Tugnology.pdf

TUGNOLOGY2009 – integration of the rudderpropeller into the ship structure

Manfred Heer - SCHOTTEL, Spay page 1

CONTENTS:

1. INTRODUCTION

2. SCOPE OF ORDER SPECIFIC CALCULATIONS

2.1 scope of calculations on thruster maker side

2.2 recommended scope of calculations and information from customer side

3. THE SCHOTTEL MODULAR SHAFTING SYSTEM

3.1 components used

3.2 the software tool

4. STATE-OF-THE-ART INSTALLATIONS

4.1 principal installation methods of the thruster

4.2 installation variants

4.3 design goal for thruster foundation stiffness

4.4 traditional and modern shaft line arrangement

5. TYPICAL PROBLEMS OR “What can go wrong?”

5.1 thruster fixation – welded version

5.2 thruster fixation – bolted version with O-ring sealing

5.3 thruster fixation – bolted version with “rubber flange” installation

5.4 thruster fixation – global resonance of the thruster in the ship structure

5.5 shaft line – lubrication problems

5.6 shaft line – “Why do shaft lines vibrate?”

5.7 shaft line – unallowable displacements

5.8 shaft line – minimum required bearing load

5.9 shaft line – maximum axial bearing loads

5.10 shaft line – unallowable bearing vibrations

6. TYPICAL EXCITATIONS OF THE NOZZLED THRUSTER

6.1 main static and dynamic force components acting on the azimuthing thruster

6.2 exemplary static and dynamic forces of the 4-bladed nozzled thruster

7. INTEGRATION OF THE THRUSTER INTO THE SHIP CONSTRUCTION

7.1 calculations with detailed and reduced models

7.2 evaluation of the results

8. SUMMARY



SYNOPSIS

The continuing trend to higher powers coupled to larger rpm can cause problems with vibrations in tugs.

Higher powers are installed into smaller and more lightweight hull constructions.

At the same time the demands for smooth operation with low vibration levels come into the foreground for

reasons of comfort together with a more proven knowledge on the influence of vibrations on component lifetime.

At each point where the propulsion systems interact with the ships, hull vibratory forces will be induced into the

construction. This is inevitable and the ship's construction should be such that it can absorb these forces. On the

other hand it should be understood that the structural design of mainly the shaft line foundation and

rudderpropeller ambient structure is significantly influencing the vibration level which is lead into the tugs hull

and should therefore be subject of precise engineering not only with regard to strength but also to stiffness.

This paper describes aspects of the mechanical integration of rudderpropeller propulsion systems for tugs.

It shows modern azimuthing propulsion systems (azimuthing propellers and shafting systems) and will highlight

the advantages and dangers of different installations of shaft lines and rudderpropeller fastenings.

It will be shown what can go wrong and will present examples of shafting systems that are less sensitive to

misalignments. The paper will also provide an insight in the typical excitation of ducted azimuthing thrusters.

It will show some results of investigations into the magnitude of these excitations.

By means of modal analyses, samples will be given of good and inadequate constructions.

1. INTRODUCTION

SCHOTTEL rudderpropellers have been serving tugs for more than 40 years and with a total number of about

800 tug boats equipped with SCHOTTEL rudderpropellers the company has gained a wide experience in

supplying complete azimuthing propulsion systems consisting of fixed and controllable pitch rudderpropellers,

shaft lines, hydraulic systems for steering, clutch and brake plus complete electric control systems including e.g.

electrically driven azimuth steering, modern bridge panels and the popular SCHOTTEL Masterstick multi-unit

control system.

Over the years the tug boat has developed from a simply designed “working horse” for use mainly in harbour

areas to a highly engineered complex ship system with a large scope of functions and specialized tasks.

For many applications working hours and power density both have increased while hull sizes are often limited

by reasons of manoeuvrability or reasons of costs (pic. 1).

„JANUS“ L: 24 m B 8.50 m

2 x 360 kW BP: 12 t

„SINALDA“ L: 33 m B: 11 m

2 x 1850 kW BP: 65 t

„WESLEY“ L: 30 m B: 12m

2 x 2350 kW BP: 75 t

This higher engineering depth of the product “TUG” requires engineering support and continuous development

efforts in all essential disciplines of mechanical, hydraulic, electrical and hydrodynamic engineering.

SCHOTTEL has therefore built up specialised development teams of experienced engineers for all these 4 key

competences supplying the companies´ sales and order management organisation with the engineering support

that is needed to serve our customers needs.

As a further contribution to a high quality technical support of each individual order a team of engineers was

added to the companies´ order management department to have a 100% technical revision of the installations

that are released to our customers.

By this organisation SCHOTTEL ensures to bring the state-of-the art knowledge from the R&D departments

directly into the projecting phase of each individual order.



The typical order specific engineering support for the standard tug application includes engineering from all

disciplines that have to be combined to a round package of system competence:

o bollard pull calculation & propeller blade design (optionally CPP propeller pitch range determination)

pic. 2 – bollard pull calculation, blade design & CFD analysis of the thruster

Design & calculations are made under consideration of motor capacity, thruster-hull-interaction and space &

draft. The detailed propeller 3D-models are given to the foundry as to ensure precise propeller geometry.

o electrical design of the interfaces, switch boxes, operation panels and superior control systems

pic. 3 – standard panel, modern compact switch-box with bus control, panel layout with Masterstick

Based upon standard drop-in panel components and a modular switch box concept any arrangement of units with

its design variants can easily be adapted. The CAN-bus system with its flexible interface possibilities allows safe

data and signal exchange. The generation 3 Masterstick system offers one hand operation of the ship.



o hydraulic & pneumatic engineering for steering, clutch, brake

pic. 4 – hydraulic plan for hydraulic steering, standard hydraulic aggregate

Hydraulic plans and 3D-construction of the hydraulic aggregate give exact interfaces for the designer.

o check of strength of the mechanical fixation of the thruster to the hull

pic. 5 – critical structural height, FEM-calculation of critical installation

design of the shaft line considering:

o strength requirements of the classification society

o distance and height difference between diesel and SRP

o axial movements of the components

o stiffnesses of the foundations and bulkheads

o excitations by unbalance and cardan joints

o thermal expansion of the shaft line itself

o Torsional-Vibration-Calculation / Lateral-Vibration-Calculation / Axial-Vibration-Calculation

The graphics next page show a typical installation space for a medium sized stern tug application, an example for

a mass-elastic-system of a tug drive line and a calculation result of a modal analysis of an intermediate shaft.



pic. 6 – typical installation space for a shaft line in a stern tug

The distance between motor and thruster, the axis offset of crankshaft to thruster input shaft and eventually the

inclination of motor and/or thruster are the first main figures needed to make up a shaft line concept.

A Torsional-Vibrations Calculation is

made for each individual project.

Focus is on the dynamic torque load of

the thruster makers´ scope of supply as

a counter check of the official TVC,

which is in the responsibility of the

motor supplier. Torsional vibrations are

not handled in this paper!

It should be understood, that excessive

torsional vibrations can also excite the

shaft line, the motor or even the thruster

which will result in vibrations measured

e.g. at the shaft line foundation.

When examining lateral or axial shaft

line vibrations, the torsional vibration

system should be studied beforehand, to

prevent faulty conclusions. pic. 7 – TVC mass elastic system of a tug drive

pic. 8 – LVC vibration mode of a typical fourfold-supported intermediate shaft

Different from the TVC which requires data exchange between motor and thruster maker only, the Lateral-

Vibration-Calculation has an interface to the designer/yard through the assumptions for bearing foundation

stiffnesses that need to be made by the calculation engineer.



Finally for all the a.m. technical issues an adequate data exchange between the technical people is substantial.

To get an overview over the organisation at SCHOTTEL that is involved into the engineering decisions during

the project phase a simplified organisation chart may help to understand responsibilities and information flow.

pic. 9 – organisation chart active in the order entry phase

This paper will address the mechanical engineering and especially the mechanical integration of rudderpropeller

(SRP) and shaft line into the tug and is therefore reduced to those aspects.

The most important document which is released to the customer as a result of the project work is a detailed

installation drawing showing the mechanical components in the ship structure supplied by the customer.

With the final acceptance of this installation drawing the order specific mechanical engineering in the order entry

phase is closed and finished. This period of time should be used by both the drive suppliers and the customer to

get all relevant aspects of the installation fixed to ensure trouble-free installation, setting into service and

operation.

The following essay shall identify aspects and problems that have to be clarified within this task.

Thumb values and examples will be given to quantify the one or the other parameter.

Finally, in spite of the use of standardised components and methods every installation requires wary engineering

and clean interface commitments.

This again requires a clear and common understanding of the questions to be clarified on both sides.

This paper shall help to identify the questions that should be asked.



2. SCOPE OF ORDER SPECIFIC CALCULATIONS

2.1 scope of calculations on thruster maker side

For each order a standard scope of calculations is made to ensure proper engineering of the propulsion system.

o calculation file of the SRP product is a standard

package supplied to the classification society

o check of strength of thruster-to-hull connection

(bolting, welding, structure)

o calculation of strength of shaft line

o lateral vibration calculation of shaft line

(with assumed or better known stiffness values)

o axial vibration calculation of shaft line

(axial stiffnesses of bulkheads and fixed bearing

foundations must be supplied by customer)

o check of deflection/displacement capacity rad. & ax.

(motor movements of elastically supported motors

must be supplied by customer)

o TVC (torsional vibration calculation) as a counter-

check to the official TVC supplied by the motor

manufacturer to the classification society

This scope of calculations is managed by a team of engineers in the order management department who are

responsible for the technical clarification of the order and thus to communicate between customer (over sales

department), classification society and with the experts in the development departments on demand.

To find the best solution for each individual application a certain scope of information is required also from

customer side.

2.2 recommended scope of calculations and information from customer side

o check of structural strength or SRP foundation

(hull-sided)

o check of modal behaviour of the the SRP units

in the hull structure

(resonances with 1st

& 2nd

blade harmonics

to be avoided)

o shaft foundation stiffness to be ensured

(design to 500 kN/mm - min. 300 kN/mm)

o supply axial stiffnesses of fixed bearing

foundations and bulkhead to thruster maker

o supply motor movements (6-degree-of-freedom

calculation) to SCHOTTEL

The above list is limited to the mechanical aspects of drive integration into the tug hull.

Strength aspects are generally under control and under tight observation of the classification societies.

With regard to thruster and shaft line dynamics there is room for improvement and data and knowledge exchange

should be improved by all parties involved.



3. THE SCHOTTEL MODULAR SHAFTING SYSTEM

3.1 components used

As can be imagined by all the a.m. aspects of design and installation variants a huge portfolio of components

agglomerates over years of application. This is not only hindering smooth logistic order handling but also bears

an increased risk of running into technical problems.

It was therefore decided to develope a standardised modular shafting system with a limited number of

components with known and proven performance in the tug application.

The “construction kit” for shaft lines comprises the following elements:

A) bearings

For the loose bearing position split cylinder roller bearings and spherical bearings both of different sections are

used.

For the fixed bearing position the same types of bearing are used with axially fixation of the outer ring.

Additionally a split double taper roller bearing type (developed and tested commonly with COOPER Roller

Bearings Co Ltd.) is available in all required sizes which allows very high axial loads at a rather low minimum

bearing load.

pic. 10 – bearing types used in SCHOTTEL shaft lines

B) shafts

Solid forged steel shafts made of 42CrMo4V with a max. length of 6 m are used most in the lower speed range,

as stub shafts or for higher speed range with short shaft lengths.

Flanges are shrink fitted with conical oil pressure assembly as a standard.

Seamless welded St52-3 with welded solid shaft/flange ends made of 42CrMo4V are mainly used as

intermediate shafts, more for the higher speed range and for large installation lengths.

Distances of ab. 6..7 m are spanned with hollow shafts with only 2 bearings.

Using split bearings allows having the shaft flanges integrated into the welded end pieces.

Composite shafts are used time by time in special applications. For the tug application it must be said that a steel

shaft is clearly more robust and thin-walled hollow steel shafts offer most of the advantages of the composite

solutions at lower cost.

pic. 11 – standard shaft types used in the modular shaft line system



C) compensation couplings

To compensate movements and deflections within the shaft line system a defined variety of components is

nowadays used for the standard installations.

Cardan shafts, with the disadvantage of own excitation are used to compensate height differences between diesel

flywheel and thruster input flange.

For the favourite straight shaft applications steel link couplings with metal joints (lower displacement capacities)

or rubber joints (higher displacement capacities) are used at different positions in the shaft line.

As these couplings are radially stiff they can only compensate axial and bending deflection when used in single

arrangement.

To compensate radial deflections or to adjust small height differences between shafts they are used in double

arrangement with a piece of shaft in between to adjust the radial displacement capacity at given max. bending

angle of the coupling.

Toothed couplings aren´t used that often because they again require regular maintenance and normally only have

limited radial displacement capacity. Nevertheless they are often used to link 2 pieces of shaft together (to e.g. a

longer intermediate shaft).

The graphics below give an overview over the typical max. displacement values of the different coupling types

and the required alignment accuracies.

pic. 12 – couplings used in the modular shaft line system

The coupling data above are general values only. Specific values have to be taken for the respective size chosen.

Special advantage of the double arrangements with steel link or rubber link couplings is that they do not produce

excitations as the cardan shaft but have sufficient displacement capacities as required from the installation.

A trivial but important thing is to differ between alignment tolerances and displacement capacity in operation.

Any alignment needs to leave sufficient margins for the operation deflections that will come on top of it!



3.2 the software tool

To make the choice of the components quick and easy and to have all necessary data available in one database

SCHOTTEL uses an own software tool for the composition of the shaft line.

The so-called “shaft line generator” allows the composition of the shaft line under consideration of shaft

strength, lateral vibration safety, required bearing lifetime and adequate cardan shaft size.

The basic strategy is to use max. 3 bearings per shaft whereas the centre bearing is the fixed bearing in this case

to assure the minimum bearing load required.

pic. 13 – user interface of the shaft line generator

The user can choose the main order data, power and speed and the dimensional restrictions of the installation

space.

By adding preferences that can be made with help functions from the programme the user comes to a limited

number of useful solutions for the power transmission task.

With the data generated the further calculations like TVC and LVC can be done upon stored databases and

installation drawings can be produced for the customer.



4. STATE-OF-THE-ART INSTALLATIONS

4.1 principal installation methods of the thruster

Dependent on the maintenance concept and constructive restrictions by the tug overall design the principal

installation method can be chosen to be:

A) from above B) from below

pic. 14 – installation of the azimuth thruster from above and below

In both cases the connection to the hull has to be designed to be:

o pressure water tight

o of sufficient strength

o of sufficient stiffness

A) installation from above:

Due to the large size of the structure supplied with the thruster there is a sufficient “part of the stiffness” already

delivered with the thruster itself. Radial web connection to the hull is therefore normally not necessary.

B) installation from below:

Also for this method the thruster only needs to be connected to the hull flange via its top plate from a strength

point of view.

With respect to stiffness and consequently the dynamic behaviour of the thruster in the structure it is sometimes

useful or necessary to also connect the brackets of the thrusters support structure radially with the trunks´

cylindrical wall (see pic. 15a).

Care has to be taken to align the brackets with the hull structures´ counter webs (see pic. 15b).

pic. 15a pic. 15b



Compared to the fixation of the main flange only the bending stiffness increases approximately by ab. 10..25 %

when the webs are connected. This will result in an offset of the pendulum mode of the thruster by ab. 5..10 % to

higher frequencies. Pendulum mode means that the underwater masses of lower gear box, propeller and nozzle

as a whole swing in the stiffness of mainly the steering tube of the thruster.

4.2 installation variants

The graphics below show the standard fixation methods of the thruster to the hull structure. From the top left to

the bottom right welded connection, bolted steel flange connection with O-ring sealing and bolted “rubber

flange” connection with spacers that define the compression of the rubber flange are shown.

A formerly used method is the epoxy cast foundation which isn´t recommended any more because the other 3

variants meet the yards requirements at lower installation effort.

pic. 16 – installation variants for the thruster-hull connection

All these installation methods can be used either with or without additional bracket connection in the trunk.

As the brackets are also used to hold the bottom cover plate in are of the hull plating they are anyhow foreseen as

a standard feature of the support structure of the thruster.

4.3 design goal for thruster foundation stiffness

Finally with a specific thruster design the mass of the thruster and its mass moments of intertia especially around

x- and y-axis (pendular modes) are known.

With the thrusters´ own vertical (z-direction) and bending stiffnesses (x-, and y-direction) linked to the hull

structures´ stiffness the Eigenmodes of the thruster are already determined.

Furthermore the main excitation frequencies of the propeller (1st, 2

nd and 3

rd order of the blade number) are given

by choice of the number of blades (4 as a standard, sometimes 5 as a requirement from the TVC).

So there are not very much parameters left to tune other than the constructive stiffness of the thruster foundation

and optionally the connection of the thruster brackets to the hull.



Considering the thruster´s own mass-elastic data and its excitation frequencies some typical required values for

the thruster foundation stiffness for the power range of 1000 to 3500 kW can be found with:

cxx & cyy > 3 x 1010

[Nm/rad] cz > 3,5 x 109 [N/m]

Detailed calculations should be made to avoid global resonances to the 3 critical mode shapes with the a.m.

propeller orders to avoid needless high vibration levels in the ship or consequential damage in mainly thruster

bearings and shaft line components.

pic. 17 – Eigenmodes of the thruster – bending in x - and y-direction, pounding in z-direction

4.4 traditional and modern shaft line arrangement

For many years the tug drives with azimuthing thrusters were built-up with a rather limited number of diesel

motor types and with a tight speed range of mainly ab. 600..1000 rpm.

The typical shaft line was constructed using:

o massive forged steel shafts only

o cardan shafts in Z- or W-arrangement

o all displacements were taken by the cardan shaft splined connections

o spherical roller bearing were used as fixed and loose bearings

o standard lithium-soaped multi-purpose grease was used

o foundation stiffnesses of 500 kN/mm were commonly reached with massive foundations

o bulkhead bearings were meant to be a “no go” due to the radial stiffness deviance compared to the

other bearing foundations

Using a standard foundation stiffness and massive cylindrical

shafts which were considered as simple “bending springs” the

choice of the bearing distance (and hence number of bearings) was

a simple job that could be solved by means of the diagram on the

rigth:

pic. 18 – selection diagram for bearing distance

by shaft diametre and speed

The higher the speed and the longer the shaft line the more

bearings were used to solve the task which went well until 1600 to

1800 rpm diesels were widely used.

In extreme cases it became necessary to use an amount of up to 6

bearings in one shaft line and up to 10 bearings in the arrangement

of one azimuth drive!

This demands a very good foundation design and requires

extensive alignment which cannot always be realised in practice.



The picture below shows such a traditional arrangement which was often used also with higher rpm. With a good

foundation design and proper alignment and maintenance many of these installations are in operation without

problems. But it must be said, that also the most severe problems happened with these arrangements in high-

speed applications; complex to analyse and with high effort to solve.

pic. 19 – extreme example for traditional high-speed arrangement with 8 bearings in the shaft line

Today the aim is to solve the power transmission with the lowest possible number of parts (especially bearings)

that require maintenance and to ease the installation and alignment.

Every bearing position is also a potential excitation contact for structural noise and requires maintenance,

alignment and foundation and is a subject to wear and limited lifetime.

In modern installations we often do not find a cardan shaft anymore because it is a vibration exciter on its own

when used with typical deflection angles of 4..8°.

By bringing the thruster into a hydrodynamic reasonable inclined position of ab. 3..6° (which also needs to be

applied for the motor) a straight shaft line is reached.

Using hollow shafts made of seamless tubes of high strength allows much wider spans of these shaft lines due to

their high bending stiffness and low mass without a reduction of the natural frequencies of the shaft line.

In this way very well-arranged installations are reached; easy to install, alignment tolerant and with low

maintenance.

pic. 20 –optimum example for modern high-speed arrangement with 1 bearing on 8 metres installation length

In the installation above the one and only bearing in the shaft line is located in the bulkhead. So the bearing

housing flanged into the bulkhead is serving as bulkhead sealing as a double function.

The only “foundation” effort for the whole shaft line is a hole cut into the bulkhead and occasionally some

stiffening flat steels to tune the bulkheads membrane stiffness.



5. TYPICAL PROBLEMS OR “What can go wrong?”

5.1 thruster fixation – welded version

The thruster installation via a fully welded connection causes only few problems in daily practise.

Yards are used to welding. Once the thruster is put into position correctly and deflections of the welding process

are controlled within permissible alignment tolerances the installation offers a very rigid connection to the ship

which is normally resonance free for the a.m. main modes of the thruster. The installation method doesn’t

require any machining of the thruster trunk!

Nevertheless this installation method makes only ab. 20..30% of the delivered units.

In spite of the advantages mentioned above design or installation failures can also happen with this installation:

o defective alignment (which is difficult to correct afterwards)

o ship-sided flange to thin, to short or with to much overhang

o insufficient connection of the webs or webs to weak

For the alignment of the thruster it is useful to

attach provisional installation arms to the

thrusters´ main flange which should be set onto

adjustable hydraulic plungers or threaded spacers

to allow accurate vertical and angular positioning

of the heavy unit prior to the fixing by welding

(pic. 23).

Independent on how the alignment procedure is

made a certain vertical, horizontal and angular

failure is unavoidable and will change again

when the vessel is put into water due to the

different hull deformation between jacked-up

and floating condition of the tug.

Using long hollow shafts with none or only one

bearing position that can be averaged between

motor and thruster makes the whole installation

of the shaft line an easy work to do.

Perfectly straight installation with almost no

unnecessary bending deflection load onto the

hollow shafts and their deflection couplings is

reached when the motor is finally positioned after the welding of the thruster by using e.g. adjustable motor

mounts (both rigid or flexible).

Care should be taken to a proper documentation of the aligned motor position when using flexible mounts. This

helps to compensate setting of the rubber elastic motor mounts after a period of time. With height adjustable

flexible mounts these settings should be followed by compensation alignment.

pic. 21 – proper installation of fully welded

installation with smooth changeover from

thruster to well flange and full connection

of brackets

pic. 22 – faulty installation of fully welded

installation with well flange to thin or with

to much overhang / connection of brackets

to weak with shortened lashes



5.2 thruster fixation – bolted version with O-ring sealing

Although this installation method requires proper

machining of the ship´s flange it is the most popular

version used by our customers especially as installation

from below (with about 50% share of orders).

The accuracy requirements of ab. 0,15..0,2 mm in

plainness and ab. Ra=3,2 µm in roughness normally

requires machining of the flange in situ which is done

by specialised companies anywhere in the world

(see pic. 24).

Problems appear when the flange or the trunk is

machined in the workshop and afterwards welded into

the structure. In most cases the flange is then out of

tolerance and the unit is bolted down to the flange by

the high clamping forces of the very strong main

bolting. This results in forced deformation of the

steering gear and causes hidden problems like wear etc.

The typical shape failure of such a machined flange associates an “8” and can easily reach deflections of few

millimetres which is obviously the multiple value of the tolerance . In such extreme cases also to main bolting of

the thruster gets unsafe because the clamping force is used for flange deformation instead of supplying correct

and equally distributed friction force over the flange circumference (see pic. 25).

pic. 25 – uneven bolt pre-stress after setting of bolts by typical flange deformation due to welding stress

5.3 thruster fixation – bolted version with “rubber flange” installation

The rubber flange method is more tolerant with regard to unevenness and surface quality of the ship flange. Due

to the spot contact of the bolting a certain compensation comes from the flanges themselves (by bending

deformation). Sealing is not a problem due to the high compression value (at least 2 mm) of the 10 mm rubber

flange. Unevenness of at least 1,5 mm can be compensated by this installation method.

Even higher unevenness values can be tolerated by using thicker rubber flanges of e.g. 15 or 20 mm whereas ab.

20% compression is used and up to 3 mm unevenness can be equalised.

pic. 24 – typical requirements for flange machining



When making use of these extended

tolerances with the thicker rubber

flanges an individual machining of the

steel spacers between thruster flange

and ship flange is required to ensure

proper loading of each bolting position

and to prevent unallowable thruster

flange deformation.

Therefore the gap of each numbered

bolting position has to be measured and

the spacers (see “washers” in drawing

right side) have to be supplied with over

length to allow quick and easy turning

to measure on the lathe.

Attention: stoppers should be used in

radial direction because centering of the

flange is not possible. pic. 26 – rubber flange installation showing rubber flange

bolts, spacers and stoppers

The rubber flange installation method has come up in the past few years and is making about 20..30% of the

deliveries today. The installation method normally allows the pre-machining of the ships´ flange in the workshop

prior to welding-in (welding sequence has to be observed).



5.4 thruster fixation – global resonance of the thruster in the ship structure

Independent from the installation method the worst thing that can happen is a global resonance of the thruster

unit(s)! in the ship structure.

The thruster together with its sister unit can perform pendulum and vertical vibrations both synchronous and in

opposite phase. This delivers normally at least 6 relevant Eigenmodes of the thruster pair in the ship structure.

pic. series 27 – relevant Eigenmodes of the thruster pair in the ship structure



To determine these Eigenmodes the structural 3D model of the tug aftship (normally available) and a reduced

model of the thruster is used to keep the whole model simple and fast to calculate.

Mass supplements for the submerged surfaces have to be made to consider that the system is oscillating in water.

When calculating pendulum modes of the thrusters the hydrodynamic mass supplement for the hull plating can

be neglected due to the more local deformation area of the hull shell.

For the thruster submerged body the water supplement should be added for more exact examinations.

For the pounding modes the water supplement gets more important for the hull shell and less important for the

thruster. As long as comfortable distances to resonances are reached (e.g. >20%) water supplements needn´t to

be applied.

There are various recommendations available for water supplements for different submerged shapes.

E.g. DNV offers values in classification note 30.5. and many other sources can be found in the world-wide-web.

The reduced thruster model is a simplification of the thruster detailed finite-element-model to a pendulum with

similar properties.

3D-mass inc.

lower gear,

propeller, nozzle

bending

stiffnesses

cxx & cyy

+ mass of

steeringcolumn

mass of non-

azimuthing

part of thruster

+ membranestiffness of

steering gear

pic. 28 – conduction of the reduced thruster model from the detailed 3D-FEM-model of the thruster

The task is to prevent resonances of the propeller blade orders with the above shown Eigenmodes. The only

parameters to be tuned are the stiffness of the hull fixation and foundation and the number of propeller blades

(normally to be chosen either 4 or 5). Concerning the no. of propeller blades also the TVC may require a specific

blade number with regard to torsional resonances in the drive train which has to be dealt with.

5.5 shaft line – lubrication problems

Roller bearings with grease lubrication are used as a standard for shaft lines in tug installations.

It may appear trivial but still lubrication problems occur regularly in practise.

Insufficient lubrication of the roller bearings due to inadequate type or insufficient amount of grease is the

typical case.

But even overfilling of bearing housings can cause problems after a while of operation/maintenance.

Most problems occur with overheated grease in high rpm applications (1600/1800 rpm).

Few cases are known with insufficient maintenance because the necessity of maintenance of the bearings seems

to be obvious for the operator and bearings are normally easy to access and under daily observation.

Bearing housings with a large “dead space” compared to the grease filling recommendation may be problematic

especially when using high temperature grease with stiffer consistence at operation temperature (pic. 29).

Fewer problems are known with bearing housings with smaller “dead space” for high speed applications.



SCHOTTEL meanwhile uses high temperature grease as a standard in all shaft-lines that are supplied ready

assembled to our customers. Standard grease is adequate and can be used for speeds up to 1000/1200 rpm.

Maintenance manuals should be observed and individual consultation of service support should be used in case

of doubtful lubricant or bearing condition or excessive operation temperatures (after running-in!).

pic. 30 & 31 – overheated grease in under loaded high speed application /

coloured bearing metal due to insufficient lubricant

Bearing temperatures measured on the housing surface in way of the outer bearing ring (on the bearing housing)

shouldn´t exceed ab. 60 °C for standard grease and ab. 80 °C for high temperature grease.

grease type standard grease high temperature grease

NLGI class 2 1,5

base oil type mineral oil synthetic

thickener lithium clay

temperature range -25..130°C -55..180°C

dropping point 190°C 310°C

ref. viscosity base oil @ 40°C 100 cst 30 cst

A general comment:

Cardan shafts contain 8 needle bearings in their joints! A regular check of these joint bearings with respect to

presence and condition of lubricant and condition of raceways is highly recommendable to prevent vibration due

to play of the joints or even worse a fatal damage with the joint flying out of the shaft line.

A regular check reduces risks and isn´t very costly as needle bearings are very cheap spares.

pic. 29



5.6 shaft line – “Why do shaft lines vibrate?”

Shaft lines do vibrate because they are excited by dynamic forces mainly from dynamic reaction forces of the

cardan shafts under torque and deflection and in rare cases by unbalance of the shaft system. These are the main

sources for self-excited vibrations of the shaft line.

Vibrations can also be induced through the shaft line foundation. These vibrations could be named indirect

vibrations and are mainly excited by e.g. diesel motor foundation excitations or thruster excitations induced via

the thruster foundation.

The shaft line is a mass-elastic system and will react with forced vibrations which will be amplified when a

resonance of the shaft line is hit. As the whole shaft line runs in air and foundations are normally welded

constructions with rigid connections of plummer blocks to foundation the whole system has a very low damping

which leads to high amplification factors in case a resonance is hit!

pic. 32 – shaft vibrations and its excitation sources

Unbalance problems appear especially with worn-out equipment or in consequence of assembly failures.

Recommendations for the balance quality of the shaft components can be found in ISO 1940 and VDI 2060.

A balancing quality of 6,3 or better should be aspired (which is easy to reach in practise).

Most regular the cardan shaft excitation and induced vibrations from the motor foundation make the problems.

pic. 33/34 – cardan shaft alignment & dynamic bearing forces under torque and deflection



Cardan shafts:

o produce dynamic reaction forces onto the shaft line bearings with higher excitation frequencies at

higher shaft speeds

(depending on W- or Z- arrangement, torque & deflection angle)

o additional dynamic components appear due to alignment failures that are unavoidable in real

installations!

o an alignment accuracy of few tenth of a degree is required for cardan shafts

o deflection angles are normally chosen below 4° (max. 8°) to limit dynamic excitation forces

o cardan shaft excitation goes with 2nd

order of the shaft revolutions (and its higher orders!)

o a sufficient distance of the 1st natural frequency of the shaft line to this cardan shaft excitation

frequency is needed

o the so-called Campbell diagram below shows the design goal

pic. 35 – shaft line resonance shown in a Campbell diagram

o SCHOTTEL designs the shaft lines with the 1st natural frequency being at least at 20(15)% distance

above the cardan shaft excitation frequency at nominal speed (undercritical design)

o overcritical design with the natural frequency being far below the nominal speed (e.g. in the range of

idle speed to 50/60% shaft speed is possible and useful but requires exact knowledge of the

parameters

For the determination of the natural frequencies of the shaft line masses and stiffnesses of the shafts and bearings

are known by the shaft line maker.

Bearing foundation stiffnesses must be agreed. SCHOTTEL requires a radial design stiffness of 500 kN/mm;

300 kN/mm are used in calculations; poor values of ab.down to 100 kN/mm are regularly found in real

installations and cause severe problems!



It has to be considered, that the foundation stiffness is a linked serial stiffness of the bearing stiffness (roller

contact), the bearing housing stiffness and the welded foundation stiffness.

The foundation stiffness in the a.m. range influences the natural frequency of the shaft line significantly.

pic. 36 – influence of foundation stiffness on 1st

& 2nd

natural frequency of an intermediate shaft

o in cases of doubt SCHOTTEL evaluates the quality of the shaft foundation on request

o very often the shaft line foundation is not defined in the order entry phase, because steelwork for

shaft line foundation is scope of work of the yard and so the designer just leaves some space for the

foundation in its steel construction; this sometimes results in “uncontrolled” poor constructions

pic. 37 & 38 – poor constructions of shaft line foundations



Bulkhead bearings are a special case. With the traditional shaft line designs with multiple bearings

bulkhead bearings were normally prevented because of the “jump” in radial stiffness and the resulting undefined

load situation for the bulkhead bearing.

With the modern shaft designs bulkhead bearings are used willingly because they:

o offer very high radial stiffness due to the bulkhead itself

o the bearing flanged into the bulkhead substitutes the bulkhead sealing

Problems occur especially when:

o the bearing is an axially fixed bearing and the axial stiffness of the bulkhead is insufficient

o the bulkhead bearing is used together with more than 1 additional bearing in the same shaft line

In the worst case the shaft line may run into resonance and an axial vibration of parts of the shaft line or the

whole shaft line will appear. This normally results in a short-term damage of the bulkhead bearing or other

axially moveable components of the shaft line like cardan shaft splined connection or deflection couplings.

pic. 39 – bulkhead in resonance and possible solutions



The picture series 39 (last page) shows what can happen and

how it can be solved.

For a 1000 rpm application the main cardan shaft order of

excitation is f = (1000 rpm / 60) × 2 = 33,33 Hz.

Obviousely (first picture) the mass-elastic-system of the shaft

line (which is connected to the bulkhead via the bulkhead

bearing) and the bulkhead (being a membrane spring) has

an Eigenfrequency at ab. 31 Hz which would result in a

resonance with the cardan excitation slightly below the

nominal speed.

This cannot be accepted and will clearly result in at least a

very short lifetime of the bearing or even a fatal damage of

the shaft line.

pic. 40 – fatal damage due to resonance problem

Such problems can be solved with low effort both in the design phase (change of axial stiffnesses or masses of

shaft components / shaft separation) and also in case of detection in operation (additional stiffeners on the

bulkhead).

In the mid picture of picture series 39 two simple flat irons running vertically left and right of the bulkhead

bearing are used to shift the Eigenfrequency to twice its original value by stiffening!

In the bottom picture the problem is solved by making the shaft more lightweight by using a hollow shaft of 220

× 8 mm instead of the solid 140 mm shaft. This results in a safe increase of the Eigenfrequency by almost 30%.

To allow the prevention of this problem the axial bulkhead stiffness should be given to the shaft line supplier.

5.7 shaft line – unallowable displacements

Displacements of the ship structure under sea load of the hull, deflections of the hull under thruster forces,

thermal expansion and movements of the motor on its elastic mounts have to be taken by the shaft line.

In modern shaft lines with a low number of bearings over the shaft length and deflection couplings between shaft

elements the radial displacements are usually not critical. More problems are found in axial direction both for

static and dynamic displacements.

The following example shall give some absolute values that can be found in a real application:

pic. 41 – axial displacements in a standard straight shaft line

In the above installation with 11 metres distance between motor and thruster the following displacements occur:

(see next page)



o axial movement of the thruster input flange under thrust load in both directions (e.g. +/- 2 mm)

o thermal expansion of shaft line at differential temperature of the shaft of ab. 10..20 K (e.g. + 3 mm)

o axial movement of the diesel motor on its elastic supports (e.g. +/- 2 mm)

(value to be taken from 6-deg-of-freedom calculation of motor maker)

o bending of aft ship structure under thrust load with effect on thruster-motor distance

(can be neglected here)

As can be seen in the above example the totally needed axial compensation capacity of the shaft line is about

7 mm for a longer shaft line. Values can be somewhat higher for large boats or softer motor mounts.

To manage these deflections compensation elements are used in the shaft line.

Most commonly steel link couplings with steel or rubber joints or toothed couplings are used for this purpose.

These couplings are normally stiff enough not to influence the TVC and require no or only low maintenance.

Most flywheel-mounted elastic couplings (chosen by Torsional-Vibration-Calculation result and usually scope

of supply of the diesel motor manufacturer) allow significant axial displacements under a given axial stiffness.

This displacement capacity can be used , but the axial reaction force onto the shaft lines´ fixed bearings has to be

observed (see pic. 42).

With bell-mounted elastic couplings with integrated output shaft the problem is often reversed which means

that the bell-bearings are the weakest part in the axial chain and the problems occur in the motor makers scope of

supply. Reaction forces resulting from the stiffnesses of the shaft line compensation couplings will apply axial

and radial loads and also bending moments onto the bell bearings.

Additionally these bearings are also suffering high vibration load and increased ambient temperature from the

diesel motor.

Nevertheless, if properly designed and if all these aspects are kept in mind this solution is very favourite from

the point of view of installation effort and the low number of components to be used.

pic. 42 – axial displacement capacity of typical double-row elastic coupling / reaction force on shaft line

pic. 43 – bell-mounted elastic coupling with integrated stub shaft / reaction forces from shaft line



As explained above the radial displacements are rather uncritical with the modern arrangements with minimum

number of bearings per shaft length.

The following example shall give an idea of what are the limits in radial deflection:

With the required safety margin of 20% the minimum 1st Eigenfrequency of a 1600 rpm shaft line is:

fmin = 1,2 ×2 × 1600 [min-1

] / 60 [s] = 0,04 × 1600 [s-1

] = 64 [Hz]

for a shaft line with cardan shaft excitation. Of course in practise the shaft line is designed to a higher

Eigenfrequency if possible to have some additional safety margin.

Case A – traditional arrangement with solid shaft line:

power P= 2200 kW

shaft speed n= 1600 rpm

shaft diamter of the solid 42CrMo4V shaft d= 140 mm

bearing distance for under critical design l= 1720 mm

pic. 44 – 1st

mode of vibration of the solid shaft at ab. 90 Hz:

pic. 45 - permissible radial displacement of the centre bearing at bending fatigue limit of the shaft is 14 mm!:



pic. 46 - permissible radial displacement of the centre bearing at bearing lifetime Lhna=40.000 hours is 9 mm:

With this design the radial displacement capacity of the shaft line is ab. 14 mm at the bending fatigue limit of the

shaft and ab. 9 mm at the required bearing lifetime of Lhna = 40.000 hours! This corresponds to a length of the

shaft line of ab. 4000 mm.

Case B – modern arrangement with hollow shaft line:

power P= 2200 kW

shaft speed n= 1600 rpm

shaft diameter of the solid 42CrMo4V shaft d= 140 mm (at bearing positions!)

bearing distance for under critical design l= 2730 mm

pic. 47 – 1st

mode of vibration of the hollow shaft at ab. 90 Hz:



pic. 48 - permissible radial displacement of the centre bearing at bending fatigue limit of the shaft:

pic. 49 - permissible radial displacement of the centre bearing at bearing lifetime limit e.g. Lhna=40.000 hours:

With this design the radial displacement capacity of the shaft line is 14 mm at the bending fatigue limit of the

shaft and 13 mm at the required bearing lifetime of Lhna = 40.000 hours!

This corresponds to a length of the shaft line of 6000 mm.

These values may explain why the radial displacement is not so problematic for the shaft line design.

In fact a certain radial displacement is always present and helps to give the bearings additional pre-loading which

reduces the risk of slipping due to under load.

Problems can anyhow occur when single bearing positions get under load due to radial movements of the

foundations. On the other hand the above given values show, that a defined offset of the bearing positions during

installation of the shaft line offers sufficient tolerance to load the bearings also by the shaft cross force under

targeted bending of the shaft.

This method isn´t common practise today but should be used without fear.



5.8 shaft line – minimum required bearing load

Any roller bearing used in shaft lines requires a minimum load to keep the rollers rolling instead of slipping or

shifting (with unavoidable consequential damages)!

If this minimum load is extensively undershot (see pic. 50) slipping of the rollers will occur within short time.

kinematic zones in a roller bearing

deceleration zone

acceleration zone

zone of nominal speed

LOAD zone

pic. 50 – slipping of a roller bearing pic. 51- slipping damage on raceway

The required catalogue values of the bearing makers are regularly not fully reached in reality even so most

shaft lines run satisfactory and do not suffer slipping damages.

This indicates, that deviations can be accepted in practise. Nevertheless if the required values are considerably

undershot it gets critical and damages can be expected.

For a given shaft size different bearing types with different section sizes and thus load capacity are used.

This results in also different minimum loads required for these bearings made for the same shaft diametre.

The formula below indicates the minimum required load for standard bearing types used in shaft lines.

For spherical roller bearings the minimum radial load needed for grease lubrication is about:

Frad,min = 0,01 × C0 [kN]

with C0 being the static load capacity of the bearing taken from the makers catalogue [kN]

For cylinder roller bearings the minimum radial load needed for grease lubrication is about:

Frad,min = kr × (6 + 4n/nr) × (dm/100)2 [kN]

with: kr being the minimum load factor

n being the actual speed [rpm]

nr being the reference speed [rpm]

dm being the average of outer and inner bearing diameter [mm]

For a given shaft diameter the minimum load of the strongest spherical roller bearing may be about 10 times the

value of the minimum load of the smallest cylinder roller bearing section. This should be kept in mind.

Forced loading of the bearings can be made by radial offsetting of e.g. the centre bearing of a triple supported

intermediate shaft or by radial offsetting of the diesel motor using the radial stiffness of the main elastic coupling

to load e.g. the stub shaft bearings.

In case the calculation results show a bearing position being radially under loaded then this bearing should be

used as the fixed bearing of the respective shaft. By this the rollers all over the bearing circumference are more

or less equally preloaded and slipping risk is significantly reduced. This axial loading can also be done by using

the downhill-slope force of an inclined installation.



5.9 shaft line – maximum axial bearing loads

Significant axial movements of the shaft line occur in any tugs´ shaft line (see 5.7). These axial movements

multiply to axial constraining forces by the axial stiffnesses of the installed components. The resulting forces

have to be taken by thruster input shaft, motor crank shaft and also by the shaft lines´ fixed bearings.

Axial forces can be very high and have to be considered when choosing the section size of the fixed bearing.

Another source for axial forces which is often underestimated is the friction force transmitted via the splined

connection of the cardan shafts.

The following example of a 3000 kW / 750 rpm installation shows the problem:

o cardan shaft size: 390.80

o shaft torque: T = 38.200 Nm

o radius of the cardan shaft splined connection: r = 92 mm

The resulting friction force that can be transmitted axially via the splined connection of the cardan calculates to:

Fax = µ × T / r = 62282 N! with µ=0,15 being the worst case friction coefficient of a poorly lubricated splined connection

Of course this load will not appear constantly because vibrations and axial movements will make the splined

connection slipping. But especially under running-up conditions with steady thermal growth of the shaft line the

high axial force value can act onto the fixed bearing over a period of time which is critical for the fixed bearing

with respect to overheating.

With the a.m. power and speed data normally a standard cylinder roller bearing could be used as fixed bearing of

the intermediate shaft considering radial loading and required lifetime.

The axial load capacity of this bearing can be calculated to ab. 20 kN which is much less than required (see value

above).

The required bearing size is therefore e.g. a spherical roller bearing of the section size 23048 with much larger

dimensions (see size comparison in pic. 52).

Conical sleeves are normally used to clamp the bearing inner ring to the shaft line. It has therefore also to be

checked, if the transmittable axial force of this clamped connection is strong enough to take the high axial load

in cases like this one.

pic. 52 – necessary over dimensioning of the bearing due to high axial load

Note that such cases normally only occur when higher axial movements are to be compensated in a shaft line in

combination with low speed = high torque applications.



5.10 shaft line – unallowable bearing vibrations

Different causes for shaft vibrations have been precedently addressed. Finally the vibration level should be

measured during the setting-into-service of the tug. Permissible values are given in the diagram below.

Values should be measured as RMS (root mean square) values over a frequency range of 4..1000 Hz.

pic. 53 – recommended evaluation chart for shaft bearing vibration

As long as the above shown values are kept a short term failure of a shaft bearing isn´t probable.

Nevertheless damages can also occur when vibration levels are in the green range (e.g. lubrication or under load

problems due to misalignment).

Advice for the execution of the vibration measurements on such shaft lines will be found in ISO 20283-4 which

is under development.

Generally it should be observed to measure the fixed bearings in all 3 main directions whereas the loose bearings

are measured in the 2 radial directions only. Nevertheless with very high and soft bearing foundations or with

loose bulkhead bearings excessive axial vibrations of the foundation can cause problems with the roller bearings

especially with spherical bearings, because the outer bearing ring cannot follow the very quick vibration

movement and will wear out in the housing and the outer ring will axially hammer into the rollers within the

axial bearing play. Both effects can cause bearing damage.

Special problem stub shaft vibrations:

The picture below shows the problem of diesel motor vibrations induced to the stub shaft bearings especially to

the bearing close to the motor. It is a general recommendation to have the stub shaft sitting on a foundation

which is well connected to the motor foundation to prevent excessive radial movements of the motor to the stub

shaft. The disadvantage of the good connection of the foundations is the regularly higher vibration level of the

stub shaft foundation in excess of values as given in (pic. 53).

The motor permissible foundation vibrations range about 25 mm/s and values in that region and in excess of it

are often measured on the stub shaft (see pic. 54 - example of vibration measurements on a standard shaft line).

pic. 54 – example of vibration situation with excessive stub shaft vibrations due to diesel



As the roller bearings accept higher vibration levels at higher preload a useful trick is to use the radial

displacement capacity of the torsional-elastic coupling on the flywheel to radially pre-load the stub shaft

bearings within its allowable limits with regard to bearing lifetime.

pic. 55 – loading of the stub shaft by defined radial offset of the motor to the stub shaft

In this example of a 1650 kW / 1000 rpm application the shaft diameter is 140 mm and the radial load of the

bearings is ab. 1800 N by the weight of the shaft.

The dynamic force component of the cardan shaft running at 4° deflection angle is already 1500 N acting in the

horizontal plane.

The radial stiffness of the double-row elastic coupling is 2,4 kN/mm.

The permissible continuous radial offset of the coupling is 7 mm.

The limit load of the spherical roller bearing type 23028 (smallest bearing section for that shaft diameter) at

Lh10 = 20.000 hours nominal lifetime is 55 kN which gives enough margin for pre-loading.

The minimum load (catalogue value) for this bearing to be safe against slipping is 6,8 kN.

So a radial displacement of about 2..3 mm can be chosen which will result in ab. 8 kN pre-load of the bearing

which will work well in practise.

Still the elastic coupling will have sufficient radial displacement margin to take tolerances from the alignment

and operation deflections of motor to shaft.



6. TYPICAL EXCITATIONS OF THE NOZZLED THRUSTER

6.1 main static and dynamic force components acting on the azimuthing thruster

pic. 56 – azimuth thruster with main static and

dynamic force components

For the strength calculation of the ship structure the

max. values of the static components of Fx, Fy, Fz and

Mx, Mz are used in combination. My is ignored

because of low relative magnitude.

SCHOTTEL supplies these values with the document

“forces and moments” on request of designer or yard.

For that purpose the resulting values of bending

moment, steering torque and horizontal force are

usually transformed to the coordinates of the centre of

the thruster well top flange.

So the values can be directly used for strength

calculation of the ship structure.

For the modal analysis of the thruster in the structure

either a reduced pendulum model or a complete 3D-

thruster model is used. As long as a fair margin to

resonances is achieved the modal analysis is the

sufficient.dynamic calculation.

For the forced damped vibration calculation of the

submerged thruster in the ship structure the dynamic

components of Fx, Fy, Fz and Mx are used.

Mz is ignored additionally to My because of the very

high damping of the steering gears which results in

negligible magnitude as well. Forced damped

calculation is only used to check the effect of

unavoidable resonances.

pic. 57 – “forces and moments” document

for a nozzled thruster

My, My,dyn

Fy, Fy,dyn

Fz, Fz,dyn

Mz, Mz,dyn

Fx, Fx,dyn

Mx, Mx,dyn

o static thrust force + dynamic thrust component

Fx & Fx,dyn o side force + dynamic component

Fy & Fy,dyn o weight + vertical dynamic component

Fz & Fz,dyn o steering torque (highly damped)

Mz o propeller torque

Mx & Mx,dyn o propeller bending moment due to wake field

My & My,dyn



6.2 exemplary static and dynamic forces of the 4-bladed nozzled thruster

For design purposes some standardised values for the a.m. force and moment components can be used for the

4-bladed nozzled thruster. It has to be considered that these values are only valid for thrusters with nozzle with

typical design data for tug application!

The tabella below gives some standard values that can be used to make own calculations on such installations.

The values are related to the centre of the lower gear box (section point of propeller and azimuth axis) and can

act in any 360° rotated direction.

static thrust force Fx P × 170 [N/kW]

dynamic thrust component

(1st blade harmonic)

Fx, dyn1 10% of Fx

dynamic thrust component

(2nd

blade harmonic)

Fx, dyn2 5% of Fx

static side force Fy 2 × Mz / Dprop

dynamic side force


Fy,dyn1 10% of Fy

dynamic side force

(2nd

blade harmonic)

Fy,dyn2 5% of Fy

weight force Fz mthruster × 9,81[m/s2]

vertical dynamic force


Fz,dyn1 10% of Fy 1)

vertical dynamic force

(2nd

blade harmonic) Fz,dyn2 5% of Fy

1)

propeller torque Mx 16 × P[kW] × Dprop[m]

dynamic propeller torque Mx,dyn ± 25% of Mx

max. azimuth torque Mz 0,02 × nprop2 × Dprop

5

1) relation to Fy is used consciously

It has to be kept in mind, that most values are considerably dependent on e.g. motor torque, steering angle and

ship speed and can therefore deviate from the above given values in practise. As a coincidental example the

diagram below shows some measured values of the relative dynamic side force component Fy,dyn at a ship speed

of ab. 9 kn.

This may illustrate, that the a.m. dynamic forces can be transiently exceeded under dynamic steering operation at

higher ship speed.

0

0,05

0,1

0,15

0,2

0,25

0,3

-40 -20 0 20 40

I

II

III

IV

pic. 58 – influences of steering angle and ship speed on exemplary load component Fy,dyn

All the values above are thumb values which are good enough for basic design.

In case of problems with resonances a more detailed view of the individual problem needs to be made!



7. INTEGRATION OF THE THRUSTER INTO THE SHIP CONSTRUCTION

7.1 calculations with detailed and reduced models

In Chapter 4.3 (pic. 17) and 5.4 (pics. 27 & 28) the modal behaviour of the thruster in the ship structure is

already shown.

Further it is explained that the thruster can be reduced to a simplified pendulum model given to the customer

which can than be used for strength calculation of the structure and for modal analysis of the thruster in the hull.

The dummy pendulum is built-up as a 3D-mass (including the mass and inertia properties of the underwater

parts of the thruster) placed on the end of a beam element with different bending stiffnesses in the 2 pendulum

directions (so it has the same properties as the thruster steering column).

This 3D mass plus beam is fixed to a circular plate representing the mass and membrane stiffness of the non-

azimuthing part of the thruster. The diametre of the circular plate corresponds to the inner diametre of the

thruster well in the ship. Thus the dummy pendulum fits into the ship structural model.

The following comparative calculation example of an exactly modeled thruster unit and its dummy pendulum

model in a ship structure may invite yards and designers to make more regular use of this kind of calculation.

The results of the reduced and simplified models normally meet the results of the much more extensive high-

resolution models by few percent only (for the rather low frequencies that are examined here).

pic. 59& 60 – exemplary calculation of stress under max. loads for detailed and dummy model

pic. 61& 62 – exemplary calculation of 1st Eigenmode for detailed and dummy model



In the above example with a very stiff integration of the thruster the first pendulum frequency (1st Eigenmode) of

the thruster in the structure is ab. 40 Hz. The 1st and 2

nd propeller blade excitation frequency of the propeller

(rotating at 230 rpm) is 15,3 resp. 30,6 Hz which means the excitation frequency is ab. 25% lower than the first

Eigenmode of the thruster. So in this case everything is fine.

If the safety margin is very small and the propeller blade excitation is close to an Eigenmode in the speed range

of 80..120% of the nominal speed measures should be taken to prevent excessive vibrations and the

consequential reduction of component lifetimes.

The below example of a forced damped vibration calculation of a thruster pair in the structure shows the system

response close to the resonance.

In this case of a large thruster the pendulum frequency of 12,5 Hz is excited by the 1st propeller blade frequency

of 12,1 Hz at nominal speed. The dynamic thrust component is applied with ab. 8% of the nominal thrust.

The vertical vibration component that can be measured on the thruster foundation is calculated as svert = 0,2 mm

which corresponds to a vibration velocity of ab. 15 mm/s.

Of course this is not critical from the strength point of view, but the resulting vibration load will be unfavourable

for thruster component lifetime and comfort onboard.

pic. 63 – thruster longitudinal pendulum mode excited by 1st propeller blade frequency near the nominal speed

7.2 evaluation of the results

The above comparison shows the minor deviations of the reduced models´ results against the detailed

calculation. Same as for the shaft line vibrations a certain “clean” speed range especially around nominal speed

is the design goal.

Classification societies do not have strict assignments whether or not a structural resonance is accepted because

this is clearly dependent on what excitation magnitude hits what mode with what damping and what is finally the

vibration amplitude of these coactions.

As a general requirement SCHOTTEL advises to have the speed range of 80..120% of nominal speed clean of

resonances of the 3(6) main modes of the thruster(s) with the 1st and 2

nd blade harmonic.

When calculated with detailed models the clean range can be reduced. In any case the range between 90..110%

should be kept clean.

Resonances in the speed range of ab. 50..80% of the nominal speed can often be accepted, because the relation of

excitation and damping is good enough to deliver uncritical vibration amplitudes only.



For the example shown under 7.1 (pic. 63) the calculation result and alternative solutions are displayed in the

Campbell diagram below.

It can be seen, that there will be a problem with the 1st blade order hitting the 1

st pendulum Eigenmode of the

thruster in the longitudinal direction (see point A).

This will require either a stiffer connection of the thruster to the hull (“undercritical” design) with a potential

shifting of the natural frequency to point B at ab.13,7 Hz (see also 4.1 – influence of connection stiffness) or

alternatively the use of a 5-bladed propeller with consequently a 25% higher excitation frequency (“overcritical”

design) which will leave the resonance point at ab. 75% of the nominal speed which can normally be accepted

(see point C).

Point C is clearly outside the required resonance-free speed range of 80..120% nominal speed.

Point B is inside this range but already higher than 110% nominal speed which could be accepted under

consideration of the high model accuracy.

Campbell Diagram

5

10

15

20

50 100 150 200 250 300

propeller speed [rpm]

fre

qu

en

cy

[H

z]

1st Eigenmode

1st Eigenmode (stiffened connection)

1st propeller blade frequency (4-bladed)

1st propeller blade frequency (5-bladed)

idle speed

A

B

C

nom. speed

pic. 64 – illustration of the resonance problem and the 2 alternative solutions in a Campbell-diagram

The examples above may illustrate, that the examination of the dynamic behaviour of the thruster in the structure

can have different calculation depths starting with a “first shot” done with a simplified pendulum model in a

coarsely meshed hull model.

If resonances are found within a speed range of 80..120% of the nominal speed constructive counter measured

should be taken to solve the problem.

If resonances are unavoidable more detailed models to examine the modal behaviour or even a forced damped

calculation can help to make a decision.



8. SUMMARY

With its long experience in the development, design and construction of azimuth drives for the tug application

SCHOTTEL today has methods and experts available to support any task of integration of the drive system into

the ship.

The responsible project engineer in the sales department accompanies the project during the important order-

entry-phase in which the most important technical questions are clarified within a short period of time of a few

weeks. Within this period of time all order specific aspects of design are handled and corresponding calculations

are made by mainly the technical branch of the order management department under assistance of the

development departments of hydrodynamic, electrical, hydraulic and mechanical engineering.

From the technical point-of-view the order-entry-phase is closed with the release of the final and official

installation drawing showing the drive system integrated into the hull and coupled to the motor.

Any relevant information from the customer side needs to be included beforehand!

For the fixation method of the thruster to the hull SCHOTTEL offers a number of 3 variants with individual

advantages to be chosen by the customer. Different premises lead to the decision for either a welded connection,

a bolted connection with O-ring sealing or a “rubber flange” installation with water beam cut rubber plates.

Any of these alternative solutions has its own specific advantages and disadvantages from the point of view of

the customer, especially the yard that has to install the thruster properly.

All the 3 variants are furthermore possible with installation from “above” or “below” which influences the

degrees of freedom for the construction above the thrusters and the maintenance possibilities.

Additionally to the main connection between thruster main flange and ship flange the radial webs of the thruster

or even the bottom plate can be connected to the hull structure. This is to a lesser extend to gain strength but

more regular to gain stiffness and to influence the modal behaviour of the thruster in the hull structure.

For the various installation methods design criteria and installation tolerances are given and hints to prevent

problems were made.

For the shaft line the differences between the traditional arrangement with inclined solid steel shafts and cardan

shafts and the modern straight designs with hollow shafts and compensation couplings are shown.

The design goal of using the lowest possible number of bearings and other components requiring maintenance is

highlighted.

The most common problems of installation and operation of shaft lines like lubrication, vibration, insufficient

foundation stiffness (including the special case of bulkhead bearings) and over- and under load mainly due to

radial and axial displacements are explained and examples are given.

The effort of SCHOTTEL to reach a mostly standardised modular shafting system which nevertheless comprises

solutions for all kind of installation is demonstrated.

The standard components used with their specific pros and cons and properties are addressed.

The SCHOTTEL modular shafting system allows the order specific design of the shaft line considering

questions of strength, lifetime and vibration safety of the drive train.

To get a better overview on what is done and expected on both sides (supplier and customer) an overview over

the standard scope of calculations and information exchanged is given.

By this it gets clearer to the parties involved what is expected from whom and what assumptions are made.

With regard to thruster dynamic in the ship structure the various force and moment components acting on the

thruster are explained and thumb values are given that can be used for dimensioning and for the basic design.

Examples are given for the modal behaviour and for forced vibrations of the submerged thruster when excited by

realistic dynamic load components.

Working with simplified models with reduced resolution is exemplary shown. SCHOTTEL offers forces and

moments plus reduced thruster models to be used by the designer for the strength and vibration calculation of the

thruster-hull system.



For further comments, questions and more detailed information please do not hesitate to contact:

Manfred Heer head of R&D department SCHOTTEL GmbH Mainzer Straße 99 D - 56322 Spay phone: +49 (0) 26 28 / 61-419 fax: +49 (0) 26 28 / 61-380 mobile: +49 (0) 170 / 63 73 325 mailto:[email protected] www.schottel.de

10th

of February 2009, Manfred Heer



sources:

1] FAG & SKF bearing maker technical catalogues

2] FAG support center – information on damages

3] GEWES technical information on cardan shafts

4] SVA Schiffbau-Versuchsanstalt Potsdam report Nr. 3292

5] SCHOTTEL internal technical papers

Tugnology.pdf

Documents

Transcript of Tugnology.pdf