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
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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
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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
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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
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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
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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
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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
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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.
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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.
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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.
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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].
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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.
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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.
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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.
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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.
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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.
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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
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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
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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.
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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.
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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