A durable mooring system for a winch-based wave energy converter

Mingming Wang

Master of Science Thesis MMK 2017:90 MKN 205

KTH Industrial Engineering and Management

Machine Design

SE-100 44 STOCKHOLM

1

Examensarbete MMK 2017:90 MKN 205

Dellösning för en vinsch-baserad vågenergiomvandlare

Mingming Wang

Godkänt

2017-06-09

Examinator

Ulf Sellgren

Handledare

Anders Hagnestål

Uppdragsgivare

KTH

Kontaktperson

Ulf Sellgren

Sammanfattning Projektet har behandlat utvecklingen av en ny teknik för en förnybar energikälla, vågenergin,

som anses vara en av de mest lovande förnybara resurserna med potential att bidra till en

energiproduktion som motsvarar cirka 10 procent av världens energiförbrukning . Ett

punktabsorberande koncept som använder en kraftuttagsenhet (PTO) omvandlar havsytans

vågsrörelser till elektricitet. På grund av hårda arbetsförhållanden ger underhållsarbete stora

problem och ett förtöjningssystem behöver utvecklas. Syftet med detta projekt är att utforma ett

hållbart förtöjningssystem för minst 20 års drift, även i en hård havsmiljö.

En geometrisk modell av förtöjningssystemet har skapats baserad på dimensionering av dess

komponenter. Flera koncept genererades och utvärderades med en Pugh-matris. En simulering av

de olika spänningar som påverkar systemets prestanda gjordes för att validera designen.

Dessutom har detaljkonstruktion av de olika delarna av systemet gjorts, så att de kan tillverkas i

ett framtida arbete.

Nyckelord: Vågomvandlare, vinsch, utmattningsberäkningar

2

3

Master of Science Thesis MMK 2017:90 MKN 205

Durable system for a winch-based wave energy converter

Mingming Wang

Approved

2017-06-09

Examiner

Ulf Sellgren

Supervisor

Anders Hagnestål

Commissioner

KTH

Contact person

Ulf Sellgren

Abstract This project has dealt with the developing a new technology for a renewable energy source, the

wave energy, which is considered as one of the renewable resources with a potential to

contribute to an energy production corresponding to about 10% of the world’s energy

consumption nowadays. A point absorber concept that is using a Power Take-off (PTO) unit

converts the sea surface wave motion into electricity thanks to a buoy at the sea surface which is

moved by the waves. Due to harsh working conditions, the maintenance would cause too many

issues, and a mooring system needs to be developed. The aim in this paper is to design a durable

mooring system for at least 20 years of operation even working in a harsh sea environment.

A geometry model of the mooring system has been built since the dimensioning of its

components was performed. Several concepts were generated and evaluated with a Pugh matrix.

An analysis of the different stresses affecting the performance of the system was made to

validate the design. In addition, the detail design of the different parts of the system has made to

allow their manufacture in future work.

Keywords: wave energy converter, winch, fatigue calculations

4

5

FOREWORD

This master thesis was carried out at the Department of Electric Power and Energy Systems at

KTH Royal Institute of Technology, starting in January 2017. The author of the project is

grateful to have had the opportunity of working with my supervisors, Anders Hagnestål and Ulf

Sellgren. Thanks for their endless support, guidance and help. It has been really interesting to

contribute in the developing of a renewable energy technology, which is a really important area

in current research worldwide. Moreover, it has also been a really satisfying experience to study

and work at KTH, with a lot of kind and lovely people around me during these months.

Wang MingMing

Stockholm, June 2017

6

7

NOMENCLATURE

Notations

Symbol Description

𝜎𝑡 Tensile stress (Pa)

𝐹 Peak force (N)

𝐴0 Sectional area of a bar (m2)

𝜎𝑡𝑚𝑎𝑥 Ultimate tensile stress (Pa)

𝐾𝑡 Stress concentration factor

𝐹𝑚𝑒𝑎𝑛 Mean force (N)

𝑑1 Inner diameter of the bearing (mm)

b1 Width of the bearing (mm)

𝐾 Specific wear rate

P Contact pressure (Pa)

V Sliding speed (m/s)

T Sliding time (hour)

n Bearing rotational speed (rpm)

𝐷𝑑 Diameter of the drum (m)

𝑉𝑑 Speed of drum (m/s)

𝑑𝑝 Diameter of the pin shaft (mm)

τ Shear stress (Pa)

Abbreviations

CAD Computer Aided Design

WEC Wave Energy Converter

PTO Power Take Off System

8

9

TABLE OF CONTENTS

SAMMANFATTNING 1

ABSTRACT 3

FOREWORD 5

NOMENCLATURE 7

TABLE OF CONTENTS 9

1 INTRODUCTION 11

1.1 Background 11

1.2 Purpose 13

1.3 Delimitations 13

1.4 Method 13

2 FRAME OF REFERENCE 15

3 THE DESIGN PROCESS 17

3.1 Requirement specifications 17

3.2 Conceptualization 17

3.3 Initial concepts 17

3.3.1 Concept 1 17

3.3.2 Concept 2 18

3.3.3 Concept 3 19

3.3.4 Concept 4 19

3.3.5 Concept 5 20

3.3.6 Concept 6 20

3.3.7 Concept 7 21

3.4 Evaluation and selection of concepts 21

3.5 Component development 24

3.5.1 Link 24

3.5.2 Pin shaft 25

3.5.3 Bearing 26

10

3.5.4 Linker 28

3.6 Validation 29

3.6.1 Simulation of Link 29

3.6.1 Simulation of Pin shaft 31

3.6.1 Simulation of Linker 34

3.7 Material Fatigue 35

4 RESULTS 37

5 DISCUSSION AND CONCLUSIONS 43

5.1 Discussion 43

5.2 Conclusions 43

6 RECOMMENDATIONS AND FUTURE WORK 45

6.1 Recommendations 45

6.2 Future work 45

7 REFERENCES 47

APPENDIX A: SUPPLEMENTARY INFORMATION 49

APPENDIX B: DATASHEET OF BEARING 51

APPENDIX C: WATER V SEAL 53

APPENDIX D: DETAIL DRAWING OF MOORING SYSTEM 55

11

1 INTRODUCTION

This chapter describes the background, the purpose, the limitations and the methods used in the

presented project.

1.1 Background

A Wave Energy Converter (WEC) is a technology that transfers the power of ocean waves to

electricity. It belongs to a branch of the technology that aims to developing the renewable

energies, and it is becoming a more and more important area on research, since with the global

attention now is being drawn to the climate change and the rising levels of CO2 in the

atmosphere. The waves in the sea are a huge and largely untapped energy resource, whose

potential to contribute to energy production is considerable.

There are several reviews of wave energy converter concepts (see references [2], [3]). Most of

the existing wave energy devices are still being investigated, at the R&D stage. It is hard to build

a wave power unit: it has to be considered that it has to withstand the harsh conditions at the sea

and at the same time, it has to produce electric power at a competitive cost. So it can be said to

be relatively immature when compared to other renewable energy technologies. However, using

waves as a source of renewable energy has a number of significant advantages over other

methods of energy generation. Compared to other renewable energy sources, ocean waves give a

higher energy density [1]. Winds generate the waves and solar energy generate the in turn. A

solar energy intensity of 2-3 kW/m2 in horizontal surfaces is converted to an average power flow

intensity of 2-3 kW/m2 on a vertical plane, perpendicular to the direction of the propagation of

the waves, just below the water surface [4]. Another advantage is also the low impact to

environment the while using the waves’ energy. What is more, there are low energy losses

during the large distance motion of the waves. Compared to wind and solar power devices, wave

power can generate energy up to 90 percent of the time [5].

According to the large variability in designs and concepts, it can be said that there are three

primary types of WEC for now: attenuator, point absorber and terminator. In this developing, the

point absorber concept of Fred Olsen’s Lifesaver [6] shown in Figure (1) is selected as the

developing area. A point absorber is a device that, when moved by the ocean surface waves, has

a Power Take-off (PTO) unit that converts the motion of the ocean surface waves to electricity. It

is a floating structure that heaves up and down on the surface of the water.

12

Figure 1: Fred Olsen’s Lifesaver concept, an example of a winch-based point absorber.

A lifesaver, pictured in Figure (2) , is one of the developed WECs by Fred Olsen. It consists of a

floater, the power Take-off (PTO) units, the primary mooring line and the secondary mooring

system. The primary mooring line is fixed to the seabed at one side and rolls around a drum at

the other side.

Figure 2: Conceptual sketch of lifesaver with the main PTO components; winch, gearbox and

generator. Courtesy Fred Olsen.

In the global energy market, there are a great number of technical challenges to increase the

performance of wave power devices. The main challenge in this research is to design a durable

mooring line for the WEC to meet the requirements. Considering how expensive can the

maintenance be, the expected operation life of the system is at least 20 years. It means that the

buoy will heave up and down about 80 million times, causing big issues on the wear and fatigue

of the transmission. The predicatable lifetime for even the most sophisticated rope products is

only of a few hundred thousand bending cycles, due to the fact that bending under tension will

cause internal abrasion between strands. Furthermore, there are some uncertain impact variables

from the enviroment that have to be considered. During a storm, the wavelength can reach up to

30 meters producing huge forces, so it has to be ensured that the designed mooring line can be

operated at a high-force oscillatory motion.

13

1.2 Purpose

The aim of the project was to design a mooring system for WECs (Wave Energy Converter) for

at least 20 years of operation in the sea environment. The primary focus was to design a durable

mooring line for a selected wave energy converter concept. Considering the impact from the

weather, the winch will oscillate badly with rough weather, such as storms. The mooring line is

rolling on a specific drum. However, the calculations of the mooring line on the drum were not

performed due to the lack of information of the drum’s dimension. A chain transmission was

considered to replace the mooring line, because it does not suffer from the bending effect.

Meanwhile, it was also possible to find out what the design requirements for a winch-based PTO

are. Furthermore, a concept was generated and evaluated so that the predicted life time is was

ensured to be over 80 million cycles. The key issue was the effect from wear and fatigue, which

needed deeper analysis and simulations when the mooring system was operating. The selection

of bearing and the arrangement of joint also needed to be considered.

1.3 Delimitations

The working conditions and environment for a mooring system is the sea. Therefore, the

manufacture and running tests were not performed. However, the theoretical calculations and

analysis are presented in the research and were validated with the results from software

simulations. Due to lack of time, the test was passed on to future work and the accuracy of the

predicted lifetime could not be proved. The whole system of the WEC has been divided into

several groups, being the schedule on each group different, so it was decided to manufacture the

prototype after future works. The electronic system of the WEC was not considered in this thesis

work, whereas the developing of a durable mooring line was based on a point absorber wave

energy converter. Since some technical data of the winch are not published, the used nominal

data was as much close to the reality as possible.

1.4 Method

An Engineering Design process and CAD (Computer Aided Design) were utilized in this project,

which consisted of the design of a mooring line for the WEC. Firstly, in order to identify the

project needs, a search and collection of relevant information in the field was carried out to

deeply understand the system. Based on it, the system requirements definition was done with the

support of the supervisor. Then, the conceptual designed was prepared and generated. At the

generation stage, brainstorming was used as a technique to keep an open-minded approach to

developing different solutions. The members who joined the brainstorm had mechanical

engineering background. The evaluation of the concept were done and discussed with the

supervisor. Modeling of the concept was performed after its selection with the help of a Pugh’s

matrix. The CAD model of the system was performed using the software SolidEdge. Finally,

engineering calculations were used to determine the dimensions of the mooring line. Using the

simulation to analysis the force distribution was demand to ensure the system working.

Furthermore, it was also requested to predict the lifetime of the mooring line, so fatigue

calculations were performed.

14

15

2 FRAME OF REFERENCE

The reference frame is a summary of the existing knowledge and former performed research on

the subject. This chapter presents the theoretical reference frame that is necessary for the

performed research, design or product development.

Designing a durable mooring line for winch-base requires the designer to have and generate

enough knowledge about the structure of the design process. Product design, business and

production are three important factors that determine the success or failure of a product. A

product’s function is a description of what the object does, which belongs to the product design

factor. Product’s form, materials and manufacturing processes are the product’s function. The

form which means the product’s shape, color, texture… and other factors that affect its structure.

The materials and manufacturing processes used to produce the product are as important as the

form. Thus the designer needs to be concerned on the function, form, materials, and

manufacturing processes firstly [7].

It is difficult to decide how to select the best concept when different solutions to a particular

problem have been generated in the design process. It is certain that it is easier to select the

wrong concept than the best concept. The design is said to be conceptually vulnerable if the

wrong concept was selected. Conceptual vulnerability was and is a major problem in any design

situation [8]. Stuart Pugh’s concept selection is a procedural tool for controlled convergence to

the best possible solution to a design problem. The group number participants concept selection

process should have good insight and awareness of the possible solutions and find that it

simulates the generation of new concepts.

A mooring line system is used to keep the floating structure in position, therefore wave energy

converters also need a mooring system to ensure its functionality and its safety. Under the effect

of environmental forces such as winds, currents and waves, station-keeping systems could limit

the motion of the floating structure to a reduced area. A station-keeping system can be a passive,

an active or a combined active-passive system up to the principle for providing the restoring

force. Design codes for mooring line system have been developed from offshore oil and gas

industry, such as DNV OS-E301[9] , API RP-2SK [10] and ISO 19901-7[11]. It is normally

considered that for the life time of the device, the fatigue limit state is used for the ultimate

design check. However, specific design codes for mooring systems of wave energy converter

have not been generated yet. DNV and Carbon published a guideline on applying the existing

codes to design and operation of WECs. However, there are normally unmanned that the

potential risk associated with mooring failure is lower for WECs. The key issues related to the

mooring system design for waves have been pointed by some researches, see [12], [13], and [14].

Since the 1970s, there are many different types of floating wave energy converters that have

been developed. The simplest concept is a point absorber, which due to the small size of the

mooring structure system has strong coupling with the WEC motion. The influence of catenary

line systems on the huge, harsh and pitch motions of a point absorber has been studied in [15,

16]. The mooring system can be designed to make its effect on the wave energy absorption

small or even such that energy capture improves as compared with a freely floating absorber

from studies. In addition, for estimating the energy absorption in a realistic manner, it is

important to know that the mooring line will show a strong damping effect on WEC motions, see

[17, 18].

16

17

3 THE DESIGN PROCESS

In this chapter the working process is described. A structured process is often called a

methodology and its purpose is to help the researcher/developer/designer to reach the goals for

the project.

3.1 Requirement specifications The aim is to design a mooring system, different from just a rope, which has two fixed degrees

of freedom and a life of about 80 million bending cycles. There are some criteria that have to be

considered, based on the requests from the WECs.

The following aspects are to be considered:

The mooring line must have two degrees of freedom at axial and vertical direction.

The cost of the product should be as low as possible.

Detail design on the selected concept.

The mooring system should be able to operate 20 years.

The load capability should be at least 200 kN.

The working environment for the system is going to be the sea.

3.2 Conceptualization

Some concepts have been generated and presented in this report as simple sketches. All the

concepts were discussed and generated with different people from the mechanical engineering

department. The conceptualization should aim at generating ideas without calculations or

mathematical evaluations, and will be used as a guide to decide on a final concept. The chosen

concept will be developed further.

3.3 Initial concepts

3.3.1 Concept 1

A universal joint is a joint or coupling which has a pair of rods that cross together and are

connected by a cross shaft. It is usually used in shafts that transmit rotary motion. As the rod

allows bending in any direction, it could be used as a linker in the mooring line. It is also

possible to develop a seal solution for this concept.

18

Figure 2: Universal joint.

A big problem with this concept, among some others, is that the peak force on the joint is around

200kN, so it is easy to for the joint to break before reaching its operational life of 20 years.

Another problem could also be that it will not easy to assemble causing high costs on the

maintenance.

3.3.2 Concept 2

A knuckle bearing is a spherical sliding bearing and consists of an inner and an outer ring which

have as a contact surface a spherical sliding surface. Depending on its different types and

structures, it could bear larger load and others load.

It can produce self-lubricating during the work due to the spherical outer surface of the inner ring

with a composite material. Generally it is used for lower speeds of swing motion, but it can also

be used for allowing tilt movement in a certain angle range.

Figure 3: Knuckle bearing.

In this concept, it is a challenge to protect the bearing from the sea water. As the size of the

component is big, it is hard to make a seal for the bearing. For the swing motion, the bearing

could tilt in certain limited angle range. This could be a problem if the requested angle exceeds

the limited range.

19

3.3.3 Concept 3

The idea consisted of two welded crossed cylinders and adding the ribs on the side in order to

make it stiffer. A pin shaft would go through and would place in the cylinder and support with a

bearing. A link connects the joint with the pin shaft.

Figure 4: two cylinder joint.

A probable problem is that the geometry structure is hard to place on the drum, and doubtfully it

could operate 20 years at high load capability, even though the ribs would increase its stiffness.

3.3.4 Concept 4

A wrist joint basically consists of a linker, a link and a pin shaft. The linker connects two links,

and the pin shaft connects the link and linker. It has a bearing in the cylinder that supports the

pin shaft so that the link can rotate. Therefore this joint has freedom of movement in the axial

and vertical directions. It is also possible to come up with a seal solution for the joint.

Figure 5 : wrist joint.

The main issue of this concept is related to the stiffness of the joint. The performance of the joint

needs to be analyzed and validated. Meanwhile, it is also important to find a reliable seal

solution.

20

3.3.5 Concept 5

A chain is a series of connected links which are tipically made of metal. It is simple and easy to

find at the market. Compared to other solutions, it is lighter and cheaper.

Figure 6: Chain.

The problem in this concept is the large wear that will appear in the contact area between each

link. It is hard to met the target operation life of 20 years and bending cycle of 80 million times.

3.3.6 Concept 6

The idea is to use a chain drive instead of the mooring, so a chain is placed on a gear shaft,

which allowing it to tilt on a certain angle.

Figure 7: Chain driver.

21

The problem in this concept is that it may request a high performance from the drum shaft. Due

to the fact that the drum shaft should have high load capability and be movable at the same time,

it is too difficult to design the drum shaft. However, it could be a selection in the future work.

3.3.7 Concept 7

A ball joint is a spherical bearing that connects the control arms to the steering knuckles. It

consists of a bearing stud and a socket.

Figure 8: Ball joint.

The problem in this concept is that it will have lots of wear when a load close to 200 kN is

applied during a period of time. The seal solution is also expected to be a challenge.

3.4 Evaluation and selection of concepts

In order to determine the best concept to start the detail design, a decision method should be

implemented. In this case, a Pugh’s matrix was created to weigh the advantages and

disadvantages of all concepts numerically. The Pugh method, also known as Decision Matrix

Method, is a method to quantitatively assess the pros and cons of a design, and it is regularly

used in engineering.

It consists of a dimensional matrix, where the current design is used as reference (values of

zero), and the rest of the concepts are set in columns next to the reference. Each row in the

matrix is a requirement (quantitative or qualitative), and each requirement is given a weight

(numerically). Afterwards, a negative, a positive or a neutral value is assigned to each concept’s

parameter, then they are summarized and a total sum is calculated. The design with the highest

sum should be the most appropriate. This shall help in the decision-making process to start the

design process of a new concept.

The current mooring line for WEC is usually a rope, so in the evaluation matrix the rope was

used as the reference to assess the rest of the concepts. From the table, it can be seen that

universal joint, knuckle bearing and wrist joint proved themselves to be the best solutions.

22

Table 1: Pugh’s evaluation matrix

Criteria

Weight

Synthetic

Rope

Universal

Joint

Knuckle

bearing

Two

cylinder

joint

Wrist

joint Chains

Linear

Chain

driven with

adjustable

shaft

Ball

joint (1-5)

High load

capability(200kN) 4 R + + + + + + -

Durability 5 E + + - + - - -

Ease of Assembly 2 F - - - - - - -

2 fixed DOF 5 E S S S S S S S

Cost 5 R + + + + + + +

Low weight 3 E - - - - - - -

Sea water

resistance 3 N - - - - - - -

Ease of

maintenance 3 C - - - - S - -

Bending 5 E + + + + - + +

Construction

simplicity 1 - - - - - - -

Sum Neutrals(0) 0 1 1 1 1 2 1 1

Sum Positive(+) 0 4 4 3 4 2 3 2

Sum Negative(-) 0 5 5 6 5 6 6 7

Total Scoring 0 -1 -1 -3 -1 -4 -3 -5

Sum weighted Positive(+) 0 19 19 14 19 9 14 10

Sum weighted Negative(-) 0 12 12 17 12 19 17 21

Total weighted Score 0 7 7 -3 7 -10 -3 -11

Further Development 0 1 1 1

The criteria for assessing the concepts are: high load capability (200 kN), durability, ease of

assembly, 2 fixed degrees of freedom, cost, low weight, sea water resistance, bending and

construction simplicity. Between them, high load capability, durability, 2 fixed degrees of

freedom, cost and bending are more important if compared to other criteria.

High load capability indicates if the component has the ability to support 200 kN force of static

moment and not break. Durability refers to how many bending cycles can the joint withstand

during its life time, which means the longer the better. 2 fixed degrees of freedom refers to the

fact that the designed system needs to have two degrees of freedom, allowing bending and swing

motions. Actually the designed system must met this request otherwise it will not work. Cost

23

refers to the how much money will be spent on manufacturing the system. As this project is

research project with limited funding, it is better and necessary to design a low cost solution.

Bending refers to the stress when the mooring line is placed on the drums. All the other concepts

have a bearing solution except the chains, so they all perform better if compare with rope.

From the previous results, a universal joint, a knuckle bearing or a wrist joint should be

considered to further development. But the solution is supposed to one, so the next step is a

further assessment of the concepts. Based on the design aspect, it is good to design a low weight

system. Considering there is not harsh weather on the sea every day, a new idea is to combine

the chain solution with a rope. Most bending cycles distribute on the normal waves, so it is good

to using chain instead of rope. This way a big quantity of mass can be saved.

Under this text is the further assessment matrix. As it is used for a deeper development of the

concept, universal joint was selected as the reference.

Table 2: Further evaluation Pugh’s matrix

Criteria

Weight

Universal

joint and

rope

Sealed

Universal

Joint and

rope

Knuckle

bearing

and rope

Two

cylinder

and

rope

Wrist

joint

and

rope

Chains

and

rope

Linear

Chin

driven

with

adjustable

shaft

Ball

joint

and

rope (1-5)

High load

capability(200kN) 4 R S - - + S + -

Durability 5 E S - - + - - -

Ease of Assembly 2 F - S - S + - -

2 fixed DOF 5 E S S S S S S +

Cost 5 R - + - S + S -

Low weight 3 E - - - - + - -

Sea water

resistance 3 N + - S + - - S

Ease of

maintenance 3 C - S S S S - -

Bending 5 E S S S S S S S

Construction

simplicity 1 - + - S + - +

Sum Neutrals(0) 0 4 4 4 6 4 3 2

Sum Positive(+) 0 1 2 0 3 4 1 2

Sum Negative(-) 0 5 4 6 1 2 6 6

Total Scoring 0 -4 -2 -6 2 2 -5 -4

Sum weighted Positive(+) 0 3 6 0 11 11 4 6

Sum weighted Negative(-) 0 14 15 20 3 8 17 22

Total weighted Score 0 -11 -9 -20 8 3 -13 -16

Further Development 0 1

24

From the evaluation results it can be seen that the wrist joint got a higher score than the others.

Therefore, the selected concept was the Wrist joint.

3.5 Component development

From the beginning, based on the fact that steel has a long-term durability and good

performance, it was decided to use structural steel as the material for the components, with a

yield stress of 250 MPa and a tensile stress of 500 MPa. The main components of the wrist joint

are a link, a linker and a pin shaft.

3.5.1 Link

The geometry of the link is a cylinder that has two holes on both ends. In the hole there is a key

slot that fixes it with the pin shaft. A simplified model of the link can be found in the following

figure.

Figure 9: Sketch of the link.

In order to prevent the link from breaking while in the operation, it is needed to determine its

minimum diameter. The required minimum diameter of the link can be calculated using this

equation (1).

𝜎𝑡 =𝐹

𝐴0 (1)

Where 𝜎𝑡 is the tensile stress of the material, 𝐹 is the peak force and 𝐴0 is the sectional area of

the bar.

It is convenient to round the shaft corner to prevent the shaft failure. Theoretical stress

concentration factors (Kt) of a shoulder fillet can be calculated for the equation for tension load.

The width of the bar is 35 mm, the diameter of the link is 90 mm and the shape factor is 2.5.

25

Figure 10: Stress concentrations for bar with fillet.

𝐾𝑡 =𝜎𝑡𝑚𝑎𝑥

𝜎𝑡 (2)

Where 𝜎𝑡𝑚𝑎𝑥 is the ultimate tensile stress of material and 𝜎𝑡 is the tensile stress of sectional area

of bar.

The dimension of link is shown in the appendixes of this project, and the calculated rounding for

the bar is 3 millimeter, which is connected to the link. The size of the keyway was designed

using the standard keyway dimension, which can be found in the appended document. Moreover

2D PMI drawings are included in the appendixes also for the purpose of tolerance checks after

manufacturing.

3.5.2 Pin shaft

The pin shaft connects the link and the linker. A simplified model is shown in Figure (11). There

are two bearings placed on each side and they provide support for the pin shaft. In the middle of

the shaft, the key locks the pin shaft to the link so that they rotate together. The size of the key

follows the standard size table, which is shown in the appendix document.

Figure 11. Sketch of pin shaft.

26

For a uniform load distribution on the shaft with a peak force of 200 kN, the demanded diameter

𝑑𝑝 of the pin shaft can be calculated from the shear stress.

𝜏 =𝐹

2𝐴𝑝 (3)

Where F is the load force, 𝐴𝑝 is the cross-sectional area.

𝐴𝑝 = (𝑑𝑝

2)

2

𝜋 (4)

Where 𝑑𝑝 is the diameter of the shaft.

The size of the pin shaft is also determined from the standard dimension of the bearing. The

value of the stress was simulated and proved with Ansys, which will be later presented in the

Validation chapter.

3.5.3 Bearing

The joint includes two bearings to support the pin shaft. Its requirements are low speed and high

radial force. The required life time for the mooring system is 20 years. It is important then to

select a bearing that can also operate for 20 years. After searching and contacting with different

suppliers, the WB802-T bearing was select to used in the mooring system.

WB802-T bearings are produced by D&E Bearing [19], which was established in 1966 and

provides one of the market’s broadest and best-stocked range of slide bearings, roller bearings,

plain bearing and associated bearing products. The properties of the WB802-material, together

with the procedure of wrapping and calibration, make this type of bearing especially suitable for

constructions, where high loads and relatively slow movements are occurring. The data sheet of

WB802-T bearing can also be found in the appendixes of this thesis.

Figure 12:WB802-T plain bearing.

The selected WB802-T bearing has an inner diameter of 50 millimeter, an outer diameter of 55

millimeter and a width of 30 millimeter. It is vital to predict the life time of the bearing so that

the mooring system can be ensured to work normally over its expected lifetime. The method to

predict the bearing life is to estimate how much percentage of the thickness can be worn while

the bearing still withstands the load. By calculating the wear caused on the bearing, it is possible

to determine the life time of the bearing.

The formula calculate the wear is

𝑊 = 𝐾 × 𝑃 × 𝑉 × 𝑇 (5)

27

Where K is the specific wear rate, P is the contact pressure, V is the sliding velocity and T is the

sliding time in hours.

With the specific wear rate depending on the lubrication conditions, the wear rate for different

lubrication conditions is shown in the following table.

Table 3: The wear rate on different lubrication

Lubrication conditions Mm/(N/mm2/s.Hr) Mm/(kgt/cm

2.m/min.Hr)

Dry 3×10-3

to 6×10-4

1 to 5×10-6

Periodic lubrication 3×10-4

to 6×10-5

1 to 5×10-7

Oil lubrication 3×10-5

to 6×10-6

1 to 5×10-8

The contact pressure can be calculated through the following equation

𝑃 =𝐹𝑚𝑒𝑎𝑛

2 × 𝑑1 × 𝑏1 (6)

Where 𝐹𝑚𝑒𝑎𝑛 is the mean load in N, d1 is the inner diameter of the bearing in mm and b1 is the

width of the bearing in mm.

As the WEC system has not been fully developed yet, it is not possible to collect the force on

each numbers of the cycle. But the expected force distribution against the number of cycles is

shown in the plot.

Figure 13: Expected force distribution against a number of cycles.

According to the assumptions made in the system definition, the maximum force is 200 kN and a

minimum force is 20kN, thus the mean force is about 110 kN. But the winch design has to take

into account the harsh weather in the sea, and ultimately the harsh weather is estimated occurs on

the 20% of the total number of cycles. The mean force in this case is 150 kN.

The equation to calculate the sliding speed is

Load cycle

kN

28

𝑉 =𝑛 × 𝑑1 × 𝜋

60 × 1000 (7)

Where n is the bearing rotational speed in rpm and 𝑑1 is the inner diameter of the bearing.

In order to determine the sliding speed, the bearing rotation speed has to be calculated. It is

estimated that the mean speed of the drum is 3 meter per second and the maximum speed is 5

meter per second. There are 5 links place on the drum on a round, so the angular displacement of

a bearing is 72 degrees.

The idea is to calculate the total time when bearing is rotating and then calculate the bearing

speed by dividing the displacement by the time.

The time during which the bearing is rotating on the drum is

𝑡𝑏 =𝐷𝑑 × 𝜋

𝑉𝑑×

72°

360° (8)

Where 𝐷𝑑 is the diameter of the drum and 𝑉𝑑 is the speed of the drum.

The bearing rotates 36º in during this time, so the bearing speed is

𝑉𝑏 =(𝐷𝑏 × 𝜋)

𝑡𝑏 ×

36°

360° (9)

The number of cycles is about 80 million times, being the time in a cycle 𝑡𝑏 and assuming 20%

present of the cycles act on the rope. Therefore, the required life time of the bearing is

𝑇 = 2 × 𝐵 × 𝑡𝑏 × 80%

At the end, the designed mooring system will have a sealing solution, so the lubrication

condition is oil lubrication. The wear rate is 6×10-6

Mm, the bearing contact pressure is 50 Pa

and the sliding speed is 0.0937 meter per second.

3.5.4 Linker

The linker plays an important role in mooring system, and its dimensions are determined by the

geometry of the link and the standard dimensions of the bearing. The linker provides two degrees

of freedom for the mooring system, with two pin shafts that rotate in two directions. The bearing

are located by the linker’s shoulder and lid, and with the geometrical structure is possible to seal

the bearing in the linker. The simple geometry structure is shown in the Figure (15).

Figure 14: Drum movement.

29

Figure 15: Sketch of the linker.

3.6 Validation

A CAD model was built in Solid Edge, whereas the performance of the components was

simulated and demonstrated in Ansys. Few physical characteristics were validated.

3.6.1 Simulation of Link

In this section, it is a simulated the process of motion when the force pulls the link. The static

structural conditions are shown in Figure (16), being the applied force the peak force of 200 kN,

and a fixed support applied at the other side.

Figure 16: Static structural of link.

30

Figure (18) shows the element and node count for the model. This was the finest mesh that could

be obtained with the license available at KTH (Royal Institute of Technology). Given the relative

simplicity of the geometry, it is assumed to be enough node resolution to run a reliable

simulation.

Figure 17: Meshed link geometry.

Figure 18: ANSYS element node count for link model.

The normal stress present in Figure (19) after simulation, gives a highest value of is 364 MPa

located on the holes edge. The estimated tensile stress for structural steel is 500 MPa and a yield

stress of 250 MPa, so when stress reaches a value bigger than 250 MPa, unrecoverable

deformation will occur. The total deformation simulation is shown in Figure (20), being the

biggest deformation is 0.2 millimeter, which will not affect the link performance a lot.

31

Figure 19: Link Normal stress simulation.

Figure 20: Link deformation simulation.

3.6.1 Simulation of Pin shaft

The motion of the link that connects and pulls the pin shaft from X direction was simulated in

Ansys. The boundary conditions were set and show in Figure (21), with an applied force of 200

kN in the X direction and the seats of the bearing modelled with cylindrical supports.

32

Figure 21: Pin shaft static structural.

Figure (23) shows the element and node count for the model. As it happened before in the

simulation of the linker, this was the finest achievable mesh with the available license of the

software, and was considered to have enough resolution due to the relative simplicity of the

parts.

Figure 22: Meshed pin shaft geometry.

Figure 23: ANSYS element node count for pin shaft model

The normal and shear stresses are depicted in Figure (25). According to the simulation, the

highest value of the normal stress is 70 MPa, and from the diagram it can be seen that it has high

stress concentrators on the edge of stepped shoulder, although the stress is not too high to lead to

failure. The highest value of the shear stress is 82 MPa and it is enough to prevent the pin shaft

33

from breaking. What is more, from Figure (26), it can be found that the maximum deformation

happened on the connect area to the link and its value is 0.012 millimeter.

Figure 24: Pin shaft Normal stress analysis.

Figure 25: Pin shaft shear stress analysis.

Figure 26: Pin shaft deformation analysis

34

3.6.1 Simulation of Linker

The motion of linker holding the pin shaft was simulated in Ansys and the boundary conditions

set is shown in Figure (27). The applied force is 100 kN at Y direction and it applied at the

surface that supports the bearings. The other holes are modelled as cylindrical supports.

Figure 27: Linker static structural.

Once again, the mesh and nodal counts shown in the following pictures is the finest and most

precise that could be achieved with the available software license at KTH. The simplicity of the

parts is enough to predict that the size of the mesh is adequate and the obtained results are

reliable.

Figure 28: Meshed linker geometry

35

Figure 29: ANSYS element node count for linker model

The normal stress is shown in Figure (30) after running the simulation. It gives the highest values

of normal stress of 201 MPa acting on the holes edge. For structural steel, the estimated tensile

stress is 500 MPa steel and the yield stress is 250 MPa, so when stress value is bigger than 250

MPa unrecoverable deformation will occur. The total simulated deformation is shown in Figure

(31), and the biggest deformation is 0.1 millimeter, present on the under edge of the hole. Due to

its small value it is can be said that it will affect the link performance a lot.

Figure 30: Linker normal stress analysis

Figure 31: Linker deformation analysis

3.7 Material Fatigue

When a component is subjected to a repetitive and variable motion with load, fatigue effects will

come out. The part of the component has different level structural damage occur when in cycle

loading. Normally, the nominal maximum stress values that cause the component’s failure is less

36

than the strength of the material itself, such as ultimate tensile stress limit or the yield stress

limit.

The Wöhler curve, shown in Figure (32), is a graphical representation of the magnitude of a

cyclic stress (S) against the logarithmic scale of cycles of cycles to failure (N). The material

performance under high–cycle fatigue situations is shown in this curve.

Figure 32: Wöhler curve

The relation between the material’s stress and the number of load cycles is that the stress will

decrease with the number of load cycles increasing until five millions of load cycles. If the

material is also affected by other factors such as corrosion, temperature, residual stresses or the

presence of notches, the fatigue limit will be even smaller.

In this case, the mooring system has 80 million of repetition movements with high load. It has to

be considered the effect from the material fatigue. The current material to calculate and simulate

is structural steel, which has a yield stress of 250 MPa and a estimated tensile stress of 500 MPa.

Based on the Wöhler curve, the stress will decrease twice after a million number of cycle. After a

million of load cycles, the stress has constant value. So in order to prevent fatigue failure, the

material of the mooring system has to be replaced for another material with a higher fatigue limit

rather than steel stress.

37

4 RESULTS

In the results chapter the results that are obtained with the methods described in the method

chapter are compiled, and analyzed and compared with the existing knowledge and theory

presented in the frame of reference chapter.

Initially, from the Pugh’s evaluation matrix, the mooring system design was selected. The final

concept was a wrist joint. The required dimension has been calculated and the geometry model

was made in Solid Edge. Based on the calculation, results show that the minimum required

sectional area of the bar on the link is 0.0004 m2 and the minimum required diameter pin shaft is

0.032 meter. According to the standard dimension of the bearing and taking into account lifetime

considerations, the diameter of the pin shaft should be 50 millimeters. The detailed designed

design detail of the three parts is depicted in the following figures in this chapter. In order to

prevent the failure of the bar on the link, the rounding on the connecting edge should be 3

millimeters. The designed mooring system mainly consists of a link, a pin shaft and a linker. The

component part list is presented later in the report. The standard components present in this

design are bearings, seals and screws.

The designed mooring system has a seal solution to prevent the sea water from leaking inside the

system, being the idea to use V-shaped seals. Therefore, it will be possible to use oil lubrication

for the bearings. The estimated the limit of bearing to wear is 60% of thickness. So the expected

life time of the bearing is 8990 hours. In 20 years, the total calculated working hours are 5960

hours, so the selected bearing will be enough to operate during the whole life of the designed

mooring system. According to the simulation results, it could be proven that the designed

mooring system could work in high load condition.

The main dimensions of the link, linker and pin shaft are shown in Figure (33), Figure (34) and

Figure (35) respectively, with more detailed drawings presented in the appendixes.

Figure 33: Dimension of link.

38

Figure 34: Dimension of linker.

Figure 35: Dimension of pin shaft.

39

After a material fatigue analysis, stainless Steel 34CrNiMoS6 (SS2541) with yield strength 900

MPa and tensile strength 1100 MPa was selected as a suitable material for the system. Its values

of yield stress and tensile stress are twice bigger than structural steel.

The method used to fix the link and linker is with a key and a key slot. Figure (36) is the section

view that shows the form the key.

Figure 36: key and key slot arrangement.

The arrangement of the bearing is shown in Figure (37), with the bearing located by an extra

bearing housing and a retaining ring. The bearing‘s housing is fixed with a lid and a linker’s

shoulder. The diameter of the jack shaft is bigger than the outer diameter of the bearing, which is

a reason for the given extra bearing housing. Considering about the seal from the sea water, there

are V seals placed between the bearing and the shaft stepped shoulder and an O-ring seal

between the lid and the linker. Section B-B has a detailed diagram of the arrangement.

The installation procedure is to locate the pin shaft with the link first, and then put water into the

bearing housing. After that, it has to be pushed the bearing though the shaft and located by the

linker’s shoulder. In the end, the O-ring is placed and the screws will fasten the lid to the linker.

40

Figure 37: Bearing and seal arrangement.

An assembly drawing with balloons for the different parts is shown in the following picture.

41

4

4

12

64

3

1

84

716

2

2

54

Figure 38: Assembly diagram of the mooring system.

Table 4: Component part list

ItemNumber

Title Material Quantity Mass (Item)

1 Link Steel 2 12,584 kg

2 Pin Shaft Steel 2 2,049 kg

3 Linker Steel 1 15,110 kg

4 Bearing 4 0,535 kg

5 V seal PVC 4 0,005 kg

6 Lid Steel 4 0,274 kg

7 Screw Steel 16 0,004 kg

8 o-ring Steel 4 0,005 kg

42

43

5 DISCUSSION AND CONCLUSIONS

A discussion of the results and the conclusions that the author have drawn during the Master of

Science thesis are presented in this chapter. The conclusions are based from the analysis with

the intention to answer the formulation of questions that is presented in Chapter 3 and Chapter

4.

5.1 Discussion

This project’s purpose was to design a mooring system for the developing WEC. A seal solution

was made, but the life time of V-seal is hard to ensure. The requested life time for the mooring

system was 20 years and the designed mooring system could operate over 20 years when the

lubrication conditions for the bearing were oil lubricant. So it still needed to investigate the life

time of the V seal. Due to the fact that the V-seal can rotate with shaft, it will lack a bit of oil

lubricant. Nowadays, the developing of the technology allows a zero environment impact if a

special oil lubricant is used. However, it is little complex to assemble, since screws are used to

fix the lid.

5.2 Conclusions

The designed mooring system has two degrees of freedom, as it can roll on the drum and move

in the axial direction. It has a high load capability since the simulations validated the

performance predicted by the calculations. It is really essential to figure out the seal problem, so

that the bearing uses oil lubricant and life time is improved. However the seal solution needs a

long-term investigation. After a material fatigue analysis, it could be ensured that the material

can still work after about 80 million cycles of load. According to the calculations and

simulations results, it was proved that the life time of bearing and wrist joint is over 20 years. In

the end, the designed mooring system could be operating for 20 year with 80 million cycles of

load.

44

45

6 RECOMMENDATIONS AND FUTURE WORK

In this chapter, recommendations on more detailed solutions and/or future work in this field are

presented.

6.1 Recommendations

In this paper, the requested design for the mooring system is a chain transmission, hence it is not

needed to compare it with the other type of transmission. But that kind of comparison could be

done from other author. Due to the using the metal instead of rode, it must be cause a lot weight

and need high performance from other system to cooperate. The other concepts, rejected in this

paper, could also be interesting to consider for future developments of the system.

6.2 Future work

It is recommended that further work complements this research with the following activities:

Manufacture of the design mooring system.

Perform a physical test on the mooring system.

46

47

7 REFERENCES

[1] Clément, A., McCullen, P., Falc˘ao, A., Fiorentino, A., Gardner, F., Hammarlund, K.,

Lemonis, G., Lewis, T., Nielsen, K., Petroncini, S., Pontes, M.-T., Schild, B.-O.,

Sjöström, P., Søresen, H. C., and Thorpe, T. Wave energy in Europe: current status and

perspectives. Renew. Sust. Energy Rev., 2002, 6(5), 405–431.

[2] Thorpe, T. W. A brief review of wave energy, Technical report no. R120,

EnergyTechnology Support Unit (ETSU), A report produced for the UK Department of

Trade and Industry, 1999.

[3] Salter, S. H. Wave power. Nature, 1974, 249(5459), 720–724.

[4] Falnes, J. A review of wave-energy extraction. Mar. Struct., 2007, 20, 185–201

[5] Duckers, L. Wave energy. In Renewable energy (Ed. G. Boyle), 2nd edition, 2004, ch. 8

(Oxford University Press, Oxford, UK).

[6] Fred. Olsen http://fredolsen-energy.com/wave?WAF_IsPreview=true (accessed date:

2017-02-15)

[7] David G. Ullman., “The Mechanical Design Process”

[8] Stuart Pugh ,. “Design decision-how to succeed and know why”,

[9] DNV. Offshore Standard - Position Mooring. DNV OSE301, 2004.

[10] API. Recommended practice for design and analysis of station-keeping systems for

floating structures. API RP- 2SK, 2005.

[11] Petroleum and natural gas industries - Specific requirements for offshore structures - Part

7: Stationkeeping systems for floating offshore structures and mobile offshore units. ISO

19901-7, 2005.

[12] H. O. Berteaux. Buoy Engineering. John Wiley & Sons, New York, 1976.

[13] L. Bergdahl and N. Martensson. Certification of wave energy plants - discussion of

existing guidelines, especially for mooring design. Proceedings of the 2nd European

Wave Power Conference, pp. 114-118. Lisbon, Portugal, 1995.

[14] R. E. Harris, L. Johanning and J. D. Wolfram. Mooring systems for wave energy

converters: a review of design issues and choices. Proceedings of the 3rd International

Conference on Marine Renewable Energy (MAREC). Blyth, UK, 2004.

48

[15] J. Fitzgerald and L. Bergdahl. Considering mooring cables for offshore wave energy

converters. Proceedings of the 7th European Wave and Tidal Energy Conference. Porto,

Portugal, 2007.

[16] J. Fitzgerald and L. Bergdahl. Including moorings in the assessment of a generic offshore

wave energy converter: A frequency domain approach. Marine Structure; Vol. 21, pp. 23-

46, 2008.

[17] L. Johanning, G. H. Smith and J. Wolfram. Interaction between mooring line damping

and response frequency as a result of stiffness alteration in surge. Proceedings of the 25th

International Conference on Offshore Mechanics and Arctic Engineering (OMAE), Paper

No. OMAE06- 92373. Hamburg, Germany, 2006.

[18] L. Johanning, G. H. Smith and J. Wolfram. Measurements of static and dynamic mooring

line damping and their importance for floating WEC devices. Ocean Engineering; Vol.

34, pp. 1918-1934, 2007.

[19] D&E bearings http://debearings.com/ (accessed date: 2017-03-25).

49

APPENDIX A: SupplemEntary INFORMATION

Standard keyway and key sizes

Unless otherwise specified, the shaft keyway is assumed to be standard. A list of standard

keyway and corresponding key sizes for shafts are listed below in Table . The common

specification dimension, Keyway Size, is highlighted.

50

51

APPENDIX B: DATASHEET OF BEARING

Technical data

Material: Homogeneous bronze Cu 91.3%, Sn 8.5%,P

0.2% Yield Point: (Rp0.2) ca 300N/mm2

Tensile strength: (Rm) ca 450 N/mm2 Hardness: ca 125-150 HB

Friction: 0.08-0.25μ Max speed: 2.5 m/s

Temperature range: -100/ +200 ˚C

Tolerances: Bearing pressed into housing H7 get tolerance H9. Recommended tolerance for the shaft IT 7 or IT 8with position of tolerance e or f.

Lubrication: Additional lubrication ought to be done through for the shafts or radial through the housings.

Benefits

High load capacity Intended to working in difficult and dirty conditions specially.

Good lubrication properties due to lubrication holes. Has high level thermal conductivity.

Wide range of stock holds standard dimensions. Optimal lubrication intervals.

Complete solution with Intergard seal.

Special:

Bearing with one seal.

In-or outside lubrication grooves. WB800 with seals.

Other positions of tolerance.

52

53

APPENDIX C: WATER V SEAL

Figure 39: Water V seal.

Specification

Standard or Nonstandard : Standard

Material: Rubber Brand Name: AUTOX Temperature: -35˚C--110˚C

Spring and punched part : Stainless Steel Pressure: ≤ 40 MPa

Feature: Good sealing/ resistant friction Style: Mechanical seal

Model Number: VA Speed: ≤ 30 m/s Medium: Water/ air

Size: Kinds of Application: Rotary rod and the surface of bearing

cover

54

55

APPENDIX D: Detail drawing of mooring system

The Linker drawing was scale 1:2, scale value 0.5 and made by Solid Edge.

125

37 125

44

5

44

20

7,5

90°

O 75

O 73

O 70UH7

O 65UH7

2

12,5

O 5

110

37

19,5

239,5

56

The pin shaft drawing was scale 2:1, scale value 2 and made by Solid Edge.

60UH

9 O

4,62

14

50UH

7 O

R3

OR2

O

35O

117O

41O

1.6 1.6

57

The link drawing was scale 1:2, scale value 0.5 and made by Solid Edge.

35

90

R20

R20

60O

24,5

R 3

14

4,62

180

400

40

110

R 30

UH7

14

45