack should be

released and the test gear removed; or

• If slippage does not occur at the design pressure, the brake setting should be

adjusted so the brake can render at the design load.

The lever should be lightweight for easy handling. Testing can be further simplified by reducing
the lever to slightly more than the drum flange radius and placing the jack directly on the winch
foundation. A graphic depicting such a test arrangement is shown in Fig. 7.9 . In place of the
heavy lever, a simple fitting attached to the drum flange suffices. The higher jack loads may pose
a problem for existing equipment, but provision for this method can easily be incorporated into
new equipment.

Once the brakes are tested and calibrated, the proper setting should be recorded. In case of
conventional screw brakes, a tag should be attached stating the proper torque. For spring-
applied brakes, the spring compression distance should be recorded and the spring adjustment
mechanism secured with a seal.

FIGURE 7.9 : SIMPLIFIED BRAKE TEST KIT

7.4.6 Brake Holding Capacity

The primary brake should be set to hold 60% of the mooring line's MBL. Since brakes may
deteriorate in service, it is recommended that new equipment be designed to hold 80% of the
line's MBL, but have the capability to be adjusted down to 60% of the line's MBL. A drum brake
holding capacity of 80% MBL is also required by Lloyds, DNV and ISO Standard 3730 with the
rope on the first layer. If a brake of an undivided drum is set to hold 80% MBL on the first layer, it
will hold approximately 65% MBL on the third layer.

When testing spring-applied brakes, it is important to bear in mind that, in service, the brake
band will rotate some quarter to half a turn beyond the initial slipping point experienced when
tested before full holding load is achieved. Thus, for spring-applied brakes, the final holding load
will be higher than the initial slipping load recorded. This load could be some 10 – 20% above
the test load due to a combination of the servo effect in the brake and the spring force acting on
the brake screw, which compensate for the elongation in the brake band and mechanism.

Section 7

Should a spring-applied brake be tested to 80% of the line's MBL, the final brake holding load
could be 90 – 100% of the MBL. It is therefore important that spring-applied brakes are tested
and adjusted to render at no more than 60% MBL, resulting in a final brake holding load of about
70 – 80% of the MBL, as illustrated in Figure 7.10 below.

FIGURE 7.10: EFFECT OF SLIPPAGE ON FINAL BRAKE HOLDING LOAD –

SPRING-APPLIED BRAKES

'80%' to change to '60%'. '90–100%' to change to '70 – 80%'

7.5 WINCH PERFORMANCE

The principle performance particulars for different size mooring winches should correspond to
values listed in Table 7.1, noting the further explanations in 7.5.1. These are in general
agreement with ISO Standard 3730, except for minimum light line speed and in-service setting of
brake.

7.5.1 Rated Pull (also called drum load or hauling load)

This is the pull that the mooring line can develop at the rated speed on the first layer. The listed
values should not be less than 22% and not more than 33% of the design line’s minimum
breaking strength. This value assures adequate force to heave in against environmental forces.
On the other hand, it is low enough to prevent line overstressing in the stalled condition,
considering that the stall pull is generally higher than the rated pull for most drive types.

7.5.2 Rated Speed (also called nominal speed or design speed)

The rated speed is the speed that can be maintained with the rated load applied to the mooring
line. The rated speed in combination with the rated pull determines the power requirement for the
winch drive. The recommended rated speed is higher for the smaller size winches because the
smaller sizes are intended for smaller ships and these warp and moor more quickly than large
ships.

7.5.3 Light-Line Speed (also called no-load speed or slack-line speed)

This is the speed that can be achieved by the winch with negligible load on the line. High speed
is essential to pass a line quickly to a shore mooring point or to bring the line quickly back on
board. The listed value recognises that during this phase the line is normally not on the first
layer, even on split-drum winches, and the actual line speed could be up to 50% higher. The

Brake Holding
Load

Drum
Rotation

Slipping point
with brake test kit

Final brake
holding load

80%

90-100%

~ ¼

~ ½

Brake Holding
Load

Drum
Rotation

Slipping point
with brake test kit

Final brake
holding load

80%

90-100%

~ ¼

~ ½

Section 7

slack line speeds achievable are different for different types of drives and this should also be
considered when specifying performance and drive type for new equipment. For example, a low
pressure two-speed vane motor can achieve only twice the rated speed.

7.5.4 Stall Heaving Capacity (also called stall pull)

This is the line pull the winch will exert when the control is in heave and the line is held stationary.
A high stall heave capacity is desirable to winch a ship onto the pier against high environmental
loads. On the other hand, the stall pull should not be so high that there is any danger of line
breakage, and should never exceed 50% of the mooring lines MBL. Achievable stall pull depends
on the drive type and control. As a rule of thumb the following ratios of stall pull to rated pull may
be assumed:

Hydraulic

1.05

Electric

1.25

Steam-driven 1.50

7.5.5 Drum Capacity

The drum should be capable of stowing the total line length. Table 7.1 lists the drum capacities
as specified in ISO Standard 3730. Two capacities, 'normal' and high' are indicated. Winches
with undivided drums could be more suitable for the 'normal' capacity since this would reduce the
number of layers required. ISO Standard 3730 also specifics that for undivided drums the total
number of layers shall not exceed five for normal capacity drums and eight for high-capacity
drums when the total hawser length is stowed on the drum. For split drums, there is no limitation
on the number of layers on the storage section. The ISO standard notes the possibility of line
damage if large loads are applied while more than four layers of rope are reeled on the drum.
This applies to wire rope, since the standard also warns of a short rope life if synthetic rope
under tension is wound in more than one layer.

On split-drum winches, it should be possible to store the total length of the mooring line on the
stowage part of the drum. This would facilitate the rapid payout of a line during mooring without
the need to transfer the line through the dividing flange slot. An allowance for 1.5 extra layers
should be made when determining the flange diameter of 'normal' capacity drums storing wire
ropes.

Section 7

TABLE 7.1: PERFORMANCE SPECIFICATION FOR MOORING WINCHES

BRAKE HOLDING

FORCE *

DRUM CAPACITY ***

RATED
WINCH

SIZE

RATED

PULL*

RATED
SPEED

MIN *

LIGHT-

LINE

SPEED

MIN **

DESIGN

ROPE

DIA.

(WIRE

ROPE)

MINIMUM

BREAKING

STRENGTH

OF ROPE

STALL

PULL

MAX.

DESIGN

VALUE

NEW

(80%)

IN-

SERVICE
SETTING

(60%)

MINIMUM

DIA. OF

DRUM

Normal

High

WIDTH

TENSION
SECTION

****

See 7.5.1

See 7.5.2

See 7.5.3

ISO 3730

ISO 2408

See 7.5.4

See 7.4.6

See 6.2.4

See 7.5.4

See 7.2

(t)

kN (t)

m/min

kN (t)

125 (12)

426 (43)

213 (21)

341 (35)

256 (26)

416

200

400

260

160 (16)

646 (66)

323 (33)

517 (53)

388 (40)

512

250

500

320

200 (20)

9.6

817 (83)

408 (41)

654 (67)

490 (50)

576

250

500

360

250 (25)

9.6

1010 (103)

505 (51)

808 (83)

606 (62)

640

250

500

400

320 (32)

7.8

1220 (124)

610 (62) 976 (100) 732 (75)

704

250

500

440

400 (40)

7.8

1450 (148)

725 (74) 1160 (119) 870 (89)

768

250

500

480

* With rope on first layer

** This is the desirable light line speed in the first layer and is in excess of the 0.5 m/s requirement of ISO Standard 3730. However, achievable light line

speed depends on the winch drive as mentioned in 7.4 and 7.6.3, and should be clarified with the winch manufacturer.

*** For split drums, this applies to the storage part of the drum.

**** Applies to split drums only

Note: Performance data also applies to winches designed for fibre ropes, but drum dimensions must be increased to suit the larger diameter of the fibre lines.

Section 7

7.6 STRENGTH REQUIREMENTS

The strength of the mooring winch structure should be based on the breaking strength of the line
and the criteria given in Section 4. The winch should therefore be clearly marked with the range
of rope strengths for which it is designed.

Although brakes should slip before the line breaks, under extreme conditions such as over-
tightening of the brake or overriding rope turns, the winch may be subjected to the full MBL of the
line. For this reason, it is recommended that all structural components of the winch including
brakes, foundations and supporting deck structure should be designed in accordance with
Section 4.4.10.

Where a drive input brake is intended to be the primary brake, all components, including the
reduction gears, must be based on the full MBL of the mooring line. For winches with undivided
drums, the line must be assumed to be at the layer that produces the highest stress in each
individual component. For split-drum winches the line shall be assumed to be on the first layer.

Ship designers should pay special attention to the support structure in way of the fixed end of
band brakes, as this is the major load transfer point. (See Section 5 for further details.)

7.7 WINCH TESTING

In addition to brake testing per 7.4.4, the following acceptance tests are recommended. The
tests are in general agreement with ISO Standard 3730.

7.7.1 Rules concerning testing at manufacturer's facility for the acceptance of the manufacturer

and purchaser:

• Type testing

One winch of each batch must be tested. This test may be replaced by a prototype test
certificate if agreed by the manufacturer and purchaser.

The test shall be carried out as follows:

(1) operation under load: alternately hauling and rendering at the rated load of the winch for

30 minutes continuously.

(2) holding test: to be tested by applying the brake holding load to a rope led off the drum,

with no rotation of the drum; this may be carried out on board ship if agreed between
purchaser and manufacturer.

(3) automatic brake system test: this test may be carried out on board ship if agreed

between purchaser and manufacturer.

(4) throughout testing, the following should be checked:

(a) tightness against oil leakages

(b) temperature of bearings

(d) power consumption

(e) speed of rotation of the drum.

Where tests are required in excess of the type test, they should be agreed between the
purchaser and manufacturer at the time of the contract.

• Individual tests

For hydraulic systems, care should be taken to ensure that there is no risk of contamination

Section 7

and lines and systems should be flushed and cleaned prior to commissioning. The standard of
cleanliness should meet the requirements of ISO 4406.

The following tests should be carried out:

(1) operation under no load: running for 30 minutes, 15 minutes continuously in each

direction, at light line speed.

(2) correct operation of brake system.

(3) throughout testing, the following should be checked:

(a) tightness against oil leakages

(b) temperature of bearings

(d) power consumption

(e) speed of rotation of the drum.

7.7.2 On-board acceptance tests and inspections

It is recommended that the following inspections and tests be carried out on board the ship using
ship's power.

• Running tests

The winch is to be run for ten minutes at light-line speed, five minutes continuously in each
direction. Bearing temperature rises must be checked.

7.8 SUMMARY OF RECOMMENDATIONS

The following is a summary of the recommendations contained in Section 7.

7.8.1 Recommendations for Ship Designers

• All winches should be controlled in the manual mode, including automatic-tensioning

winches as presently designed, see 7.1.1.

• In selecting drive systems, the speed-pull relationship for various types should be

considered, see 7.3.

• Minimum rated winch pull should be not less than 22% and not more than 33% of the

design line’s MBL, see 7.5.1.

• Minimum rated speed should be selected on the basis of ship size, from minimum of

7.8 m/min (0.13m/sec) for large ships and minimum of 12 m/min (0.20m/sec) for small
ones, see 7.5.2 and Table 7.1.

• Minimum light line speed should preferably be 45 m/min (0.75m/sec), but will be

dependent on the chosen drive system, see 7.5.3 and Table 7.1.

• Maximum stall pull should be 50% of the design rope's MBL, see 7.5.4.

• New winch brakes should be designed for holding 80% of the design rope's MBL,.

Operating load is to be set to 60% of the design rope's MBL, see 7.4.6.

• Stopper arrangements on the tightening screw, like locking nuts on the threaded end,

are NOT acceptable due to safety reasons, see 7.4.5.2.

• Winch drums should have a minimum diameter of 16 times the design rope's

Section 7

diameter, see 6.2.4.

• The tension section of a split drum should be wide enough for ten turns of the design

wire or unjacketed high modulus rope. For other fibre rope types, a minimum of 5 or 6
turns, see 7.2 and 7.2.1.

• Winch brakes should be of the spring applied type with brake setting indicator, see

7.4.2.5. Manually operated brakes should have the provision for correct setting and
re-setting, see 7.4.5.

• To improve consistency of brake performance, stainless steel should be considered

for brake drum surfaces, see 7.4.2.4.

• Winch brake testing provisions should be incorporated into the winch design, see

7.4.5.

• The strength of all structural winch components, including drum brakes and deck

support, should be based on the design rope's MBL. See Section 4.4.

• Winches should be clearly marked with the rope MBL for which they are designed.

7.8.2 Recommendations for Ship Operators

• The instructions for the particular mooring winches should be followed carefully. Items of

equipment which appear similar may require very different operation.

• Automatic tension mooring winches should only be used in the manual mode, see 7.1.1.

• Winch brakes should be tested annually to hold 60% of the MBL of the mooring lines.

After testing , a tag with proper setting values should be attached to the winch. Spring-
applied brakes should be provided with a brake setting indicator to prevent incorrect
setting, see 7.4.6.

• Mooring lines must be spooled onto the drum in the correct direction, since band brakes

are designed to work in one direction only, see 7.4.2.6. Winch drums should be marked
to show correct reeling or payout direction.

• Winches that are provided with both a drum brake and an input brake should be operated

with one brake only while the ship is moored. On multiple drum winches, this will always
be the drum brake. On single drum winches, the instruction book should be consulted to
determine which brake is the primary brake designed to hold 60% of the line's MBL.
Where the drum brake is the primary brake, the clutch between drum and shaft should be
disengaged while the ship is moored. This applies also to winches without input brakes.

• When adjusting mooring lines under high load, the winch must be placed in gear and set

to heave prior to the brake being slacked off. Failure to do this could result in damage to
winch clutch mechanisms.

Section 8

Section 8

Mooring Fittings

8.1 INTRODUCTION

Many national, shipyard and vendor standards exist for mooring fittings. However, one difficulty in
establishing the general acceptability of a particular fitting is the inconsistency in load and design
parameters between various standards. This can be partially compensated for by the continuing
adoption of ISO Standards as the basis of national or vendor specifications. Unfortunately, the
ISO Standards do not cover all mooring fittings, nor do they always provide the detail included in
some national standards. Moreover, some ISO and national standards have not been updated to
keep pace with industry progress, adding further to the difficulty in knowing which strength criteria
are relevant to a fitting. It is therefore recommended that the criteria detailed in Section 4.4 are
adopted.

Any fitting of undocumented or incomplete strength characteristics should be verified for
compliance with the strength criteria recommended in Section 4.4 by detailed calculations and
formal type-approval processes.

The following guidance highlights some of the critical elements associated with the selection and
installation of acceptable deck mooring fittings.

8.2 MOORING BITTS

ISO 3913 contains standards for bitts (double bollards) with diameters from 100mm to 800mm.

The bitts should penetrate the baseplate rather than just be welded to the top of the baseplate,
and strengthening rib plates should be fitted in the base.

The tabulated 'single rope maximum loading' quoted in the ISO Standard is the SWL when the
rope is belayed in a figure-of-eight fashion. According to the ISO Standard two ropes of this value
may be applied in figure-of-eight fashion near the base, or alternatively a single rope of twice the
load may be applied, as a loop, at heights up to 1.2 times the bitt nominal diameter.

The reason that the SWL depends on the method of rope belaying is that certain belaying
methods tend to pull the two posts together and thus induce a higher stress in each barrel than
that produced by an eye laid around a single post. With figure-of-eight belaying, the loading in
each post corresponds to the sum of all forces in the successive rope layers, which can be higher
than the maximum rope load. Experienced mariners are aware of this phenomenon and have
devised methods that effectively distribute the external load over the two posts (for instance, by
taking one or two turns around the first post before starting to belay in figure-of-eight fashion).
Fig. 8.1 illustrates the two methods of belaying a rope around bitts.

When belaying unjacketed high modulus lines around bitts, for example, when making fast a tug's
line, two turns should be taken around the leading post prior to turning the line up in a figure-of-
eight fashion.

Section 8

TABLE 8.1: MAXIMUM PERMISSIBLE ROPE LOADING OF BITTS

Note: change 'bollard' to 'bitt' in table. Delete 'bollard' in Note 1

8.3 CRUCIFORM BOLLARDS

ISO 3913, Addendum 1, covers bollards with a barrel diameter from 70mm to 400mm.

Similarly to bitts (double bollards), the tabulated 'single rope maximum loading' value quoted in
ISO for double cruciform bollards is the SWL when the rope is applied in figure-of-eight fashion.
Since a single cruciform bollard cannot be overstressed if certain belaying methods are used,
their SWL is twice that of a double bollard. No information on height of rope application is given in
the standard. However, since the scantlings are the same, it is assumed to be the same as for
bitts.

Cruciform bollards are fitted in the vicinity of tanker manifolds. These fittings typically have a SWL
suited for hose-handling operations and they may not be suitable for use for mooring applications.

Section 8

FIGURE 8.1: METHODS OF BELAYING A ROPE ON BITTS

8.4 CLOSED AND

PANAMA-TYPE CHOCKS

Although the terms 'closed chock' and 'Panama chock' are often used interchangeably, not all
closed chocks are 'Panama chocks'. Panama chocks must comply with Panama Canal

Section 8

regulations, which stipulate, among others, a minimum surface radius (178mm), a minimum
throat opening (300 x 250mm for single, 350 x 250mm for double type) and a safe working load
(32 t for single, 64 t for double type (314 kN or 628 kN)).

Closed chocks or Panama-type chocks are either deck mounted or bulwark mounted. The
strength of these fittings does not appear to be a problem due to their substantial design.
Nevertheless, the method of attachment to the deck or bulwark is important, and it is
recommended that the criteria listed in Section 4.4.4 be applied.

Some standards quote 'enlarged' type closed chocks. These are usually large throat opening
size and large radius fittings. They are especially useful with large diameter wires where the
effects of bend radius are significant. Where soft rope tails are used, the size of the throat
opening may have to be of the enlarged type to allow for connector shackles.

When closed chocks are used in conjunction with HMPE mooring lines, consideration must be
given to ensuring that contact surfaces are smooth and free from chafe points. (See Section
6.4.7.2).

Specific requirements are in place for closed chocks that are used in conjunction with the
emergency towing equipment required under Regulation ChV/15-1 of SOLAS, details of which
are provided in Section 3.4.

FIGURE 8.2: CLOSED CHOCK

8.5 ROLLER FAIRLEADS AND PEDESTAL FAIRLEADS

The rollers resemble sheaves and may be mounted near the edge of the deck to serve as
mooring fairleads or they may be mounted upon a pedestal elsewhere on the deck to provide a
fair lead to a winch drum or warping drum. Deckside roller fairleads may be of the open or closed
type. In the case of the open type the roller pin is a cantilever attached at the base only. Almost
all pedestal fairleads are of the cantilever pin type. Experience has shown that the cantilever pin
and its attachment are very critical and pin failure has been the cause of serious accidents.
Pedestal fairleads have also failed at the pedestal-to-deck connection due to improper design or
workmanship.

Roller fairleads and pedestal fairleads should be designed to meet the strength requirements
contained in Section 4.4.x and undergo a formal type approval process.

Section 8

8.6 UNIVERSAL ROLLER FAIRLEADS

Universal roller fairleads consist of several cylindrical rollers, or a combination of rollers and
curved surfaces. Possible arrangements are shown in Figs. 8.3 , 8.4 and 8.5

The basic four-roller type may have to be modified to suit extreme inboard or outboard line
angles. Sometimes the inboard lead to the winch or bitts requires an additional vertical roller, or
in rare cases, two additional vertical rollers as shown in Fig. 8.5. Extreme outboard angles can
be accommodated with chafe plates as shown in Fig. 8.4.

Some line angles are impractical and seldom occur. For example, the inboard lead seldom runs
in an upward direction and the outboard angle is upward only at terminals with a large difference
in tide or when moored in canal locks. In this case the Type C shown in Fig. 8.2 would be
adequate and result in a lower overall height of the fairlead.

Care should be taken when installing fairleads on sloping bulwarks, such as those in the bow
area, to avoid line chafing at the upper outboard edge of the frame. Type A-2 or A-3 of Fig. 8.4
would be suitable for this case.

Apart from line leads, the following considerations apply when selecting universal fairleads:

• Roller diameter. A small diameter will reduce the strength of the line (refer to Section

6.2.4). For use with wire rope with independent wire rope core, the roller diameter
should be about 12 times the rope diameter. For synthetic ropes, including HMPE,
advice should be sought from the rope manufacturer.

• Opening size (dimensions W1 and W2 in Fig. 8.3 ). The minimum size is determined

by the space required to pass the eye of the line or end fittings (such as those
required for tails) through the fairlead. The following minimum dimensions are
recommended:

Width

= 12 x d (wire and HMPE)

4 x d (conventional synthetic fibre)

Depth

= 8 x d (wire and HMPE)

2 x d (conventional synthetic fibre)

• Strength. Strength should be the main criterion. The recommended strength criteria

is shown in Section 4.4.6 and 4.4.7. Many existing designs do not comply with these
criteria. Often the frame is not strong enough to resist longitudinal forces such as
those applied to spring lines. Adequate frame strength, in the areas indicated by the
letter 'K' in Figure 8.3, should therefore be provided by the designer to ensure that
the fitting is suitable for such applications.

• Installation/material. As mentioned in Section 5, the type of frame construction can

seriously affect the ease of installation. Designs with a closed base member
effectively connecting the end posts are preferable. Also if the frame is made of steel
of higher strength than the adjoining hull structure, elaborate hull reinforcements will
be required.

All rollers should have lubrication-free bearings or bush bearings provided with grease nipples
and provisions for turning.

Section 8

FIGURE 8.3 : TYPES OF UNIVERSAL ROLLER FAIRLEADS

Section 8

FIGURE 8.4 : ADDITIONAL CHAFE PLATES FOR TYPE A FAIRLEADS

FIGURE 8.5 : UNIVERSAL FAIRLEADS WITH ADDITIONAL INBOARD ROLLERS

8.7 STOPPERS

The most common loose fitting is a stopper, which is required to temporarily hold the load of a
line while the line is in the process of being belayed on a set of bitts.

Common methods of stopping-off wire or synthetic and natural fibre ropes are shown in Fig. 8.6 .

Stopping off a line entails risk. High loads in the line must be avoided, since the stoppers have
less strength than the mooring line. Only a properly sized 'carpenter's' stopper can approach the
line's strength.

All chain stoppers should be provided with certification attesting to their construction and SWL.

8.8 SELECTION OF FITTING TYPE

At the ship design stage, a decision should be taken regarding the types of fairleads employed at
the shipside for use with mooring lines. Roller-type chocks result in less line wear and improve
the winch hauling capacity because they reduce friction between line and chock. On the other
hand, roller-type fittings require more maintenance and can be very large if the rollers are of
proper size for the intended service. A large bend radius is much easier to realise in Panama-
type chocks and for this reason they are often used in combination with winch-mounted wire
ropes.

Section 8

Winch-mounted conventional fibre lines should ideally be used with properly-maintained roller-
type fairleads, since the friction created by the fixed fairleads can lead to rapid line damage.
However, it has been shown that HMPE ropes, due to their low coefficient of friction, may not
turn the roller leads, even though they may be well maintained. In such situations, rapid line
wear may result.

FIGURE 8.6 : STOPPERS

Appendix A

Appendix A

Wind and Current Drag

Coefficients for VLCC’s and Gas

Carriers and Example Force

Calculation

A.1 INTRODUCTION AND GENERAL REMARKS

Environmental loads acting upon a ship due to wind and current should be derived from specific
model test data for that ship design or from the general non-dimensional drag force coefficients
for oil tankers and gas carriers contained in this Appendix. These coefficients were extracted
from the OCIMF publication "Prediction of Wind and Current Loads on VLCC’s", 2

Edition

(1994) and the OCIMF/SIGTTO publication "Prediction of Wind Loads on Large Liquefied Gas
Carriers", 1

edition (1985), both of which are out of print.

The wind and current drag coefficients contained in "Prediction of Wind and Current Loads on
VLCC’s" were defined originally for tankers above 150,000 tonnes deadweight. More recent
model test data on modern tanker forms confirms that these same coefficients are, in most
cases, sufficiently accurate when applied to smaller ships, and that they may therefore be used
for a range of ships down to approximately 16,000 tonnes deadweight. These drag coefficients
may also be appropriate for smaller ships provided there is geometric similarity. This means the
ratio of freeboard to hull breadth should be similar for the transverse wind drag coefficients and
the ratio of draft to hull breadth should be similar for the transverse current drag coefficients,

For gas carriers, the wind coefficients apply to membrane and spherical tank designs in the
range 75,000 m

to 125,000 m

. For gas carriers outside this range, generally acceptable wind

drag coefficient data is not available and, unless model tests are to be carried out in each case,
a conservative extrapolation from the published gas carrier data should be adopted.

Some general guidelines for adjustments if the actual ship geometry is outside the model range
are included herein. For instance, the wind tunnel oil tanker models correspond to typical pre-
MARPOL tankers with a ratio of overall length to loaded freeboard of 50 – 60, and a ratio of
ballast freeboard to full load freeboard of 3. It is suggested that wind drag coefficients for
MARPOL (SBT and double hull) tankers, which generally have higher freeboard than pre-
MARPOL tankers, be obtained by interpolation or extrapolation on the basis of the ratio of
midships freeboard to hull breadth. The transverse coefficients will generally be higher for newer
tankers in full load condition and may be higher in ballast conditions depending on actual ballast
draft and trim.

Likewise, current drag coefficients contained in "Prediction of Wind and Current Loads on
VLCC’s" are based on tankers with a length to beam ratio of 6.3 - 6.5. Some tankers have a
lower ratio and earlier studies indicate that the longitudinal current drag coefficient may be 25%
to 30% higher with a length to beam ratio of 5.0. It should be noted that, unlike the longitudinal
wind drag, the longitudinal current drag is calculated in terms of the hull length times draft.

Where more relevant drag force data is available for a specific ship, this should be used in
preference to the general data from the earlier OCIMF and SIGTTO publications. When

Appendix A

comparing such other data, it should be noted that the OCIMF and SIGTTO drag data has been
increased above the original measured mean results, to allow for scatter in the raw data, scaling
effects, and variations in hull geometry. The wind drag coefficients for VLCC’s were all increased
by 20%, except the transverse and yaw wind coefficients for ballasted VLCC’s, which were
increased by 10%. The wind drag coefficients for LNG ships were also all increased by 10%. No
increase in the measured data was made to any of the current drag coefficients.

A.2 SYMBOLS AND NOTATIONS

The following symbols and notations are used in this Appendix:

Symbol

Longitudinal (broadside) wind area

Transverse (head-on) wind area

Above water longitudinal hull area

Above water transverse hull area

B Beam

Longitudinal current drag force coefficient

Non-dimensional

Lateral current drag force coefficient

Non-dimensional

XYc

Current yaw moment coefficient

Non-dimensional

Longitudinal wind drag force coefficient

Non-dimensional

Lateral wind drag force coefficient

Non-dimensional

XYw

Wind yaw moment coefficient

Non-dimensional

Longitudinal

current

force

Newtons

YFc

, F

YAc

Force at forward and aft of the ship due to current

Newtons

YFw

, F

YAw

Force at forward and aft of the ship due to wind

Newtons

Lateral

current

force

Newtons

Longitudinal

wind

force

Newtons

Lateral

wind

force

Newtons

High above water/ground surface

Current velocity correction factor

Non-dimensional

Length between perpendiculars

Length

overall

MD Moulded

depth

XYc

Current yaw moment

Newton metres

XYw

Wind yaw moment

Newton metres

Water depth measured from water surface

T Draft

(average)

Current velocity (average)

m/s

Current velocity at depth s

m/s

Wind velocity at 10 metre elevation

m/s

Wind velocity at elevation h

m/s

WD Water

depth

Current angle of attack measured from longitudinal axis of ship degrees

Wind angle of attack

degrees

Density in water

kg/m

Density in air

kg/m

Density for salt water is taken as 1025 kg/m

and for air 1.28 kg/m

Approximate conversion factors of 10 kN = 1 tonne.f and 1 m/s = 2 knots are used.

Appendix A

A.3 WIND AND CURRENT DRAG COEFFICIENTS FOR LARGE

TANKERS

Environmental loads induced on large tankers (for the purposes of this Appendix considered to be
tankers in the 150,000 to 500,000 DWT range) by wind and current can be computed with the
procedures described in this Appendix. The forces and moments generated are suitable for a
computer analysis of the required mooring restraint. The following non-dimensional coefficients
are used throughout:

= longitudinal force coefficient

= lateral force coefficient

= yaw moment coefficient.

The subscripts w and c are used to distinguish between wind and current.

The sign convention and coordinate system adopted are illustrated in Figure A1. The sign
convention for the axis of coordinates for ships refers to a wind or current direction as 0 degrees
when it flows parallel to the hull from stern to bow. Positive angles increase in an anti-clockwise
direction.

Curves of the force and moment drag coefficients as a function of wind and current angle of
attack are provided in Figures A2 to A4 and Figures A5 to A14, respectively. The wind
coefficients are based upon data obtained from wind tunnel tests conducted at the University of
Michigan in the 1960's. The values of the current coefficients are based upon data from model
tests performed at the Maritime Research Institute Netherlands (1975-1991) and for the deep
water case ( WD/T ≥ 6) the Design Manual of the United States Navy (“Harbor and Coastal
Facilities”, NAVDOCKS DM-26, Bureau of Yards and Docks, Department of the Navy). The forces
and moments computed will be in units of Newtons and Newton metres respectively, with their
point of action at the intersection of the transverse and longitudinal centre line of the tanker.

Two bow configurations are given in Figure A15 illustrating the effect of an elliptic-cylindrical bow
versus a more conventional bulbous bow. An elliptic-cylindrical bow resembles a segment of an
elliptic-cylinder. For convenience, the more rounded elliptic-cylindrical bow is referred to as a
"cylindrical" bow and the sharper conventional bulbous bow as a "conventional" bow.

A.3.1 Major Factors Affecting Wind Loads

The coefficients are only valid for ships with superstructure at the stern. The coefficients and
areas (A

and A

) must be consistent for the particular tanker condition being investigated. The

transverse and longitudinal areas A

and A

, respectively, are measured from the head-on and

broadside projections of the above water portion of the tanker.

For wind loads on tankers with other structures adjacent, such as tandem mooring arrangements,
the calculated data should be treated with caution as the coefficients do not take into account the
effects of sheltering.

Changes in freeboard have the most significant impact on the wind coefficients. The deviations in
the coefficients result from differences in the relative force contributions of the hull and
superstructure to the final loads, and also the ratio of freeboard to breadth. The extreme cases
are represented by the fully loaded and ballasted conditions of the tanker, therefore, separate
curves for these conditions have been developed.

Variations in bow configuration also produce a substantial difference in the longitudinal force
coefficient for a ballasted tanker. The configuration changes are characterised by tankers with so-
called "conventional" bow shapes versus a "cylindrical" bow shape. The coefficients depicting the
effects of the bow shape have been plotted in Figure A2.

Appendix A

A.3.1.1

Effects of Trim

The wind drag coefficients assume the trim is zero in the fully loaded condition and 0.8 degrees in
the ballast condition. The effects of trim will be greatest on the yaw moment and when the trim
differs significantly from that assumed above, a correction to the yaw moment may be necessary.

A.3.2

Current Drag Coefficients

Curves of the current drag coefficients C

, C

, and C

XYc

, are presented in Figures A5 to A14 for

the following conditions:

• Current angle of attack: 0 degrees at the stern to 180 degrees bow on..

• Water depth to draft ratio ( WD/T):

1.10
1.20
1.50
3.00
≥ 6.00 (deep water)

• Two bow configurations

• Distinction is made between a cylindrical bow configuration without a bulbous bow and a

conventional bulbous bow configuration. For a cylindrical bow with a bulb, it is
recommended to use the data for the cylindrical bow without a bulb. For the conventional
bow shape without bulb, the larger coefficient, with or without bulb should be used.

• Loading condition (draft = 100% loaded draft and 40% loaded draft).

• Zero trim

For the later series of tests, the coefficients of the WD/T case of 1.05 were not measured. The
results for the F

, and M

XYc

coefficients for the WD/T case of 1.05 case are taken from the

original, 1977 edition of "Prediction of Wind and Current Loads on VLCC’s". In cases of WD/T
ratio of less than 1.10, it is recommended that surge force coefficients are treated with caution.

A.3.2.1

Current Velocity

An average current velocity ( V

) over the draft of the ship should be used in computing the

current forces and moment. If the vertical current velocity profile is known, the definition of the
average current velocity over the draft of the ship can be obtained from the following equation:

where:

- average current velocity (m/s)

T - draft of ship (metres)
v

- current velocity as a function of water depth s (m/s)

s - water depth measured from the surface (metres).

If the velocity profile is not known, the velocity at a known water depth should be adjusted by the
factors provided in Figure A16 to obtain the equivalent average velocity over the draft of the ship.
The example calculation in A.5 shows use of this procedure.

A.3.3

Major Factors Affecting Current Loads

Underkeel clearance has the greatest influence on the current drag coefficients. The lateral force
coefficients for a water depth to draft ratio of 1.05 are approximately three times larger than the

Appendix A

coefficients for a water depth to draft ratio equal to 3.0. This increase is primarily due to the
blockage effect of the tanker that causes a proportionately larger volume of water to pass around
rather than under the tanker as the underkeel clearance decreases.

The magnitude of the current forces and moment is also influenced by the tanker bow form in a
similar manner to the wind. Two curves are provided to represent a "conventional" versus a
"cylindrical" bow shape wherever the difference in the coefficients is significant.

A.3.3.1

Effects of L/B Variations

The test program mainly concerned L/B ratios between 6.3 and 6.5 to reflect the majority of
existing VLCC's. However, more recent VLCC's tend to have L/B ratios in the range from 5.0 to
5.5.

The drag coefficients presented in this Appendix however are still valid for these types of tankers,
with the possible exception of the longitudinal coefficients. The longitudinal drag coefficients tend
to higher values with decreasing L/B ratios.

For a VLCC with an L/B of 5.0, an increase in the longitudinal drag coefficients of maximum 25 to
30% is to be expected for the smaller current angles (up to max. 15 degrees).

A.3.3.2

Effects of Trim

The trim is assumed to be zero for all the current drag data, and the effects of trim on current
coefficients were not investigated. However the effect of trim will be most pronounced for the yaw
current coefficients for ballasted tankers in shallow water. The reader should be cautious in using
the yaw current coefficients under these conditions when the trim exceeds one degree.

Appendix A

FIGU

RE A1

SIG

ENTION

CO-

ATE

STE

Note:

effici

ents presente

d in the Section are fo

r win

and current

angle

s of attack from 0 –

180 de

gre

but

are al

so valid

for angle

s fro

181 – 359

degr

s with

the approp

ria

te sign chan

Appendix A

FIGURE A2: LONGITUDINAL WIND DRAG FORCE COEFFICIENT

Appendix A

FIGURE A3: LATERAL WIND DRAG FORCE COEFFICIENT

FIGURE A4: WIND YAW MOMENT COEFFICIENT

Appendix A

FIGURE A5: LONGITUDINAL CURRENT DRAG FORCE COEFFICIENT – LOADED TANKER

(WD/T = 1.1)

Appendix A

FIGURE A6: LONGITUDINAL CURRENT DRAG FORCE COEFFICIENT – LOADED TANKER

(WD/T = 1.2)

Appendix A

FIGURE A7: LONGITUDINAL CURRENT DRAG FORCE COEFFICIENT – LOADED TANKER

(WD/T = 1.5)

Appendix A

FIGURE A8: LONGITUDINAL CURRENT DRAG FORCE COEFFICIENT – LOADED TANKER

(WD/T = 3.0)

Appendix A

FIGURE A9: LONGITUDINAL CURRENT DRAG FORCE COEFFICIENT – LOADED TANKER

(WD/T >4.4)

Appendix A

FIGURE A10: LATERAL CURRENT DRAG FORCE COEFFICIENT – LOADED TANKER

Appendix A

FIGURE A11: CURRENT YAW MOMENT COEFFICIENT – LOADED TANKER

Appendix A

FIGURE A12: LONGITUDINAL CURRENT DRAG FORCE COEFFICIENT – BALLASTED

TANKER (40% T)

Appendix A

FIGURE A13: LATERAL CURRENT FORCE DRAG COEFFICIENT – BALLASTED TANKER

(40% T)

Appendix A

FIGURE A14: CURRENT YAW MOMENT COEFFICIENT – BALLASTED TANKER

40% T, BASED ON MIDSHIPS

Appendix A

FIGURE A15: VARIATION IN BOW CONFIGURATION

Appendix A

FIGU

RE A1

ENT VELOCIT

RRE

CTI

N FA

CTO

Appendix A

A.4 WIND AND CURRENT DRAG COEFFICIENTS FOR GAS
CARRIERS

A.4.1

Introduction

The coefficients used to compute current loads on VLCCs are also generally applicable to the
computation of current loads acting on liquefied gas carriers in the 75,000 to 125,000 m

class.

Thus, separate current coefficients were not developed for gas carriers.

Available reports dealing with the prediction of wind loads on gas carriers exhibit widely scattered
results. Various reasons for the differences can be cited; test bases are rarely similar, and the
techniques and procedures used in obtaining the loads vary considerably. Added to this, the
inherent differences in the facilities at which the test programs were conducted make attempts at
generalising existing data for design purposes a very difficult task. Additionally, the existing wind
coefficients for VLCCs were not considered applicable to gas carriers due to inherent differences
in hull shape and wind area distribution. However, based on the relative similarity of the
submerged hull geometry between these gas carriers and VLCCs with "conventional bows," it
was felt that the current coefficients could be used with more confidence.

As a result of the above, it was recommended that appropriate wind force and moment
coefficients be developed for gas carriers. Wind tunnel tests similar to those conducted for
VLCCs had been carried out in the industry. The results of these tests were obtained by OCIMF
and SIGTTO and subsequent data analysis led to the development of the wind force and moment
coefficients presented in this Appendix. The analysis was similar to that used to prepare the wind
drag coefficients for VLCCs.

Some of these drag coefficients differ from those previously derived for specific gas carriers. The
coefficients presented in this Section are generally more conservative than previously obtained
data. This level of conservatism is believed to be justified since the coefficients in this report were
developed with the intention of safely applying them to the general range of gas carrier forms,
sizes and mooring configurations. It does not necessarily follow that the previous use of less
conservative coefficients for specific ships with specific mooring configurations is invalidated by
this new data.

A.4.2

Wind drag coefficients for gas carriers 75,000 to 125,000m

Wind loads induced on moored gas carriers in the 75,000 to 125,000 m

(cubic meter) class can

be computed based on the drag coefficients described in this Section.

The following non-dimensional coefficients are used throughout in the calculation of design loads:

= Longitudinal Force Coefficient

= Lateral Force Coefficient

XYw

= Yaw Moment Coefficient

The subscript, w, is added to indicate that the coefficients are due to wind.

The wind drag coefficients are based upon data obtained from wind tunnel tests. Curves of the
wind drag coefficients C

, C

and C

XYw

are presented in Figures A17, A18 and A19. In the

original OCIMF/SIGTTO publication "Prediction of Wind Loads on Large Liquefied Gas Carriers",
the sign convention and coordinate system was different from that used for the VLCC data. The
present figures have been edited to be consistent with the VLCC

data.

The following conditions are covered:

• Wind angle of attack: 0 degrees at the stern to 180 degrees bow-on. The coefficients are

equally applicable to winds from 181 to 359 degrees with appropriate changes in the
signs of the coefficients.

Appendix A

• Any operating draft condition: the coefficients presented in this report are applicable to

draft conditions ranging from ballasted to fully loaded as the changes in freeboard are
relatively small (see Table A1). Zero trim is assumed in all cases.

• Two cargo tank types: the curves illustrate the effect of spherical tanks versus

conventional prismatic tanks. Where the effects between these two tank types are
insignificant, the curves have been combined and presented as a single curve.

While the coefficients do not vary with the ship's draft, the forward and side areas must be
consistent for the particular ship load condition and tank type under investigation.

A.4.2.1

Gas carrier wind areas

The transverse and longitudinal areas A

and A

respectively, are measured from the head-on

and broadside projections of the above water portion of the ship. Areas and principal dimensions
for a typical 75,000 m3 (prismatic tanks) gas carrier and typical 125,000 m

gas carriers

(prismatic and spherical tanks) are given in Table A1.

A.4.2.2

Major factors affecting wind loads

The presence of spherical tanks on gas carriers has the most significant impact on the wind drag
coefficients. The deviations in the coefficients result from differences in the relative force
contribution and distribution due to the configuration of the spherical tanks. Therefore, separate
curves for conventional prismatic and spherical tanks have been developed where the deviations
are significant. Differences in ship load condition are not significant due to the relatively small
change in draft from a ballasted to a fully loaded condition for the sizes of gas carriers reviewed.

Ship
Size

(1) L

B MD

Load
Condition

(2)

x (10)

Full 10.0

3.3

0.9

2.7

0.3

75,000 PR 230.0 220.0 36.0 22.4

Ballast 8.0

3.8

1.0

3.2

0.5

Full 11.0

4.5

1.2

3.5

0.6

125,000 PR 288.6 274.0 44.2 25.0

Ballast 9.0

5.0

1.3

4.1

0.7

Full 11.0

6.6

1.3

3.5

0.6

125,000 SP 288.6 274.0 44.2 25.0

Ballast 9.0

7.1

1.4

4.1

0.7

TABLE A1: PRINCIPAL DIMENSIONS/CHARACTERISTICS OF TYPICAL LIQUEFIED GAS CARRIERS

Notes: (1) Cargo tank type: PR signifies prismatic tanks; SP signifies spherical tanks

(2) Values in table represent draft conditions for fully loaded and ballasted drafts, both having a trim

angle of 0 degrees while moored.

Appendix A

FIGURE A17: LONGITUDINAL WIND DRAG FORCE COEFFICIENT – GAS CARRIER

Appendix A

FIGURE A18: LATERAL WIND DRAG FORCE COEFFICIENT – GAS CARRIER

Appendix A

FIGURE A19: WIND YAW MOMENT COEFFICIENT – GAS CARRIER

XYw

Appendix A

A.5 EXAMPLE FORCE CALCULATIONS FOR VLCC

Objective:

To compute the wind and current drag forces and moments on a 280,000 DWT tanker with a
"cylindrical" bow configuration. The tanker is fully loaded with a 22.3 metre draft and is moored in
24.5 metres of water.

(a) The wind velocity is 34 m/s measured at a 20 metre elevation. The angle of wind attack is 30

degrees off the bow i.e. at an angle of 150 degrees.

(b) The average current velocity is 1.03 m/s (2 knots). The direction of the current is 10 degrees

off the bow i.e. at an angle of 170 degrees.

The sign convention is the same as Figure A1.

Wind Load Calculations:

Step 1: Determine ship characteristics

Particulars for a 280,000 DWT tanker in a fully loaded condition.

= 3160 m

= 1130 m

= 325 m

Step 2: Obtain wind drag coefficients for θ

= 150 degrees

NOTE: "Cylindrical" bow is not a significant factor for the wind loads on a fully loaded tanker.

= -0.73 Figure A2

= 0.31 Figure A3

XYw

= -0.032 Figure A4

Step 3: The wind velocity is calculated using:

where V

= 10 metre wind velocity (m/s)

= the wind velocity at elevation h (m/s)

= elevation above ground/water surface (metres)

which leads to V

= 30.9 m/s. Substituting these values in the following equations gives:

Appendix A

Substituted:

Current Load Calculations:

Step 1: Determine ship characteristics

Particulars for 280,000 DWT tanker in fully loaded condition

Step 2: Obtain the current coefficient for θ

= 170 degrees

From the water depth draft ratio WD/T= 1.10 and the form of the bow (cylindrical) we get:

= 0.00 Figure A5

= 0.67 Figure A10

XYc

= 0.15 Figure A11

Step 3: Substituting these values in the following equations will give:

Substituted:

Appendix A

If the current velocity had been given at 30% of the draft of the ship, the following procedure
would be used to obtain the average current velocity:

Step 4: Obtain the current velocity correction factor from Figure A16 for WD/T= 1.10

K = 0.948

Step 5: Compute the average velocity

Recompute using equation from Step 3:

Summary of Wind and Current Loads

The following table summarises the wind and current loads calculated above and also the effect
of the wind and current for other conditions.

Wind loads on laden 280,000 DWT tanker for 34 m/s wind at 20 m elevation:

WIND ANGLE OF ATTACK

Full Load

180

Degrees

150

Degrees

120

Degrees

in kN

- 655

- 504

- 228

196

449

511

in kN

599

1,216

1,390 1,303 830

XYw

in kNm

-20,082

- 52,700

- 71,500 - 103,00 - 90,300

Current loads on a laden 280,000 DWT at a water depth to draft ratio = 1.10, cylindrical bow:

CURRENT ANGLE OF ATTACK

Full Load

180 Degrees

2.57 m/s

170 Degrees

1.03 m/s

90 Degrees

0.514 m/s

in kN

- 809

in kN

2,522

3,200

XYc

in kNm

192,101

25,514

The values indicated in bold are taken from the example calculation.

Appendix B

Appendix B

Rope Over-strength

B.1 INTRODUCTION

This Appendix has been prepared as background to the treatment of rope over-strength in
this 3

edition of Mooring Equipment Guidelines. It was prompted by concern about the fact

that HMPE ropes show an increase in strength above MBL during the early part of their
service life. This over-strength could typically be apparent for many months on a mooring line
designed for a service life of 10 to 15 years. This Appendix considers the implications of this
over-strength for fitting design for new ships and for fittings when wire ropes are to be
replaced by HMPE on an existing ship. For original equipment, it would be possible to
increase the size of fittings to carry the maximum strength of the rope rather than its MBL, but
for replacement ropes this is hardly an option. The guiding principle of Section 4 is that fittings
shall be able to carry rope breaking load without permanent deformation, and it is considered
essential to maintain this principle even when over-strength ropes were in use. It was thus
necessary to make an informed choice between simply accepting rope over-strength in the
existing recommendations, or amending them by adding a factor to recommend extra strength
in fittings carrying potentially over-strength ropes.

100

120

140

% MB

Years in Service

FIGURE B1 : DEPICTION OF HMPE MOORING LINE RESIDUAL STRENGTH

There are two aspects to this problem, one being the loading applied by the rope, and the
other being the strength or resistance of the fitting. Both of these are subject to statistical
variations of various types, and the loading, based as it is on rope breaking load, is also time
dependent.

It should be noted that rope manufacturers catalogue minimum breaking strengths
using different methods, for example, ISO shows MBL as unspliced strength whereas
USA manufacturer’s standards use spliced strength. In addition, some manufacturers
catalogue average minimum break strength, whereas others use 2 standard deviations
below lowest actual break strength. This can result in variations up to 20% above
catalogue strengths when new. for the purposes of these Guidelines, the ISO definition
has been adopted.

Appendix B

B.2 LOADING

The breaking of a mooring rope on a fitting is a rare event and Mooring Equipment Guidelines
has the objective of ensuring that fitting failure is a very much rarer one. A given fitting on a
given ship has a very small chance of experiencing a rope failure during the life of a ship, and
hence, because the odds against experiencing more than one rope failure on a given fitting
during a ship’s life are so high, a large safety margin against permanent deformation of a
fitting is not necessary, even though the consequences of fitting failure may be very serious.

The actual load needed to break a rope is influenced by a number of factors:

• The fact that Minimum Break Load for a new rope is just that, the minimum guaranteed

value in a population of breaking load values which has a mean value greater than the
MBL, and a standard deviation about the mean.

• Rope type. Steel wire has less variability (smaller standard deviation) than fibre rope.
• Age. The strength of all rope types eventually declines with age. The HMPE problem is

that there is an initial increase of strength of around 20% as the rope works in from new.

•

The fact that the strength in an in-line tensile test is higher than that obtained in a test
over a curved surface, where strength reductions of 15% can occur at D/d ratios of 8 or
less, and such ratios occur in a number of fitting types. Even on a winch, with a more
benign D/d ratio, strength reductions are known to occur.

B.3 RESISTANCE

The calculated ability of a fitting to withstand without permanent deformation the force exerted
on it by a breaking rope is influenced by:

• The refinement of the calculation model employed.
• The fact that the actual yield strength is larger than the specified minimum yield stress

(SMYS). The SMYS is, in fact, the minimum of a population of yield stress values which
has a mean value by definition larger than the minimum, and a standard deviation about
the mean. For steels, the variability can be quite large, particularly in thin sections and
plate.

• The use of 85% of SMYS as the limit on the stresses caused by the load on the fitting.

This concentration on linear elasticity and avoidance of yield is to overlook all the benefits
of the well-known ductility of steel. Reaching yield stress at hot spots (as long as it is not
done repeatedly) does not mean that there is no strength reserve available. For example,
if a rather thick-walled tube in bending (e.g. a bollard) reaches surface yield at a bending
moment of, say, 100 kNm, it will continue to accept increasing bending moments up to
about 130 kNm before the entire cross-section has reached yield and a “plastic hinge”
forms with deformation growing at constant moment. Even then strain hardening provides
further reserves of strength. The ratio of 1.3 rises to 1.5 for solid rectangular cross-
sections, although it is smaller for structural I -sections. In terms of bending moment, the
ratio of 1.3 to 0.85 is close to 1.5, so that the real reserve against noticeable permanent
deformation at 85% of yield is not a factor of 1.18 but one of 1.5. Values close to 1.5
apply for shear loading, although for pure tension the factor of 1.18 would be more
appropriate. Fortunately pure tension does not often occur in typical fittings.

• In passing, there is no such thing as a welded structure where yield is not present from

day one. Any deposit of weld metal is brought into tensile yield as it cools, although some
of these stresses may be shaken out by service loads on the structure.

This discussion suggests that it would not be surprising if test results for fittings designed to
the 85% of yield criterion in Section 4 showed reserves of strength of 50% or more above the
design load before permanent deformation was noted, and even greater reserves before
fracture occurred.

Appendix B

B.4 COMPARISON OF RESISTANCE AND LOADING

The remarks above show that loading and resistance are both statistical distributions, with the
loading having time-dependence as the rope ages. The objective is to minimise the overlap
between the upper tail of the loading distribution and the lower tail of the resistance
distribution. Minimise, rather than eliminate, to avoid expensive over-design of the fittings.

The worst case scenario seems to involve a HMPE rope breaking at that point in its life when
it has maximum over-strength. In round figures, breaking at 1.2 MBL, eroded by a minimum of
10% by being broken on a curved surface. The designed resistance factor is 1.18 on yield,
and up to 1.5 allowing for spread of plasticity short of permanent deformation. Even
comparing 1.2 MBL (with some reduction due to the curved surface effect) and 1.18 on yield,
there does not seem to be a real problem for an event that occurs so rarely on any given
fitting. With ductility considered, and the fact that the over-strength persists for only a fraction
of the rope service life, it seems extremely unlikely that fitting failure will ever precede rope
failure, even if HMPE ropes are used to replace original equipment of a similar rating at some
stage in the life of a ship. There appears to be a more than adequate reserve to cope with the
worst case where a stronger than expected rope is placed on a fitting whose steel is only just
up to the specified minimum yield stress.

It is also worth noting that fitting failure has apparently been a very rare occurrence in the
experience of operators in recent years, much rarer than the already rare rope failures. This
would suggest that the safety criteria of the 2

edition of Mooring Equipment Guidelines has

given good service, and can be carried forward to the 3

edition, following careful review, with

some confidence.

B.5 CONCLUSION

In the light of the discussion presented above, it was concluded that it was not necessary to
add a specific factor in the revision of Section 4 to cope with rope over-strength early in
service life.

Appendix C

Appendix C

Guidelines for Handling,

Inspection and Removal from

Service of

Wire Mooring Lines

C.1 HANDLING AND INSTALLING

C.1.1 General

The procedure for installing the rope should be planned in accordance with manufacturer's
recommendations.

The rope should be checked to verify that is not damaged when unloaded and when
transported to the storage compound or site.

C.1.2 Offloading and Storage

To avoid accidents, ropes should be off-loaded with care. The rope reels or coils should not
be dropped and care should be taken to prevent the rope being struck by a metal hook or the
fork of a fork-lift truck.

After delivery, the rope identification and condition should be checked and it should be
verified that it is in accordance with the details on the certificates and/or other relevant
documents. The rope diameter should be checked and terminations examined to ensure that
they are compatible with the equipment to which they are be fitted.

For storage, a clean, well ventilated, dry, undercover location is preferred. If outdoor storage
is unavoidable, the rope should be covered with waterproof material. The rope should be
stored in a location where it is not likely to be affected by chemical fumes, steam or other
corrosive agents and protected in such a manner that it will not be exposed to any accidental
damage.

During long periods of storage, the reel should be rotated periodically, particularly in warm
environments, to prevent the migration of the lubricant from the rope. It should be ensured
that the rope does not make any direct contact with the floor and there is a flow of air under
the reel. Consideration should be given to supporting the reel on a simple A-frame or cradle,
located on the ground.

Ropes in storage should be examined periodically and, when necessary, a suitable dressing
which is compatible with the manufacturing lubricant should be applied. The rope supplier
should be contacted for guidance on the type of dressings available and the methods of
application. After the dressing has been applied, the rope should be re-wrapped, unless it is
considered that this will be detrimental to rope preservation.

C.1.3 Certification and Marking

It should be ensured that the relevant Certificate has been obtained before putting a rope into
service and verified that the marking on the rope or its package matches the Certificate. The

Appendix C

Certificate should be retained in a safe place for identification of the rope when carrying out
subsequent periodic examinations in service.

C.1.4 Handling and Installation

The handling and installation of the rope should be carried out in accordance with a detailed
plan and should be supervised by a competent person. Personnel should wear suitable
protective clothing such as overalls, industrial gloves, helmet, eye protectors, safety footwear
and, where the emission of fumes due to heat is likely, a respirator.

C.1.4.1 Verification of Wire Rope Specification and Condition

It should be ensured that the correct rope has been supplied by checking to see that the
description on the Certificate is in accordance with that specified in the purchase order. The
rope itself should be checked by measurement to confirm that the nominal diameter is in
accordance with the nominal size stated on the Certificate. For verification purposes, the
diameter should be measured by using a suitable rope vernier gauge fitted with jaws broad
enough to cover not less than two adjacent strands. Two sets of measurements should be
taken, spaced at least 1 metre apart, ensuring that they represent the largest cross-sectional
dimension of the rope. At each point, measurements should be taken at right angles to each
other. The average of the four measurements should be within the tolerances specified in the
appropriate Standard or Specification.

The rope should be examined visually to ensure that there is no damage or signs that
deterioration may have taken place during storage or transportation to the installation site.

The rope should be carefully transferred from the storage area to the installation site.

C.1.4.2 Removing Wire Ropes from Coils and Reels

Rope coils should placed on the ground and the wire rolled out straight, ensuring that it does
not become contaminated with dust/grit, moisture or any other harmful material. If the coil is
too large to physically handle it may be placed on a 'swift' turntable and the outside end of the
rope pulled out while allowing the coil to rotate.

When the rope is on a reel, a shaft should be passed through the reel and the reel placed in a
suitable stand which allows it to rotate and be braked to avoid overrun during installation.
Where multi-layer coiling is involved, it may be necessary for the reel to be placed in
equipment which has the capability of providing a back tension in the rope as it is being
transferred from reel to drum. This is to ensure that the underlying (and subsequent) layers
are wound tightly on the drum.

The reel and stand should be positioned such that the fleet angle during installation is limited
to 1.5 degrees.

The release of the outboard end of the rope from a reel or coil should be done in a controlled
manner. On release of the bindings and servings used for packaging, the rope will want to
straighten itself from its previously bent position. Unless controlled, this could be a violent
action and it should be ensured that personnel stand clear.

The rope should be monitored carefully as it is being pulled into the system and checks
should be made to ensure that it is not obstructed by any part of the structure.

If a loop forms in the rope, it should be ensured that it does not tighten to form a kink.

This entire operation should be carried out carefully and slowly under the supervision of a
competent person.

C.1.4.3 Cutting Wire Ropes

Particular care should be taken when the rope is required to be cut and manufacturer's
instructions should be followed. Secure servings should be applied on both sides of the cut

Appendix C

mark and the length of the servings should be at least equal to two rope diameters. One
serving either side of the cut is normally sufficient for preformed ropes.

Prior to cutting, the rope should be arranged and positioned in such a manner that at the
completion of the cutting operation the rope ends will remain in position, thus avoiding any
backlash or any other undesirable movement.

The rope should be cut with a high speed abrasive disc cutter. Other suitable mechanical or
hydraulic shearing equipment may be used, although their use is not recommended when a
rope end is required to be welded or brazed.

Adequate ventilation should be ensured to avoid any build-up of fumes from the rope and its
constituent parts, including any fibre core or rope lubricants. The products used in the
manufacture of steel wire ropes for lubrication and protection present minimal hazard to the
user in the form shipped. However, the user should take reasonable care to minimise skin
and eye contact and also avoid breathing associated vapours and mists.

C.1.4.4 Securing the Ends of Wire Ropes

Any fittings, such as clamps or fixtures, should be checked to be clean and undamaged
before being used for securing rope ends.

When terminating a rope end with a wedge socket, it should be ensured that the rope tail
cannot withdraw through the socket by securing a clamp to the tail, in accordance with
manufacturer's instructions. The tail length should be a minimum of 6 rope diameters.

C.1.4.5 Transferring Wire Ropes to Winch Drums

When coiling a rope on to a plain (or smooth) barrel drum, it should be ensured that each lap
lies tightly against the preceding lap. The application of tension in the rope will greatly assist
in the effective coiling of the rope.

With plain barrel drums it is difficult to achieve satisfactory multi-layer coiling beyond three
layers.

The direction of coiling of the rope on the drum is important, particularly when using plain
barrel drums, and should be related to the direction of lay of the rope in order to induce close
coiling (Figure C1 refers).

FIGURE C1: PROPER METHOD OF LOCATING ROPE ANCHORAGE POINT ON A

PLAIN DRUM

Appendix C

When multi layer spooling has to be used, after the first layer is wound on a drum the rope
has to cross the underlying rope in order to advance across the drum in the second layer. The
points at which the turns in the upper layer cross those of the lower layer are known as the
cross-over points and the rope in these areas is susceptible to increased abrasion and
crushing. Care should be taken when installing a rope on a drum and when operating to
ensure that the rope is spooled and layered correctly.

The condition of re-usable rope end terminations should be checked before use to confirm
their size, strength and cleanliness before use. When re-using a socket, depending on its type
and dimensions, the existing cone should be pressed out or, if necessary, removed by using
heat.

It should be checked that the new rope is spooling correctly on to the drum and that no slack
or cross laps develop. As much tension as possible (2%-5% of the MBF of the rope) should
be applied to ensure tight and even spooling, especially on the first layer. As multi-Iayer
spooling is unavoidable, succeeding layers should spool evenly on the preceding layers of
rope.

C.2 INSPECTION OF WIRE ROPES

At routine intervals, the entire length of rope should be inspected by a competent person with
particular attention paid to those sections that experience has proven to be the main areas of
deterioration. Excessive wear, broken wires, distortion and corrosion are the usual signs of
deterioration. For a more detailed examination special tools are necessary which will also
facilitate internal inspection

In cases where severe rope wear takes place at one end of a wire rope, the life of the rope
may be extended by changing round the drum end with the load end, i.e. turning the rope 'end
for end' before deterioration becomes excessive.

Broken wires should be removed as they occur by bending backwards and forwards using a
pair of pliers until they break deep in the valley between two outer strands. Protective clothing
such as overalls, industrial gloves, helmet, eye protectors and safety footwear should be worn
during this operation.

The number and position in the rope of any broken wires should be recorded.

The authorised person responsible for carrying out wire rope maintenance should ensure that
the ends of the rope are secure. At the drum end this will involve checking the integrity of the
anchorage and, at the outboard end, the integrity of the termination should be checked.

Following any periodic examination or routine or special inspection where any corrective
action is taken, a record made of the defects found, the extent of the changes and the
condition of the rope should be appended to the rope's Certificate.

The following procedure should be applied if it is necessary to prepare samples from new or
used rope, for the purpose of examination and testing to destruction:

• Check that the rope end, from which the sample will be taken, is secured by welding or

brazing. If not, select the sample length further away from the rope end and prepare new
servings.

• Serve the rope, using the buried wire technique and apply a rope clamp or grip as close

to the cut mark as practically possible. Do not use solder to secure the servings.

• Ensure that the sample is kept straight throughout the whole procedure and ensure that

the minimum sample length is 1 metre unless otherwise specified.

Appendix C

• The rope should be cut with a high speed abrasive disc cutter or an oxyacetylene torch.

Weld the rope ends of the sample, after which the clamp or grip can be removed.

• The identification of the rope should be established and the sample suitably marked and

packed. It is recommended that the sample is kept straight and secured to a wood pallet
for transportation.

C.3 REMOVAL OF WIRE ROPES FROM SERVICE

C.3.1 General

The safe use of wire rope is qualified by the following criteria:

• the nature and number of broken wires;

• broken wires at the termination;

• localised grouping of wire breaks;

• the rate of increase of wire breaks;

• the fracture of strands;

• reduction of rope diameter, including that resulting from core deterioration;

• decreased elasticity;

• external and internal wear;

• external and internal corrosion;

• deformation;

• damage due to heat or electric arcing;

• rate of increase of permanent elongation.

All examinations should take into account these individual factors, recognising the particular
criteria. However, deterioration frequently results from a combination of factors, giving a
cumulative effect which should be recognised by the competent person, and which reflects
the decision to discard the rope or to allow it to remain in service.

The individual degrees of deterioration should be assessed, and expressed as a percentage
of the particular discard criteria. The cumulative degree of deterioration at any given position
is determined by adding together the individual values that are recorded at that position in
the rope. When the cumulative value (C.3.2 to C.3.8) at any position reaches 100%, the rope
should be discarded.

C.3.2 Nature and Number of Broken Wires

Broken wires usually occur at the external surface, but there is a probability that some of
broken wires will occur internally and are "non-visible" fractures.

One valley break may indicate internal deterioration, requiring closer inspection of this
section of rope. When two or more valley breaks are found in one lay length, the rope should
be considered for discard.

Particular attention should be paid to any localised area, which exhibits a dryness or
denaturing of lubrication.

If the number of visible broken wires found in a rope is more than 4 over a length of 6d, or 8
over a length of 30d, the rope should be discarded, 'd' being the nominal diameter of the rope

C.3.3 Broken Wires at Termination

Broken wires at, or adjacent to, the termination, even if few in number, are indicative of high
stresses at this position and can be caused by incorrect fitting of the termination. The cause
of this deterioration should be investigated and, where possible, the termination should be
remade, shortening the rope if sufficient length remains for further use, otherwise the rope
should be discarded.

Appendix C

C.3.4 Localised Grouping of Broken Wires

Where broken wires are very close together, constituting a localised grouping of such breaks,
the rope should be discarded. If the grouping of such breaks occurs in a length less than 6d
or is concentrated in any one strand, it may be necessary to discard the rope even if the
number of wire breaks is smaller than the maximum number referred in C.3.2 above.

C.3.5 Rate of Increase of Broken Wires

If the predominant cause of rope deterioration is fatigue, broken wires will appear after a
certain period of use, and the number of breaks will progressively increase over time.

In these cases, it is recommended that careful periodic examination and recording of the
number of broken wires be undertaken, with a view to establishing the rate of increase in the
number of breaks.

C.3.6 Fracture of Strands

If a complete strand fracture occurs, the rope should be immediately discarded.

C.3.7 Reduction of Rope Diameter Resulting from Core Deterioration

Reduction of rope diameter resulting from deterioration of the core can be caused by:

a) internal wear and wire indentation.
b) internal wear caused by friction between individual strands and wires in the rope,

particularly when it is subject to bending,

c) fracture of a steel core.

If these factors cause the actual rope diameter to decrease by 10%, the rope should be
discarded, even if no broken wires are visible.

Note: New ropes will normally have an actual diameter greater than the nominal diameter.

Low values of deterioration might not be so apparent from normal examination. However, the
condition can result in a high loss of rope strength, so any suggestion of such internal
deterioration should be verified by internal examination. If such deterioration is confirmed, the
rope should be discarded.

C.3.8 External Wear

Abrasion of the outer strands of the rope results from rubbing contact, under load, with
fittings such as chocks and fairleads.

Wear is promoted by lack of lubrication, or incorrect lubrication, and also by the presence of
dust and grit.

Wear reduces the strength of ropes by reducing the cross-sectional area of the steel strands.

If the actual rope diameter has decreased due to external wear by 7% or more of the nominal
rope diameter, the rope should be discarded, even if no wire breaks are visible.

C.3.9 Decreased Elasticity

Under certain circumstances, usually associated with the working environment, a rope can
sustain a substantial decrease in elasticity and is thus unsafe for further use.

Decreased elasticity is difficult to detect. If the examiner has any doubt, advice should be
obtained from a specialist in wire ropes. However, it is usually associated with the following:

• reduction in rope diameter;

• elongation of the rope lay length;

• lack of clearance between individual wires and between strands, caused by the

compression of the component parts against each other;

• appearance of fine, brown powder between or within the strands;

• increased stiffness.

Appendix C

While no wire breaks may be visible, the wire rope will be noticeably stiffer to handle and will
certainly have a reduction in diameter greater than that related purely to wear of individual
wires. This condition can lead to abrupt failure under dynamic loading and is sufficient
justification for immediate discard.

C.3.10 External and Internal Corrosion

C.3.10.1 General

Corrosion occurs particularly in marine and polluted industrial atmospheres. It will diminish the
breaking strength of the rope by reducing the metallic cross-sectional area, and it will
accelerate fatigue by causing surface irregularities, which lead to stress cracking. Severe
corrosion can cause decreased elasticity of the rope.

C.3.10.2 External Corrosion

Corrosion of the outer surface of the wire can be detected visually.

Wire slackness due to corrosion attack/steel loss is justification for immediate rope discard.

C.3.10.3 Internal Corrosion

This condition is more difficult to detect than the external corrosion which frequently
accompanies it, but the following indications can be recognised:

• variation in rope diameter;

• loss of clearance between the strands in the outer layer of the rope, frequently

combined with wire breaks between or within the strands.

If there is any indication of internal corrosion, the rope should be subjected to internal
examination, carried out by a competent person.

Confirmation of severe internal corrosion is justification for immediate rope discard.

C.3.11 Deformation

C.3.11.1 General

Visible distortion of the rope from its normal shape is termed "deformation" and can create a
change at the deformation position, which results in an uneven stress distribution in the rope.

C.3.11.2 Basket or Lantern Deformation

Basket or lantern deformation, also called "birdcage", is a result of a difference in length
between the rope core and the outer layer of strands. Different mechanisms can produce this
deformation.

If, for example, a rope is running onto the drum under a great fleet angle, it will touch the
flange of the drum groove first and then roll down into the bottom of the groove. This
characteristic will unlay the outer layer of strands to a greater extent than the rope core,
producing a difference in length between these rope elements. The drum will then be able to
displace the loose outer strands and bring the length difference to one location in the reeving
system where it will appear as a basket or lantern deformation.

Ropes with a basket or lantern deformation should be immediately discarded.

C.3.11.3 Core or Strand Protrusion/Distortion

This feature is a special type of basket or lantern deformation in which the rope imbalance is
indicated by protrusion of the core between the outer strands, or protrusion of an outer strand
of the rope or strand from the core.

Rope with core or strand protrusion/distortion should be immediately discarded.

C.3.11.4 Wire Protrusion

Appendix C

In wire protrusion, certain wires or groups of wires rise up in the form of loops.

Rope with wire protrusion should be immediately discarded.

C.3.11.5 Local Increase in Diameter of Rope

A local increase in rope diameter can occur and might affect a relatively long length of the
rope. This condition usually relates to a deformation of the core and consequently creates
imbalance in the outer strands, which become incorrectly oriented.

If this condition causes the actual rope diameter to increase by 5% or more, the rope should
be immediately discarded.

C.3.11.6 Kinks or Tightened Loops

A kink or tightened loop is a deformation created by a loop in the rope which has been
tightened without allowing for rotation about its axis. Imbalance of lay length occurs, which will
cause excessive wear, and in severe cases the rope will be so distorted that it will have only a
small proportion of its strength remaining.

Rope with a kink or tightened loop should be immediately discarded.

C.3.11.7 Bends

Bends are angular deformations of the rope caused by external influence.

Rope with a severe bend should be immediately discarded.

C.3.11.8 Damage due to Heat or Electric Arcing

Ropes that have been subjected to exceptional thermal effects, externally recognised by the
colours produced in the rope, should be immediately discarded.

Section Criteria

Discard

Criteria

Visible wire breaks

C.3.2

Number in length of 6d or
30d

Discard if over 4 in length 6d
or 8 over 30d

Wire breaks at
termination

C.3.3

Evidence of broken wires

Remake termination or
discard rope

Fracture of Strand

C.3.6

Strand fracture

Discard if present

Reduction of rope
diameter

C.3.7

% reduction

Discard if diameter decreased
by 10%

Abrasion of outer
wires

C.3.8

Degree of deterioration (%)

Discard if over 7%

TABLE C.1: SUMMARY OF THE MAJOR CRITERIA FOR THE INSPECTION AND

DISCARD OF WIRE ROPES

C.3.12 Removal of Rope from Service

Only qualified and experienced personnel, taking the appropriate safety precautions and
wearing the appropriate protective clothing, should be responsible for removing wire rope
from service.

Care should be taken when removing the condemned rope from drums as it may be grossly
distorted, lively and tightly coiled.

Discarded rope should be stored in a safe and secure location and should be suitably marked
to identify it as rope which has been removed from service and is not to be used again.

Appendix C

The date and reason for discard should be recorded on the rope's Certificate before it is filed
for future reference.

Attention should be paid to any Regulations affecting the safe disposal of steel wire rope.

Examples of fatigue failure of a wire rope which has been subjected to heavy loads through
under-sized fairleads.

Wire fractures at the strand or core interface
caused by failure of core support.

FIGURE C2: EXAMPLES OF ROPE DAMAGE WITH BROKEN WIRES

FIGURE C3: REDUCTION IN WIRE ROPE DIAMETER

Undamaged

Rope Section

Damaged

Rope Section

Strands bind and take

on an oval shape if the

core has failed.

Note the decrease

in lay angle when

the core fails.

Normal

Diameter

Reduced
Diameter

Appendix C

FIGURE C4: WIRE ROPE CRUSHING DAMAGE

Normal Undamaged Rope

One Rope Lay

Stretched rope shows increased lay length

FIGURE C5: ROPE STRETCH LEADING TO DECREASED ELASTICITY

Appendix C

FIGURE C6: CROSS SECTION DEPICTING SUBSTANTIAL WEAR AND SEVERE

LATERAL CORROSION

FIGURE C7: BASKET OR LANTERN DEFORMATION

A open kink in a rope caused by

improper handling.

Examples of wire ropes exhibiting severe damage resulting from service after being

kinked, leading to localised wear, distortion and misplaced wires .

FIGURE C8: AN OPEN KINK AND EXAMPLES OF DAMAGE CAUSED

Appendix D

Appendix D

Guidelines for Inspection and Removal

from Service of Fibre Ropes

D.1 INSPECTION OF ROPES

D.1.1 General

One of the most frequent, as well as the most important, questions asked about ropes is how to
visually inspect the rope in order to estimate the useful residual strength.

FIGURE D1: NEW ROPE

FIGURE D2: USED ROPE

FIGURE D3: DAMAGED ROPE

If there is no actual fibre damage or distortion, there is no positive method by which the residual
strength of a used rope can be determined visually. A laboratory analysis and tensile test is the
best way to establish residual strength.

D.1.2 Estimating Damage and Strength Degradation with Different Rope Constructions

The following guidelines are suggested for use in estimating damage and strength degradation,
brought on by normal wear:

It is important to understand that a rope will lose strength during use in any application. Ropes
are serious working tools and used properly will give consistent and reliable service. The cost of
replacing a rope is extremely limited when compared to the physical damage or personnel injury
a worn out rope can cause.

• Before inspection, identify the rope by its label or permanent marking, and consult any

previous inspection records.

Appendix D

• Visually inspect the rope over its entire length, identifying any areas requiring in-depth

investigation.

• Splice terminations should also be inspected to ensure they are in 'as-made' condition.

In synthetic fibre ropes, the amount of strength loss due to abrasion and/or flexing is directly
related to the amount of broken fibre in the rope's cross-section. After each use, look and feel
along the length of the rope inspecting for abrasion, glossy or glazed areas, inconsistent
diameter, discoloration, inconsistencies in texture and stiffness.

It is important to understand the design of the rope in use. Most ropes are designed to have
features specifically tailored to their application. These features can lead to misconceptions
during visual inspections. When a rope has a braided cover, it is only possible to visually inspect
the cover.

In laid and 8-strand rope constructions, all strands have an intermittent prominent surface
exposure, usually referred as the “crowns". Thus, they are susceptible to damage.

12-strand braided ropes are similar to the 8-strand rope mentioned above. However the “crowns”
of the strands are less prominent and therefore less susceptible to surface damage.

Double-braided rope construction has an independent inner-core, possessing approximately 50%
of the total rope strength. This core, since it is not subjected to surface abrasion and wear, tends
to retain a larger percentage of its original strength, over a longer period of time. Thus, wear on
surface strands does not constitute as large a percentage of strength loss as in other
constructions.

In a parallel strand rope construction, the core represents 100% of the rope strength. The outer
braided jacket acts as a protection against external abrasion for the strength member and
therefore massive damage to this outer braid does not dramatically reduce the overall strength of
the rope.

Ropes are also subject to internal abrasion.

D.1.3 Rope Retirement

There are so many variables that affect rope life that only a continuous process of examination by
a competent person, during and after each use, will give them the ability to retire the rope before
it reaches a critical point.

Many factors affect a rope in service and all should be taken into consideration in assessing the
remaining rope life. Factors such as load history, abrasion, bending radius and chemical attack
all need to be considered when assessing retirement criteria.

It is recommended that, in the absence of any other information, mooring ropes are replaced
when their residual strength has reduced to 75% of the original MBL This reduction can be
ascertained either by destructive testing or by visual examination, as indicated in the attached
inspection check-list. (See Section D2). Tails should be retired when their residual strength has
reduced to 60% of the original MBL.

The amount of strength loss due to abrasion and/or flexing relates to the percentage of yarns
broken in the rope's cross-section. For a conventional mooring rope, a 25% reduction will equate
to at least a 25% loss of strength in the rope. For high modulus synthetic fibre mooring lines, the
proportional loss of strength is greater than the percentage of the damage. Figure D4 provides a
graphical depiction of the relationship between rope damage and residual strength for 8 strand
polyamide/polyester and 12 strand HMPE moorings.

Appendix D

FIGURE D4: RESIDUAL STRENGTH TO ROPE DAMAGE RELATIONSHIPS

Note – 'nylon' to change to 'polyamide'

D.1.3.1 External Abrasion

When a rope is first put into service the outer filaments will quickly take on a furry appearance.
This is a normal occurrence as the surface filaments break due to slight abrasion in service. This
furry surface however acts to protect the underneath fibres in the rope construction.

This surface abrasion needs to be examined regularly to ensure what is a normal occurrence is
not mistaken for more serious damage being caused to the rope by other means. A rope left lying
in the water for instance will suffer from 'water wash', where the action of the sea works the rope
continuously under very low load, resulting in flex fatigue which also causes fibre damage and
furring. Another cause of abrasion can be from rust build up on untreated surfaces.

D.1.3.2 Internal Abrasion

Abrasion can also occur between strands and yarns in a rope and is indicative of the amount of
work the rope has done. Therefore, a rope should be opened up, where this is practical, to
inspect for internal wear. One of the signs to look for is powdered fibre which is indicative of
internal wear and will indicate a reduction in rope strength

D.1.3.3 Other Damage

Ropes can be damaged by heat and on the surface this is indicated by glazed areas where the
fibres have melted together. The strength loss can be much greater than the surface appearance
would indicate. This is the case with most conventional ropes.

With HMPE ropes, there can be another non-damaging form of glazing resulting from the
compression which typically occurs when the rope is wound on to a winch drum, around bitts or
fairleads, etc. This type of glazing will disappear when the rope is flexed.

Appendix D

Ropes should be inspected for inconsistency in diameter which can be either increases or
reductions. With ropes which have separate core and sheath constructions, inconsistency in
diameter can indicate internal damage from overloading or shock loads and can indicate that a
rope needs to be replaced.

All ropes become dirty in use but patches of discolouration along a ropes length need to be
investigated in order to determine the cause as this could indicate chemical contamination.

Localised areas of stiffness along a rope normally indicate that the rope has been subjected to
shock loads. Shock loads are simply a sudden change in tension from a state of relaxation or low
load to one of high load. The rope should be considered for retirement.

Excessive dirt or grit may have embedded in the rope causing localised stiffness. This stiffness
should not be confused with glazing as described earlier.

An occasional pulled or cut yarn will have very little detrimental effect on the strength of the rope.
However this damage is usually caused by localised external forces, which very rarely damage
only one yarn, and therefore the cumulative effect of the damage needs to be assessed by
regarding a pulled yarn as broken.

Appendix D

D.2 MOORING ROPE INSPECTION CHECK-LIST

If any one of the following criteria are observed, the rope should be discarded or the damaged
area should be cut out and the rope re-spliced.

HMPE mooring lines should not be spliced on board as they require specialist splicing ashore.

Generally, a conventional mooring line should be discarded if it has more than two splices within
its length and an HMPE mooring should be splice-free.

DISCARD CRITERIA FOR MOORING LINES

MATERIAL LOSS THROUGH ABRASION:

Construction

Conventional

Fibre

HMPE

4 strand construction

25%

15%

6 strand/7 strand construction

25%

15%

8 strand construction

25%

15%

12 strand construction

25%

15%

Double braid construction sheath

50%

n/a

Parallel strand rope and jacketed construction sheath

100% 100%

INCONSISTANT DIAMETER:

Localised reduction in diameter

Localised increases in diameter

INCONSISTENT FLEXIBILITY:

Localised areas of stiffness

HEAT FUSION:

Extended areas of heat fusion

DISCOLOURATION:

Areas caused by chemical contamination

The following figures show fibre ropes in various conditions.

Appendix D

FIGURE D5:

FIGURE D6:

SURFACE ABRASION

PLUCKED STRAND IN COVER

FIGURE D7:

FIGURE D8:

SINGLE CUT STRAND

MULTIPLE CUT STRANDS

FIGURE D9: GLAZED, NO FIBRE DAMAGE (BENT ROPE)

FIGURE D10: GLAZED, NO FIBRE DAMAGE (FLAT ROPE)

Appendix D

FIGURE D11: SAME ROPE AS IN FIGURES D9 AND D10: AFTER FLEXING NO

PERMANENT DAMAGE

FIGURE D12: ACTUAL MELTING DAMAGE, OFTEN BLACK HARDENED YARN END THAT

CAN NOT BE FLEXED BACK. IN THIS PICTURE APPROX. 50% OF ONE STRAND IS

ACTUALLY MELTED AWAY.

Appendix E

Tanker Mounted SPM Fittings

The recommendations apply to the number of mooring connections used, the safe working loads
of fittings, the dimension of chafe chains and attendant fittings and the type and location of
securing devices and fairleads on board. The recommendations only deal with those features that
are necessary to ensure the correct matching of equipment used for mooring at SPMs.

These recommendations are extracted from the fourth edition of the OCIMF publication
‘Recommendations for Equipment Employed in the Bow Mooring of Conventional Tankers at
Single Point Moorings' which should be referred to for additional information.

E.1 BOW CHAIN STOPPERS

Existing ships delivered before 2009 likely to trade to SPMs should be equipped with bow chain
stoppers designed to accept 76 mm chafe chain in accordance with the following table:

Ship Size

Number of

Bow Chain

Stoppers

Minimum

SWL

(tonnes)

150,000 tonnes DWT or less

(approx. 175,000 displacement)

Note that ship in this size range may elect to
fit two stoppers to ensure full range terminal
acceptance.

200

Over 150,000 but not greater

than 350,000 tonnes DWT

200

(approx. 175,000 – 400,000 displacement)

Over 350,000 tonnes DWT

250

(approx. 400,000 displacement)

Note: The safety factor on yield of bow chain stoppers on existing ships should be a minimum of

1.50 SWL.

New ships delivered during or after 2009 likely to visit SPMs should be equipped with bow chain
stoppers designed to accept 76 mm chafe chain in accordance with the following table. Owners
of ships under construction before 2009 are encouraged to consider fitting bow chain stoppers in
accordance with the recommendations for new ships.

Appendix E

Ship Size

Number of

Bow Chain

Stoppers

Minimum

SWL

(tonnes)

100,000 tonnes DWT or less
(approx. 120,000 displacement)

Note that ship in this size range may elect to
fit two stoppers to ensure full range terminal
acceptance.

200

Over 100,000 but not greater

than 150,000 tonnes DWT
(approx. 120,000 – 175,000 displacement)

Note that ship in this size range may elect to
fit two stoppers to ensure full range terminal
acceptance.

250

Over 150,000 tonnes DWT
(approx. 175,000 displacement)

350

Note: The safety factor on yield of bow chain stoppers on new ships should be a minimum of 2.0

SWL.

Bow chain stoppers, their foundations and associated ship supporting structure should be
demonstrated as adequate for the loads imposed. The ship should hold a copy of the
manufacturer's type approval certificate for the bow chain stopper(s) confirming that the bow
chain stopper(s) are constructed in strict compliance with a recognised standard that specifies
SWL, yield strength and safety factors. The ship should also hold a certificate attesting to the
strength of the bow chain stopper foundations and associated ship supporting structure
substantiated by detailed engineering analysis or calculations and an inspection of the
installation. An independent authority such as a Classification Society should issue both
certificates. Bow chain stoppers, associated foundations and supporting structures should be
subject to periodic survey at least once every five years and maintained in good order. Bow chain
stoppers should be permanently marked with the SWL and appropriate serial numbers so that
certificates can be easily cross-referenced to the fitted equipment.

Bow chain stopper manufacturers should provide basic operating, maintenance and inspection
instructions. Where appropriate, manufacturers should also provide guidance on maximum
component wear limits. At many terminals, common malpractice with one particular design of
bow chain stopper is the use of wedges between the retaining pin and the tongue.

Shuttle tankers with bow mooring systems designed for specialised offshore terminals may be
fitted with specialised hydraulic bow chain stoppers with interlocks and emergency shutdown
systems integral to the bow loading system. It is recommended that owners of shuttle tankers
assess the suitability of specialised bow chain stoppers for use at conventional terminals and
provide procedures and crew training to prevent accidental or inadvertent release of the chafe
chain.

These recommendations are based strictly on ship size as being the only general criteria that can
be used. Although terminal operators may change the chafe chain and equipment to take account
of local environmental conditions, it is essential that the dimensions of the material remain as
detailed in Section 5 of Reference 1 in order that it matches shipboard equipment.

In practice, bow chain stoppers have been found to be safe and easy to use and maintain. The
bow chain stopper should be designed to secure standard 76 mm diameter stud-link chain when
the chain engaging pawl or bar is in the closed position. The bow chain stopper should be
designed to freely pass a standard 76mm diameter stud-link chain and associated fittings when
the chain engaging pawl or bar is in the open position.

Appendix E

FIGURE E1: TYPICAL TONGUE-TYPE BOW CHAIN STOPPER

The following recommendations regarding positioning of the bow chain stoppers should be
complied with: -

•

Bow chain stoppers should be located between 2.7 and 3.7 metres inboard from the
bow fairlead (see Figure E2).

•

When positioning, due consideration must be given to the correct alignment of bow
chain stoppers relative to the direct lead between bow fairlead, the winch storage
drum and, if necessary, pedestal rollers (see Section E3).

•

Bow chain stopper support structures should be trimmed to compensate for any
camber and/or rake of the deck. The leading edge of the bow chain stopper base
plate should be faired to allow for the unimpeded entry of the chafe chain.

•

Improperly sited bow chain stoppers and/or ancillary equipment will hamper mooring
operations and may result in ships being rejected from terminals until the
arrangements have been modified to conform with these recommendations.

Appendix E

FIGURE E2: POSITIONING OF FORWARD FAIRLEADS, BOW CHAIN STOPPERS AND

PEDESTAL ROLLER LEADS

Appendix E

E.2 BOW FAIRLEADS

Bow fairleads should be of at least equivalent SWL to the bow chain stoppers that they serve.
The safety factor on yield of bow fairleads on existing ships delivered before 2009 should be at
least 1.50 SWL. For new ships delivered during or after 2009, it is recommended that the safety
factor on yield of bow fairleads is increased to at least 2.0 SWL. Owners of ships under
construction before 2009 are encouraged to consider upgrading the safety factor to 2.0 SWL.
The load position should be based on hawser angles up to 90 degrees from the ship's centreline
both starboard and port in the horizontal plane and to 30 degrees above and below horizontal in
the vertical plane.

Bow fairleads, their foundations and associated ship supporting structure should be
demonstrated as adequate for the loads imposed. The ship should hold a copy of the
manufacturer's type approval certificate for the bow fairleads confirming that the bow fairleads
are constructed in strict compliance with a recognised standard that specifies SWL and safety
factors. The ship should also hold a certificate attesting to the strength of the bow fairlead
foundations and associated ship supporting structure that is substantiated by detailed
engineering analysis or calculations and an inspection of the installation. An independent
authority, such as a Classification Society, should issue both certificates. Bow fairleads,
associated foundations and supporting structure should be maintained in good order and subject
to periodic survey at least once every five years.

In practice, it has been found that the use of a single central bow fairlead often creates problems
when attempting to heave the second chafe chain inboard, as the first chain tends to obstruct the
direct line of pull. A certain amount of interaction between mooring hawsers, thimbles and
hawser floats also occurs with this arrangement, frequently resulting in chafing and damage to
flotation material. As a result, two bow fairleads are recommended for ships fitted with two bow
chain stoppers.

The following recommendations refer to the size, location and type of bow fairleads (See Figure
E2):

• All bow fairleads should measure at least 600 x 450 mm.

• Bow fairleads should be spaced 2.0 metres centre-to-centre apart, if practicable, and in

no case be more than 3.0 metres apart. Ships of 150,000 tonnes DWT or less at
maximum summer draft fitted with only one bow chain stopper, need only provide one
bow fairlead, which should be on the centre line.

•

Bow fairleads should be oval or round in shape and adequately faired when fitted in
order to prevent chafe chains from fouling on the lower lip when heaving inboard or
letting go. Square bow fairleads are not suitable.

E.3 POSITION OF PEDESTAL ROLLERS AND

WINCH STORAGE DRUMS

See Figure E2 for drawings of the various arrangements described in this section.

Existing ships delivered before 2009 likely to visit SPMs should be equipped to safely handle
pick-up ropes taking into consideration safety and protection from risk of whiplash injury to
mooring personnel. It is recognised that existing ship mooring arrangements will normally require
the use of pedestal rollers to achieve the desired lead through the bow fairlead and bow chain
stopper to the winch storage drum. It is essential that pedestal roller(s) are correctly positioned,
relative to the winch storage drum and the centre of the bow chain stopper to enable a direct
lead from the centre of the bow fairlead to the centre of the bow chain stopper while allowing the
pick-up rope to be stowed evenly on the storage drum. There should be at least 3 metres
distance between the bow chain stopper and the closest pedestal roller to allow for the pick-up

Appendix E

rope eye, connecting shackle, shipboard-end oblong plate and a number of chafe chain links.

Owners of existing ships and ships under construction before 2009 are encouraged to consider
the practicality of adopting the recommendations for new ships.

New ships delivered during or after 2009 likely to visit SPMs, wherever possible, it is
recommended that winch storage drums used to recover the pick-up ropes are positioned in a
direct straight lead with the bow fairlead and bow chain stopper without the use of pedestal
rollers. This relative positioning of the tanker SPM mooring equipment in a direct straight lead is
considered the safest and most efficient arrangement for handling the pick-up ropes. However,
recognising that not all new mooring arrangement designs will permit direct straight leads to a
winch storage drum, consideration of safety and protection from risk of whiplash injury to
mooring personnel should take priority in determining number and positioning of pedestal rollers.
In addition to the pedestal roller arrangement recommendations for existing ships, the number of
pedestal rollers used for each bow chain stopper should not exceed two and the angle of change
of direction of the pick-up rope lead should be minimal.

Remote operated winch storage drums may afford some additional whiplash injury protection for
the winch operator.

Winch stowage drums used to stow the pick-up rope for existing and new ships should be
capable of lifting at least 15 tonnes and be of sufficient size to accommodate 150 metres of 80
mm diameter rope. Use of winch drum ends (warping ends) to handle pick-up ropes is considered
unsafe and should be avoided.

Appendix F

Appendix F

Strength of Chain Tensioned

over a Curved Surface

The chafing chain at the end of a tow line or SPM hawser normally is guided through a chock at
the edge of the tanker deck. This chock restrains the chain from excessive lateral movement and
provides a curved surface for the chain to bear against. The diameter of this curved surface must
be large enough to prevent overstressing of the chain when it is subjected to high towing or
mooring loads. The problem of defining a criteria for the minimum diameter of chock surfaces
was addressed by both analysis and experiment in this study.

F.1 THREE CASES OF CHAIN ON A CURVED SURFACE

Analysis of the stresses in a chain tensioned over a curved surface is more complex than first
appears. Three cases of chains tensioned over curved surfaces are illustrated in Figure FI. Case
1 is that of an ungrooved surface, such as a chock. The chain links lie at angles to the surface,
alternatively lying to one side and then the other, with all links bearing against the surface. This
is the case of primary interest in this study.

The second and third cases apply to grooved surfaces, such as anchor chain windlasses and
chain sheaves. On grooved surfaces, every second chain link projects into a groove in the
surface and has its transverse axis perpendicular or upright to the surface. Intervening links lie
with their transverse axis parallel or flat on the surface. In case the groove is so shallow that the
upright link rests on the bottom of the groove and the flat link is lifted free of the surface. In this
case bending stresses are exerted on the upright links only. In case 3, the groove is so deep that
the upright links do not touch bottom, and the flat links bear against the curved surface. Only the
flat links are subjected to bending stresses in this case.

Grooved surfaces are discussed to distinguish them from the ungrooved case and to allow
discussion of other analyses and recommendations. An intermediate case can exist in which
both the flat and upright links touch the surface. In the case of chain lying over a very small
ungrooved surface with only one or several links in contact, alternating links may lie flat and
upright to the surface as in the second and third cases.

F.2 DEFINITION OF ANGLES AND OTHER TERMS

Dimension and angles of chain links lying on a curved surface are shown in Figure FI. The chain
diameter, d, is the nominal diameter of the bar from which the common stud link is formed. For
common stud links the overall link length L is 6 times d and the link width B is 3.6 times d.

The diameter of the curved surface over which the chain is bent will be defined as D. This should
not be confused with the diameter of the opening in the bow chock. For case 2, the shallow-
grooved surface with contact at the bottom of the groove, D, is analysed as the diameter to the
bottom of the groove.

The ratio of bending surface diameter to chain diameter D/d will be used as a criteria for surface
diameter. This diameter ratio should not be confused with the ratio of the surface radius to chain
diameter which is sometimes referred to in recommendations made by others.

Appendix F

FIGURE F1: THREE CASES OF A CHAIN BENT OVER A CURVED SURFACE

Certain angles, shown in Figure F2, are of interest in analysing chain tensioned over a curved
surface. The angle between the centre lines of two adjacent chain links will be defined as

α. This

is also the angle subtended by rays from the centre of surface curvature to the centres of
adjacent links. The angle between the transverse axis of a chain link and the surface will be
defined as ß. The angle α can be found as a function of the diameter ratio D/d and the angle ß by
the following equation:

Appendix F

FIGURE F2: GEOMETRY OF A CHAIN BENT OVER A CURVED SURFACE

Unfortunately, the angle ß cannot be determined by a closed-form analytical solution. The
problem is a very complex solid geometry problem involving the interlocking of two toroids
intersecting at an angle which depends on angles α and ß. As α increases then the angle ß
decreases. Figure F3 shows the approximate relationship between the angle ß and the angle α.
These values were measured with small chain having proportions similar to large stud link chain.
The angle ß decreases slowly from approximately 40° to approximately 30° as α increases to
about 70°, and then it drops rapidly as α

increases to 90 °

Appendix F

FIGURE F3: APPROXIMATE RELATION BETWEEN ANGLE

AND ANGLE

The angle α as a function of the diameter ratio D/d and various angles ß is shown in Figure F4.
The angle α cannot, therefore, be uniquely determined without first having determined ß, and ß
unfortunately depends on α. Thus a closed form solution is not possible.

F.3 FORCES ON THE CHAIN LINK

A free-body diagram of half a chain link, showing the forces and moments applied when bent
over a curved surface, is given in Figure F5. The effect of the stud is ignored in this analysis.

The force P is the tension in the chain. For the purpose of analysis, it is assumed the tension
force P is applied at the centreline of the chain. In the case of chain on an ungrooved surface,
the tension force P between chain links acts at an angle α/2 to the chain centreline, because
every chain link bears on the surface. This is an important distinction from the case of chain on a
grooved surface, in which every other chain link is not in contact with the surface and thus is
under tension only. The angle at which the tensile force P acts on the bent chain link in the
grooved surface case is α. Only the solution of the ungrooved surface case will be addressed
here.

The reaction force of the surface against the chain link is R. For the half chain link free body
shown, only half this reaction force applies. Because the chain lies at an angle to the surface, the

Appendix F

FIGURE F4: ANGLE

AS FUNCTION OF D/d FOR VARIOUS ANGLES

FIGURE F5: FREE BODY ALALYSIS OF HALF CHAIN LINK

Appendix F

reaction applies a moment about the centreline, which must be resisted by the interlocking chain
link. The manner in which this resisting moment is applied is complicated. For the purpose of
analysis, a resisting moment N between chain links will be assumed to act about the chain link
centreline.

The vertical component of the interlink tension is P sin α /2. This force applies a bending moment
M at the centre of the chain link. From the dimensions of the link, this moment is

Considering the chain link as a beam consisting of two cylinders, each with its centreline a
distance u from the centroid, the moment of inertia is

The maximum tensile stress in the chain is at the top centre of the link. The distance to the outer
fibre at that point is

By superposition of the tensile force which is exerted by the horizontal component of P acting on
the area of the link and the maximum tensile force which is produced by the vertical component
of P applying a moment M, the total stress at this point is

To facilitate plotting and comparison of the stress level, a non-dimensional stress parameter is
defined by multiplying the above equation by d

/ P

This factor has been plotted in Figure F6 as a function of the diameter ratio for various angles ß.
For convenience the surface diameter corresponding to 76 mm (3 in.) chain is also indicated on
the abscissa of the figure.

F.4 INCREASE IN MAXIMUM STRESS

Marsh and Thurston (Reference 8) analysed the stress distribution in a stud link chain under
straight tension using stress equations which are more sophisticated than those used here. They
measured these stresses by strain gauging chain links. The results of their analysis (Figure 3 of
Marsh and Thurston reference) show the maximum tension stress in straight tension occurs at
the outside of the link at a position about 70º from the end of the link. The non-dimensional value
of this stress, αd

is 2.

This non-dimensional stress value is plotted as a horizontal dashed line in Figure F6. It indicates

that the maximum stress due to straight tension is higher than that due to being tensioned over
the surface for surface to chain diameter ratios greater than about 7 when ß = 25º. The actual
diameter ratio at which stress due to tensioning over the surface becomes greater than that due
to tensioning over the surface becomes greater than that due to straight tension depends on the
angle ß, which is undetermined.

Appendix F

FIGURE F6: NON-DIMENSIONAL STRESS FACTOR AS A FUNCTION OF D/d FOR

VARIOUS ANGLES

F.5 COMPARISON OF GROOVED

AND NON-GROOVED CASES

The two cases of chain tensioned over a curved surface with a groove can be analysed by the

stress equations given in the preceding section by replacing α/2 by α. Also, recall that for the
case of the upright link supported at the bottom of the groove, the surface diameter D must be
defined as the radius to the bottom of the groove. Figure F7 shows the non-dimensional stress
parameter αd

plotted for these two cases, together with the non-grooved surface case for ß=25º

. P

As reported by Buckle (Reference 9), the two grooved surface cases were analysed by Lloyds
Register of Shipping. The analysis was done for stud-link chain with a finite-element technique.

Appendix F

FIGURE F7: COMPARISON OF GROOVED AND UNGROOVED SURFACE CASES

The inter-link angle α was approximately 30°. The non-dimensional stress factors, as defined
above, were approximately 2 and 4.6 for the upright link (ß = 90°) and flat link (ß = 0°) cases
respectively (Figure 12 of Buckle reference).

Comparing the Buckle results with those of the present analysis gives some indication of the
adequacy of the present calculations. For the 0° case (flat link), Buckle gives a stress only
slightly lower than the present analysis. For the 90° case (upright link) Buckle gives a stress
approximately 70% higher. This higher value may be due to the action of the stud putting a
concentrated load on the middle section of the link.

The comparison indicates the simplified analysis done here may not be as accurate as the finite
element analysis performed by Buckle. Correlation is good for the 0° case, but poor for the 90°
case. The intermediate cases of interest, i.e. ß = 25°, may be less inaccurate.

In the present study, a more accurate analysis of the case of a chain tensioned over an
ungrooved curved surface could have been obtained using finite-element methods. However, a
more exact analysis of the forces between two interlocking links would be necessary. A precise
definition of the inter-link forces is a major problem. Further analysis does not appear to be
warranted.

Appendix F

F.6 TESTS OF CHAIN TENSIONED

OVER CURVED SURFACE

An experimental programme was conducted to determine the load at which stud link chain
breaks when tensioned over a curved surface. Specimens from two samples of flash-welded
grade-2 steel stud-link chain from different manufacturers were tensioned around pins having
surface diameters ranging from 2.3 to 14 times the chain diameter. They were loaded until they
failed. Properties of the chain specimens are given in Table F .1.

TABLE F1: CHAIN TENSIONED OVER CURVED SURFACE, PROPERTIES OF CHAIN SAMPLES

A typical test set up is shown in Figure F8. Load was applied by a large hydraulic cylinder
mounted horizontally. One end of the chain specimen was attached to the cylinder head by a
detachable link.

The other end was run over the test pin, mounted with its axis horizontal between two plates
extending from a stationary frame and connected through a detachable link to a chain swivel
bolted to a strong rail in the test floor.

A total of 24 tests were conducted. The set-up and results of each test is summarised in Table
F.2. The first 17 tests were conducted with chain A. The results of these tests indicated the chain
specimens might be untypically ductile. Therefore chain B was obtained and specimens from it
were tested.

F.7 RESULTS CHAIN TESTS

When loaded in straight tension, both chain samples broke at loads significantly higher than their
rated break test loads. These straight break strengths were used as the basis for defining
percent of strength for those chains loaded on the curved surfaces. There was very good
agreement between identical tests with chain A. Thus tests were not repeated with chain B.

There was wide scatter of results depending on the geometry of the chain in relation to the
surface. When interlocking chain links were arranged to contact the surface at angles near 45º,
the chains broke at the interlocking point at loads somewhat lower than the straight break test
load. When the interlocking chain links were alternating upright and flat to the surface, as
happens with smaller surface diameters, chain strength was not significantly reduced.
Sometimes the straight chain links free of the surface broke before the bent chain links did.

Figure F9 shows the tests results. The lowest breaking load obtained for each diameter ratio
divided by the straight breaking strength is plotted as a function of the surface diameter to chain
diameter ratio. In the case of identical tests, the average of the two loads is plotted.

Appendix F

TABLE F8: TEST SET-UP. TEST 15,

= 135º, D/d = 4, 8 LINKS

Although chain A was at first thought to break at untypically high loads, because of greater than
normal ductility, chain B performed even better and appeared to be even more ductile.

It was thought that links which were deformed around the curved surface would have lower
strength when loaded in straight tension. This did not occur. In the tests in which deformed links
were loaded in straight tension, the chains broke at points other than the deformed links.

The results of the chain tests should be generally applicable over a wide range of chain sizes,
even though only 32 and 35 mm (1

~ and 1⅜ in.) chains were tested. The diameter ratio

results have

been treated as non-dimensional quantities. Percent reduction in strength as a

function of the ratio of surface diameter to chain diameter chain link proportions is independent of
the experiment scale. Also, material properties and manufacturing processes are the same for
the chains tested as for larger chains. The findings of the chain tests generally agree with those
of the analytical study. Therefore, the results of these studies should be applicable to larger
chains.

The curve in Figure F9 is believed to be a reasonable indication of the manner in which surface
diameter to chain diameter ratio affects chain strength. The results indicate that strength
reduction due to tensioning chain over a curved surface is not significant for surface diameter to
chain diameter ratios greater than about 6.

Appendix F

TABLE F2: SUMMARY OF TEST RESULTS, CHAIN TENSIONED OVER CURVED SURFACE

Appendix F

*Indicates break occurred in straight chain link which was not in contact with curved surface.

TABLE F2: (continued)

FIGURE F9: RESULTS OF TESTS OF CHAIN TENSIONED OVER CURVED SURFACE

Appendix F

F.8 RECOMMENDATIONS FOR CHOCK

SURFACE DIAMETER

Practicality must be considered in specifying a chock diameter for SPM and towing chains.
Retrofitting or modification might be necessary on many tankers. If an impractical or too large
diameter is required, many operators and owners may choose to resist or defer making
modifications.

Many recommendations have been made in the past regarding the safe diameter for surfaces on
which chains bear. Review of these past recommendations indicates that only some of them
apply to chain tensioned over an ungrooved surface, the case of interest in SPM and towing
chain chocks. The Lloyds investigation reported by Buckle, the only other known analysis,
applies only for a grooved surface. No other experimental work of chain tensioned over
ungrooved surfaces is known. Thus most past recommendations are apparently only
conservative assumptions based on experience in various dissimilar services.

In the present investigation, both analytical and experimental findings indicate that stud link
chain will not be significantly weakened provided the diameter of the surface over which it is
tensioned is not less than 7 times the diameter of the chain.

Chocks with a 600 by 450 mm (24 by 18 in.) opening and a surface diameter of 560 mm (22 in.)
are available. Such chocks are known as Mississippi fairleads, and are manufactured and
marketed by several companies. When used with the 76 mm (3 in.) chain, recommended for
SPM and towing chafing chain, the 560 mm surface curvature of such chocks provides a surface
diameter ratio of 7.4. Thus, the recommended surface diameter to chain diameter ratio of 7 is
practical.

Many tankers are already fitted with bow chocks which comply with this recommendation. Some

tankers probably have chocks with smaller surface diameters. The most common commercially
available 600 by 450 mm (24 by 18 in.) opening chock has a surface diameter of 360 mm. Such
chocks provide a surface diameter to chain diameter ratio of only 4.7 with 76 mm chain. However,
such chocks can be readily replaced with larger surface diameter chocks.

Index

Aramid fibre

characteristics 6.4.3.2

Automatic tension winch

method of operation 7.1.1

Barge mooring

requirements 3.7

Berth design

recommendations for mooring facilities 1.12.1

Bitts (Double bollard)

- Definitions
connection to deck structure 5.5
design loads, safety factors and strength 4.4.1, 8.2, Table 8.1
impact of method of belaying on load 4.4.1, 8.2
methods of belaying Table 8.1
used for harbour towing 3.6

Band brake

condition of linings 7.4.2.2
effect of applied torque on holding power 7.4.2.1, Fig 7.5
friction coefficient 7.4.2.4
general 7.4.2
holding capacity 7.4.6
sensitivity in reeling direction 7.4.2.6
spring applied 7.4.2.5, Fig 7.6, Fig 7.7
testing 7.4.5

Breasting dolphin

positioning 1.12.1

Breast line

arrangement at piers and sea islands 3.2.2
function 1.3
orientation 1.5

Canal transit

requirements 3.8

Certification and inspection

of mooring fitting support structure 1.7

Chock

- Definitions
connection to rounded gunnel 5.9
design loads, safety factors and strength 4.4.4

for harbour towing 3.6
for ship-to-ship transfer 3.9
for tug escort/pull-back use 3.4.1
for use with emergency towing arrangement 3.4
general description 8.4, Fig 8.2
Panama 8.4
securing to hull structure 5.3, 8.4

Coefficient

wind and current drag coefficients App. A,

Connecting shackle

for attaching tails 6.5.4, Fig 6.7

Conventional buoy mooring

see Multi buoy moorings

Conventional fibre mooring line

bend radius 6.3.3
characteristics of materials Table 6.1
combinations of materials 6.3.1.4
construction 6.3.2, Table 6.4
general 6.3.1
handling 6.3.4, App D
inspection and removal from service App D
load extension characteristics Fig 6.5
minimum breaking forces Table 6.2
polyamide 6.3.1.2
polyester 6.3.1.1
polypropylene 6.3.1.3
safety factors Table 6.6
stoppering-off 6.3.4
storage 6.3.4

Cow hitch

for attaching tails 6.5.4, Fig 6.8

Cruciform bollard

design loads, safety factors and strength 4.4.2, 8.3

Current

calculation of forces 2.3
force 1.2.1, Fig 1.3

Design Basis Load

description 4.1

Berth Design

recommendations for mooring facilities 1.12.1

Disc brake

general 7.4.3

Elasticity

- Definitions
effect on restraint capacity Fig 1.6
of lines 1.4
of mooring system 2.4.1.3
use of tails 6.5.1

Emergency

unberthing of ship 1.11

Emergency towing arrangement

equipment requirements 3.4

Emergency towing-off pennant

- Definitions
recommendations 3.12

Environmental data

standard environmental criteria 2.2
site-specific data 2.5

Fairlead

- Definitions
- see 'Chock', 'Pedestal fairlead' or 'Universal fairlead'

Fire wire

- see ‘Emergency towing-off pennant’

First line ashore

- Definitions
contribution to restraint capacity 1.5
description 3.2.2

Fleet angle

- Definitions
for mooring winches 3.15, Fig 3.17

Forces

acting on the ship 1.2
wind and current 1.2.1

Geometric factor

- Definitions
application 4.2, 4.4,
description 4.1

Harbour towing

equipment requirements 3.6

Head line

effectiveness 1.3, 1.5

High modulus fibre mooring line

- Definitions
aramid fibres 6.4.3.2
chafe protection 6.4.7.2
chemical resistance 6.4.5.3
coefficient of friction 6.4.6.4
construction 6.4.4, 6.4.6.2
data concerning MBL Table 6.5
elasticity 6.4.5.2, 6.4.6.3
general 6.4.1
high modulus polyethylene 6.4.3.4
inspection and removal from service App C
installation 6.4.7
liquid crystal polymer 6.4.3.2
load extension characteristics Fig 6.5
properties 6.4.2, Table 6.3
safety factors Table 6.6
strength 6.4.5.1, 6.4.6.1
trade names 6.4.3.1
use on mooring winches 6.4.7.3

High modulus polyethylene, HMPE

- Definitions
characteristics 6.4.3.4

High mooring load condition

precautions applicable 1.9

Input brake

general 7.4.4

Instrumented mooring hook

function 1.7.4

Line load measurement apparatus

- see Instrumented mooring hooks

Line tending

checks on effectiveness 1.7, 1.7.4
effect of line length on tending requirements Fig 1.8
objective 1.8.1

Line-up

of equipment and fittings 3.15

Liquid crystal polymer

characteristics 6.4.3.3

Load/deflection characteristics

of mooring lines and breasting dolphins 2.4.1.2

Mixed moorings

in similar service 1.4

Mooring

arrangements and layouts Section 3
augmentation in exceptional conditions 3.11
calculations see Mooring restraint calculations
pattern Fig 1.1, 1.3, Fig 1.4
principal objectives 3.1
principles Section 1
management by ship 1.8

Mooring boats

operating limits 1.10

Mooring fittings

basic strength philosophy 4.2
certification and inspection 5.10
design loads, safety factors and strength Section 4
recommended design criteria 4.4
selection of fitting type 8.8
standards 8.1
special installation considerations 5.9
strength testing 4.5
SWL related to line MBL 4.1

Mooring lines

analysis, examples Table 1.2, Table 1.3
calculation of forces 2.4.1.3
certificates 6.1.2
conventional fibre 6.3
effectiveness 1.3
fatigue and service life 6.4.7.5
general considerations for selection 6.1
general safety hazards 6.1.1
high modulus fibre 6.4
number, size and type 3.2.1
optimum length 1.6, 1.12.1
pre-tensioning 1.8.1
record keeping 6.1.2
safety factors Table 6.6
wire 6.2

Mooring plans and diagrams

mooring layout plans 4.6
preparation by terminal 1.7

Mooring restraint calculation

basic principles of mooring calculations 2.4.1
dynamic analysis 2.5

Mooring system

elasticity 2.4.1.3

Mooring winch

acceptance tests 7.7
brakes 7.4
design loads, safety factors and strength 4.4.x
drives 7.3
drum capacity 6.4.7.3, 7.5.5
foundations 5.2
function 7.1
marking 7.6, 7.8
strength requirements 7.6
summary of recommendations 7.8

Most probable maximum load

calculation of 2.5.1

MSC Circ. 1175

minimum requirements for mooring fittings 4.3

Multi-buoy mooring

- Definition
mooring requirements 3.5
operating limits of boats 1.10

Operating limits

mooring limits 1.7.2
terminal responsibility 1.7.1

Orientation

effect on restraint capacity Fig 1.5
general guidelines 1.5

Pedestal roller fairlead

- Definitions
connection to deck structure 5.4, Fig 5.5, Fig 5.6, Fig 5.7
design loads, safety factors and strength 4.4.x
general description 8.5

Piers and Sea Islands

mooring arrangement 3.2,

Polyamide

characteristics 6.3.1.2, Table 6.1
minimum breaking force Table 6.2

Polyester

characteristics 6.3.1.1, Table 6.1
minimum breaking force Table 6.2

Polypropylene

characteristics 6.3.1.3, Table 6.1

minimum breaking force Table 6.2

Precautions

applicable in high mooring load conditions 1.9

Pre-tensioning

of lines 1.8.1

Principles of Mooring

Section 1

Quick release hook

SWL 1.12.1

Recessed bitt

design loads, safety factors and strength 4.4.3
on gas carriers for harbour towing 3.6
structural arrangements 5.6

Record keeping

of mooring lines 6.1.2

Responsibility

terminals 1.7

Restraint requirements

calculation of 2.4
standard requirements 2.4.1.4

Rounded gunnel

connection of chocks 5.9.1

Safe working load

- Definitions
marking on fittings 1.12.2, 1.12.3, 3.4.1, 3.6, 4.4, 4.6
of mooring fittings 4.4

Safety factors

of mooring fittings 4.1

Ship design

summary of recommendations for moorings 1.12.3

Ship-to-ship transfer

mooring requirements 3.9
offtaker 3.9.1
discharge ship 3.9.2

Single Point Mooring (SPM)

- Definitions
connection of fittings to hull structure 5.7
handling of pick-up ropes 7.2.3

mooring requirements 3.3
required fittings for mooring at 3.3, App.E

Ship/shore safety checklist

checking of moorings 1.7
detailing operating limits 1.7.1
joint terminal/ship meeting 1.7.3

Shore moorings

for augmenting mooring system 3.11.1
pulley system 3.11.2

Smit bracket

- Definitions
connection to hull structure 5.7

Snap-back

of broken mooring lines 6.1.1, Fig 6.1

SOLAS

emergency towing requirements 3.4

Specified minimum yield stress

- Definitions
use for establishing stress limits of fittings Section 4

Split winch drum

brake holding capacity 7.4.1
description 7.2, Fig 7.1, Fig 7.2
drum capacity 7.5.5
operation 7.2.1
use with high modulus synthetic fibre lines 6.4.7.3, Figs 6.6

Spring lines

- Definitions
arrangement at piers and sea islands 3.2.1
function 1.3
orientation 1.5
tending 1.8.1

Standard environmental criteria

details 2.2

Static equilibrium

principle 2.4.1.1

Stern lines

- Definitions
effectiveness 1.3, 1.5

Stopper

- Definitions
carpenter's Fig 8.6
chain Fig 8.6
for synthetic fibre lines 6.3.4, Fig 8.6

Structural reinforcement

basic considerations 5.1

Synthetic tail

see 'Tails'

Tail

- Definitions
general 6.5.1
impact on elasticity of assembly 1.5, Fig 1.7
length 6.5.2
methods of connecting 6.5.4, Fig 6.7
retirement criteria 6.5.3
use for STS operations 3.9.1, 6.5.1

Terminal

layout of mooring facilities 1.12.1
marking of mooring equipment 1.7, 1.12.2
mooring system management 1.7
responsibilities 1.7, 1.7.1

Testing of winch brake

frequency 7.4.5.2
general 7.4.5.1
method of testing 7.4.5.6
simplified brake test kit 7.4.5.6, Fig 7.9
supervision of testing 7.4.5.4
test equipment 7.4.5.5, Fig 7.8, Fig 7.9
test specification 7.4.5.3

Tug push points

shell reinforcement 5.8

Tugs

fittings for tug escort duties 3.4.1
limitations on use 1.10

Underkeel Clearance

effect on current force 1.2.1, Fig 1.3

Undivided winch drum

brake holding capacity 7.4.2
description 7..2

Universal fairlead

construction 8.6, Fig 8.3, Fig 8.4, Fig 8.5
design loads, safety factors and strength 4.4.x, 4.4.x
securing to hull structure 5.3, Fig 5.1, Fig 5.2

Wave frequency vessel motions

general description 2.5

Winch brake

band brakes 7.4.2
description 7.4
disc brakes 7.4.3
effect of applied torque on holding power 7.4.2.1, Fig 7.5
holding capacity 7.4.1, 7.4.6
input brakes 7.4.4
rendering 4.2, 7.4
testing 7.4.5

Winch drive

general 7.3
electric 7.3.3
hydraulic 7.3.1
self-contained electro-hydraulic 7.3.2
steam 7.3.4

Winch drum

capacity 7.5.5
marking of reeling or pay-out direction 7.8.2
minimum drum diameter 7.2
split and undivided drums 7.2
use for stowing SPM pick-up ropes 7.2.3

Winch performance

drum capacity 7.5.5
general 7.5, Table 7.1
light-line speed 7.5.3
rated pull 7.5.1
rated speed 7.5.2
stall heaving capacity 7.5.4

Wind

calculation of wind forces 2.3, App.A
force 1.2.1, Fig 1.2

Wire mooring line

bend radius 6.2.4
construction 6.2.2, Fig 6.2
corrosion protection 6.2.3
effects of bending on strength Fig 6.3
handling, inspection and removal from service 6.2.5, App C
load extension characteristics Table 6.5
material 6.2.1
safety factors Table 6.6
standard specifications 6.2.6

Document Outline

Introduction
Contents
List of Figures
List of Tables
Glossary of Terms and Abbreviations
Bibliography
Section 1 Principles of Mooring
Section 2 Mooring Restraint and Environmental Criteria
Section 3 Mooring Arrangements and Layouts
Section 4 Design Loads, Safety Factors and Strength
Section 5 Structural Reinforcements
Section 6 Mooring Lines
Section 7 Winch Performance, Brake Holding Capacity and Strength Requirements
Section 8 Mooring Fittings
Appendix A Wind and Current Drag Coefficients ....
Appendix B Rope Over-strength
Appendix C Guidelines for Handling, Inspection and Removal from Service of Wire Mooring Lines
Appendix D Guidelines for Inspection and Removal from Service of Fibre Ropes
Appendix E Tanker Mounted SPM Fittings
Appendix F Strength of Chain Tensioned over a Curved Surface
Index

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