Martin_Ness_Bachelor_Thesis

Resolving gasket- and bearing
failure in a TBV96S valve
Bachelor thesis conducted at Stord/Haugesund university college
Mechanical engineering - Design of Marine Structures
Mart´ın Ness Candidate 13
Helene F. Skage Candidate 02
Haugesund Spring 2016

University college
Stord/Haugesund
Study for engineering
Bjørnsonsgt. 45
5528 HAUGESUND
52 70 26 00
Thesis title:
Resolving gasket- and bearing failure in a TBV96S valve
Written by: Submission date:
Mart´ın Ness & Helene F. Skage May 4, 2016
Profile: Study:
Mechanical engineering - Design of maritime structures Mechanical
Supervisors: Grading:
Einar A. Kolstad & Vladimir Mitin Open
Abstract
This report is written for Valves of Norway and addresses the selection process for
gaskets and bearings for the failed valve TBV96S. In addition a mount is designed
and optimized for this specific usage. The changes made on the gasket and bearings
will effect some components and these will be stressed analyzed with both hand and
computer aided calculations and recommendations for optimization is suggested based
on the result.

Preface
In the final semester of the bachelor degree for the students at Stord/Haugesund University
College they are required to write a thesis. The thesis will treat issues related to the
fields that have been lectured trough the whole bachelor degree. The thesis is written
for a company with the intention to show that the students can use new information in
combination with existing knowledge to plan and execute an independent task, formulate
questions and analyze the information at hand to answer these.
The task planned along with the firm involves choosing materials for both bearings and
gaskets, stress analyzing and additionally designing a mount for the TBV96S valve. This
will give the students challenges in several areas. The students will have to get both a
theoretical and practical understanding of the subjects at hand, and furthermore applying
their creative skills to the task.
The courses which are especially applied for this task is:
• Mechanical Design I and II
• Statics and Strength of Materials
• Process Technology I
• Materials and Manufacturing
• Underwater Technology and Hydraulic Systems
A sincere appreciation and gratitude goes to the external supervisor from Valves of Norway,
Vladimir Mitin, and the internal supervisor at Stord/Haugesund University College, Einar
Arthur Kolstad.
Mart´ın Ness Helene F. Skage
i

Contents
Preface i
Glossary vii
Summary viii
1 Introduction 1
1.1 Background for the thesis . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1
1.2 About Valves of Norway . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1
1.3 The goal for the thesis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2
1.4 Methodology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2
1.5 Software . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3
1.6 Limitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3
1.7 Regulations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5
1.8 Finite element method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6
1.9 Ball valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8
2 Gasket 10
2.1 Seal type . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11
2.2 Material . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11
2.3 Solution . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13
3 Bearings 14
3.1 Material . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15
3.2 Reaction forces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16
3.3 PV factor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18
3.4 Selection from supplier . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19
4 Calculations 21
4.1 Friction torques from components . . . . . . . . . . . . . . . . . . . . . . . . 24
4.2 Stress Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27
4.3 Finite element analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35
4.4 Stress on bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36
4.5 Conclusion stress calculations . . . . . . . . . . . . . . . . . . . . . . . . . . 37
5 Valve Actuator 38
5.1 Types of actuators . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38
5.2 Commonly used mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . 40
5.3 Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 42
5.4 Alternative design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 50
6 Discussion 52
7 Conclusion 54
Bibliography 56
ii

2D Drawings i
Appendices xxvi
A Tolerances from ISO-2768 . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxvi
B Bearing material specs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxviii
C FEROGLIDE dimensions . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxix
D AST formulas . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxx
E Previous 2D assembly drawing . . . . . . . . . . . . . . . . . . . . . . . . . xxxi
F Metric tap- and clearance drill dimensions . . . . . . . . . . . . . . . . . . . xxxii
G 316 Stainless steel properties . . . . . . . . . . . . . . . . . . . . . . . . . . xxxiii
H Friction coeﬃcient for stainless steel bolts . . . . . . . . . . . . . . . . . . . xxxiv
I Torque Constant K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxxiv
J Email regarding PEEK . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xxxv
K Calculations torsion, shear and moment. . . . . . . . . . . . . . . . . . . . . xxxvi
L Calculation method for distributed loads . . . . . . . . . . . . . . . . . . . . xlvi
M FEM report for the Bearing forces . . . . . . . . . . . . . . . . . . . . . . . xlviii
N FEM report for mount with bend . . . . . . . . . . . . . . . . . . . . . . . . l
O FEM report for closed valve . . . . . . . . . . . . . . . . . . . . . . . . . . . liii
P FEM report for open valve . . . . . . . . . . . . . . . . . . . . . . . . . . . lvii
List of Tables
2.1 Fluoropolymer materials for gaskets . . . . . . . . . . . . . . . . . . . . . . 12
3.1 Material for bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15
3.2 Data for calculating PV factor . . . . . . . . . . . . . . . . . . . . . . . . . 18
3.3 FEROGLIDE static load capacity . . . . . . . . . . . . . . . . . . . . . . . 19
3.4 FEROGLIDE dynamic load capacity . . . . . . . . . . . . . . . . . . . . . . 19
3.5 FEROGLIDE bearing dimensions . . . . . . . . . . . . . . . . . . . . . . . . 20
4.1 Friction torque on the bearings . . . . . . . . . . . . . . . . . . . . . . . . . 24
4.2 Diameters of circular sections . . . . . . . . . . . . . . . . . . . . . . . . . . 27
4.3 Torsional stress with closed and open valve . . . . . . . . . . . . . . . . . . 29
4.4 Shear stress . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31
4.5 Lengths . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32
4.6 Moment stress . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
4.7 Von mises stress . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34
4.8 Utilization with closed and open valve . . . . . . . . . . . . . . . . . . . . . 34
4.9 Stress from FEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35
4.10 Utilization from FEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35
4.11 Stress on bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36
4.12 Utilization on bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36
iii

List of Figures
1.1 Mesh refinement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7
1.2 The main components of the TBV96S Valve. . . . . . . . . . . . . . . . . . 8
2.1 The damaged gasket . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10
2.2 Illustration of old and new gasket solution . . . . . . . . . . . . . . . . . . . 13
3.1 The two main types of bearings . . . . . . . . . . . . . . . . . . . . . . . . . 14
3.2 Free body diagram . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17
3.3 FEROGLIDE wear and friction graphs . . . . . . . . . . . . . . . . . . . . . 20
4.1 Illustration of the directions when the ball is in open position . . . . . . . . 21
4.2 FEM setup . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22
4.3 Notations used in calculations . . . . . . . . . . . . . . . . . . . . . . . . . . 23
4.4 Illustration of the components used on the stem . . . . . . . . . . . . . . . . 23
4.5 Torsion diagram . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28
4.6 Shear force diagram . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30
4.7 Moment diagram . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32
4.8 Result stress analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37
5.1 Illustration of how the fluid actuator work. . . . . . . . . . . . . . . . . . . 39
5.2 Illustration of the scotch yoke. . . . . . . . . . . . . . . . . . . . . . . . . . 40
5.3 Illustration of the sliding crank. . . . . . . . . . . . . . . . . . . . . . . . . . 40
5.4 Illustration of the lever solution. . . . . . . . . . . . . . . . . . . . . . . . . 41
5.5 Illustration of the rack and pinion mechanism. . . . . . . . . . . . . . . . . 41
5.6 Sketches of concepts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43
5.7 Render of mount - version one . . . . . . . . . . . . . . . . . . . . . . . . . . 44
5.8 Render of mount - final version . . . . . . . . . . . . . . . . . . . . . . . . . 45
5.9 Illustration of a and b . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 46
5.10 Von mises FEM result for mount . . . . . . . . . . . . . . . . . . . . . . . . 47
5.11 Preload and friction force . . . . . . . . . . . . . . . . . . . . . . . . . . . . 49
5.12 Illustration of the circular gear mechanism . . . . . . . . . . . . . . . . . . . 50
5.13 Illustration of the linear gear mechanism . . . . . . . . . . . . . . . . . . . . 51
5.14 Illustration of the ”wrench” mechanism . . . . . . . . . . . . . . . . . . . . 51
iv

List of Variables
Variable Description Unit
D Outer diameter mm
d Inner diameter mm
p Pressure N/mm2
A Area mm2
I Moment of inertia mm4
W Section modulus mm3
L Length mm
P Load capacity N/mm2
V Velocity m/s
F Applied load N
c Cyclic velocity cpm
θ Oscillating angle rad
k Spring constant N/mm
l0 Length at start mm
l1 Length at end mm
µ Coeﬃcient of friction
β Seat incline °
n Amount
γ Safety factor
σy Yield stress N/mm2
σt Tensile strength N/mm2
σp proof strength N/mm2
σa Allowed stress N/mm2
τ Shear stress N/mm2
P Pitch mm
N Bolt count
K Bolt constant
As Tensile stress area mm2
Fi Pre-tension force N
J Torsional constant mm4
v

List of Subscripts
Subscripts is the lower letters behind a variable to describe further what exactly is calculated.
These subscripts are often self explanatory since the subscripts and the abbreviation of the
part names often are the same. Many subscripts indicates positions on the valve, these are
shown in ﬁgure 4.3 on page 23.
Table: Subscripts used in the report.
Subscript full form Refers to the:
g gasket Gasket placed in the seat
b ball Hollowed out sphere in the middle of the valve
pb pressure ball Pressure working on the ball
ls lower stem Position with the lower stem bearing
us upper stem Position with the upper stem bearing
t trunnion Position with the upper stem bearing
h hexagon Hexagon section on the stem
lc lower circular Lower circular section on the stem
gr glide ring Position of the top glide ring on the stem
tb thrust bearing Position of the thrust bearing on the stem
sg spring gasket Force from the springs on the gasket
sh stem hole Stem hole diameter
vi

Glossary
alloy A mixture of metals or a metal and another element. 15, 34
brine A solution of salt in water. 10
CAD Computer-aided design. 3
CAE Computer-aided engineering. 3
CAM Computer-aided manufacturing. 3
centroid Center of a geometry, equal to center of mass when uniform density. 31
neutral axis An axis in the cross section of a shaft/beam along which there are no
longitudinal stresses or strains. 33
oscillatory motion The repeated motion in which an object repeats the same movement.
18, 38
outer most ﬁber The material ﬁber with the longest distance to the axis. 33
peripheral speed The distance a given point on the perimeter of a rotating circular
object travels, expressed in m/s. 18
pivot point The center point of any rotational system. 41
polymer Polymers are made up of molecules all strung together to form long chains or
structures. 11, 12
projected area The rectilinear parallel projection of a surface of any shape onto a plane.
25, 36
self-lubricating Oil-impregnated sintered metal or plastic to provide lubrication on
friction. 14, 15
static load Any load that does not change in magnitude or position with time. 14
tensile strength The maximum stress a material can handle before failure due to breaking.
26
von mises It is widely used by designers to check whether their design will withstand
a given load condition. It is part of a plasticity theory that applies best to ductile
materials, such as metals. 34, 47
yield strength The stress level where a material reaches plastic deformation. 28, 34, 48
vii

Summary
This report is the bachelor thesis in the mechanical engineering study with specialization
in Design of Marine Structures, and addresses bearing and gasket failures with the valve
TBV96S which is constructed by Valves of Norway. The task is given by Valves of Norway
a valve producing company located by the west coast of Norway, outside Bergen.
Related to the gasket and bearing problems, the assignment also is to find the reason for
the failures and the modifications needed to avoid these. The components affected by the
changes will be stress analyzed with both hand and computer aided calculations. Lastly a
mount will be designed for this specific valve.
The TBV96S valve is used for mud and has a problem with leakage between the seat and
the ball inside the valve. The reason for the leakage was found to be that the hardness
rating of the gasket was to low to handle the corrosion from the hard particles in the mud.
This was found through examining the failed valve, and researching the materials used for
gaskets. The material that is recommended for further use in the valve is PEEK.
The damaged bearings on the stem and trunnion is found to be overloaded from both the
work pressure and from the actuator force. A new material is assigned to the bearings,
a woven PTFE reinforced and bonded to metal backing from TENMAT called FEROGLIDE.
A maximum actuator force is then found through several finite element analysis. Further a
mount is designed to counteract the friction torques generated by the gasket and bearings.
The mount is designed for simplicity and low cost. It will convert the linear motion from the
actuator to a rotational motion on the stem. The design is presented using 3D renderings
and 2D drawings.
In case of a situations where a higher safety factor or a smaller arm is needed several
proposals for mounts that will cancel out the actuator force are suggested.
viii

1 Introduction
1.1 Background for the thesis
The oil and gas industry is a large part of the Norwegian industry. The oil and gas industry
started with the discovery of the oilfield ”Ekofisk” in 1969. Due to high global development
in the oil- and gas industry a large number of platforms and subsea production systems
are built.
These production systems contains a large amount of pipelines for transport and control of
the well contents. To control the flow inside these pipelines ball valves is often used. A ball
valve works by rotating a hollowed out sphere. When the hole is inline with the flow it is
open, and closed when perpendicular to the flow. The purpose of the valve is not to adjust
the flow and should only be used for shutoff or fully opening the flow. At subsea the valve
is often controlled by a remote controlled actuator, where the actuator is either fluid, air
or electrically powered and controlled. In some cases where operations of the valve is not
done on a regular basis, it can be operated by an ROV (remotely operated vehicle).
Valves at subsea must be robust and maintenance free were possible. This is because
of a high cost in ROV maintenance operations and/or replacement of valves. Valve of
Norway have a ball valve that failed under usage and the group has been tasked to make
the necessary modifications to prevent this in further use.
1.2 About Valves of Norway
Valves of Norway was founded in 1987 by engineers with experience from the North Sea oil
sector. It started with maintenance and repairs on valves but they developed to a valve
manufacturer. They are located at the rim of the North Sea, ˚Agotnes outside Bergen, and
have around 20 employees.
Valves of Norway is a manufacturer of standard and customized valves in a variety of
materials. They produce ball, gate, check, globe, double block and bleed/DBB/combination-
/modular valves. They are ISO 9001 and PED (European pressure equipment directive)
certified.
They have production facilities and a machine park in their company building. All valves
are manufactured and assembled locally. Every part of the valve is made in-house and this
gives them control over quality and planned lead times. The company’s main goals are
to strengthen their position world-wide and keep their reputation as an innovative valve
manufacturer.
Valves of Norway is in the process of changing their company name to Norvalves. They
will still be refereed to as Valve of Norway since this transition started after the assignment
was given.
1

1.3 The goal for the thesis
The ball valve at hand had defects on both the bearings and gaskets due to faulty assembly
and wrong usage. The valve contains three bearings one at the trunnion, and two at the
stem. These take up forces from the pressure and the actuator. The gasket is placed
around the ball in the valve, preventing leakage. The goal with this thesis is to modify the
valve to meet the requirements of the application it is used in.
This will be accomplished by identifying the reasons for the failures and find the necessary
changes needed. An overview of the materials is then presented, compered and the most
suitable material will then be suggested. The material and dimensional changes will affect
several components of the valve. These will be analyzed for stress with both hand and
computer aided calculations. Since this is dependent on the actuator force, which is not
defined, the maximum allowed actuator force is therefore found.
In addition to the tasks above. A new actuator mount that can manage the forces and
moments is to be designed for the same valve. First the limitations is set, and information
about the mechanisms on the market today are found, discussed and compared. Further
a design will be built on top of this research. This will also be 3D modeled and stress
analyzed with both hand and computer aided calculations.
1.4 Methodology
The students got a tour at Valves of Norway’s facilities so that they could get an overview
of the way valves were manufactured, which equipment the company had and also a better
understanding of the manufacturing process. The valve and its parts was examined, which
helped define the main cause of the valves failure and to give a better picture of how the
valve functioned. Research was done, firstly with a focus on the material choice for the
gasket and bearings and later the actuator and mount. Companies were contacted about
the gasket material and dimensions to get their point of view.
Some meetings were held between the group and the external sensor Vladimir Mitin, to
discuss around the calculations of the forces and friction forces in the thesis. This was
formulas that was not documented in a clear way and not found in known academical
books or the web.
Finite element analysis was then used in combination with hand calculations to find the
most accurate solution. This was done on the stress calculations for the trunnion, stem,
bearings and the actuator mount. The solution was then rendered and 2D drawings were
made.
2

1.5 Software
The main programs used to produce this thesis was overleaf, Autodesk inventor and fusion
360. Several support programs were also used including Power Point, Keyshot, Photoshop
and Soulver.
Overleaf
Overleaf is a collaborative writing and publishing system that makes the whole process
of producing academic papers much quicker for both authors and publishers. Overleaf is
an online collaborative LaTeX editor with integrated real-time preview. LaTeX is based
on Donald E. Knuth’s TeX typesetting language. LaTeX was first developed in 1985 by
Leslie Lamport. It is a comprehensive and powerful tool for scientific writing and it is not
a word processor. Instead, LaTeX encourages authors focus on the content and not the
appearance of their documents. Overleaf was used to write the whole report.
Autodesk
Autodesk is one of the worlds leading companies in 3D design, engineering- and entertainment
software and has been on the market since 1982. Autodesk software are used to design,
visualize, and simulate ideas before they’re built. Autodesk inventor is a 3D parametric
history-based computer aided design program. It is used for creating 3D digital prototypes
used in design, visualization and simulation of products. In this report it was used for the
2D drawings.
Autodesk Fusion 360 is an all in one tool with CAD, CAM and CAE. It is a multi-platform
cloud-based program making it easy for collaboration. The simulation part of the program
gives the the ability to predict how the product will handle the environments and situations
it is designed for before it is produced. Creating virtual prototypes can reduce both cost
and time in the prototyping phase. In this report Fusion 360 is used for 3D modeling and
the finite element analysis done on both the valve and mount.
1.6 Limitations
Some factors are given in advance for the valve. The fluid type running through the valve
is mud. Working pressure in the valve is 10 MPa, and the temperature is between 0°C and
70°C. The rotational speed and rotation angle of the ball and stem is 15 cycles/min and
90° respectively.
In the report several measurements needed from the valve is taken from the 3D model
made for the previous version of the valve. This is provided by Valves of Norway, see
appendix E for the 2D assembly drawing. It is assumed that all the information from the
sources referred to in the report are reliable and accurate. All abbreviations, symbols and
formulas that are not mentioned in the list of variables and subscripts, are believed to be
generally known or known to the target audience of this thesis.
3

Gaskets
The company will keep their seats that are made from the material A276 S23750, a super
duplex stainless steel. The materials that will be examined are the one that are most used
in gaskets. The gaskets are made at valve of Norway. Valve of Norway prefers if the design
is changed from two to only one gasket in each seat. The dimensions of the gasket is given
by the company, see section 2.3.
Bearings
It is desired that the changes in bearing dimensions are as small as possible such that the
modification of the trunnion, ball and stem is kept at a minimum. The company have a
preferred bearing supplier which is TENMAT. These will be used were possible. As stated
above, the rotational speed are set to 15 cycles/min. This is a maximum and will in reality
most likely be much lower since repeatedly operations will happen on a rare basis. The
bearings must be designed to be maintenance free, and wear resistant in corrosive and
highly frictional environments.
Calculations
The safety factors used in the calculations are found in the NORSOK standard, ISO 10423
to be 1.5 for pressure containing equipment and 1.2 for the bolts. ISO 10423 is further
discussed in section 1.7. The friction factor from trunnion and lower stem bearings are set
to 0.1 because they partly run in mud. The loads from the bearings will be calculated as
point loads. The actuator force is calculated as a perpendicular force to the flat surface on
the stem. An assumption was that the deflection is at such a small scale in comparison to
the tolerances that the thrust bearing would not take up any moment. The actuator mount
height on the stem is adjustable but is assumed to be placed at the middle of the square
section of the stem. Stresses will be calculated for both the open and closed position, it is
assumed that these will generate the maximum amount of stress. Each position experiences
different forces and torques acting on the components.
Actuator mount
Valve of Norway will not make the actuators, but will produce the mount in their machine
park. There is an assumption that the machine park can produce the design that are
proposed in the report. The report only considers solutions with one actuator mounted to
the valve. Solutions with several actuators exist on the market, but this thesis will focus on
a single actuator driven solution. Some assumptions are made for the calculations, these
includes what height the mount is on the stem. This is assumed to be 40 mm from the
top, which is in the middle of the square section of the stem.
4

1.7 Regulations
A standard is an agreed way of doing something. That be making of materials, products,
managing or delivering a service. It is a document that provides the necessary guidelines,
requirements, specifications or characteristics to make sure the quality, safety and reliability
is sufficient. All tolerances are found in ISO 2768-f in appendix A if not otherwise specified.
Iso standard
The International Organization for Standardization (ISO) has developed standards since
1947 for the majority of the sectors. Norway collaborates on several important areas in the
ISO standards. Most of the ISO standards contains technical specification or guidelines to
secure the quality and strength of materials, products, processes and services. Each ISO
standard is often concentrated on a small and specific area.
ISO 13628-4
ISO 13628-4: Petroleum and natural gas industries - Design and operation of subsea. Part 4:
Subsea wellhead and tree equipment. It contains specifications regarding subsea wellheads,
mudline wellheads, drill-through mudline wellheads and both vertical and horizontal subsea
trees. It also contains guidelines in design, materials, welding, quality control, marking,
storing and shipping. Section 7.10 in ISO 13628-4 with title: Valves, valve blocks and
actuators, is the only section that has relevance for this thesis. This mostly refers to the
ISO 10423 standard.
ISO 10423
ISO 10423: Petroleum and natural gas industries - Drilling and production equipment -
wellhead and christmas tree equipment. It gives recommendations for technical specifications
in regards of dimensions, functionality, design, materials, testing, inspection, welding,
marking, handling, storing, shipment, purchasing, repair and re manufacture of wellhead
and christmas tree equipment for use in the petroleum and natural gas industries. The
safety factors given in the standard is 1.5 for all pressurized equipment and 1.2 for the
bolts.
5

1.8 Finite element method
The mathematics of the finite element analysis is complex and will not be explained in
detail for this thesis, but the methods will be mention for further reading. In this thesis
finite element analysis is used to find the reaction forces in the original and new version of
the valve. For the new version both the closed and the open position in addition to the
mount will be analyzed.
The finite element method (FEM) also referred to as finite element analysis (FEA), is
a numerical technique for finding approximate solutions to boundary value problems for
partial differential equations. Finite meaning that it is a limited amount of elements. It is
used to build predictive computational models of real-world scenarios. The method was
first proposed in 1943 by Richard Courant, but did not gain attention before the 1950s and
1960s [1]. It has since been researched, tested and improved upon and is now an accurate
and reliable method to use for calculating complex problems in several fields.
The finite element analysis cuts/divides a structure into several elements, it then reconnects
the elements at connection points known as ”nodes”. The nodes work as pins or drops
of glue that sticks the elements together. It uses the concept of piece wise polynomial
interpolation to simplify the problem. Furthermore the software uses calculations such as
matrix manipulations, numerical integration and a set of simultaneous algebraic equations
at the nodes to estimate the solution.
The main advantages of finite element analysis over hand calculations is that it can handle
very complex geometry, loads and constraints in a wide range of engineering fields. The
disadvantage is that it gives a closed-form solution, meaning that parameters can not
be changed in post process to examine the affect each parameter has on the system. It
also only approximates solutions. In FEM user error is fatal, this is also true for hand
calculations, but there they are more easily noticed. FEM also have ”inherent” errors such
as simplifying the geometry. Finite element is used for problems such as thermal, fluid flow,
electrostatics, elastic and many more. In this case it is used for elastic problems which will
be the process in focus.
Process
Before starting the analysis, some factors needs to be defined. These are the loads,
constraints, contacts and mesh. The order these are set is not important. In addition the
material properties must be defined to minimize the hand calculations.
A load causes stresses, deformations, and displacements in components. It can be a force,
pressure, moment, torsion or a remote force. A remote force is a force that is not on a
component in the analysis but still affects it. This may be used to reduce computing time
if the stress analysis for some parts are unnecessary. The force from gravity can also be
included if the mass is of a significant size.
Contacts defines which relationship the bodies or components have to each other. There
are several contact types to chose from, bonded, separation with no sliding, sliding with no
separation or separation and sliding. Bonded is used where the components are rigid in
6

relationship, for example where the components are welded or bolted together. Separation
is when components can as the name implies separate from each other. Sliding is when the
components can move freely along a surface but have continually contact with it, like a
bearing on a shaft.
Constraints are limitations or restriction of movement. A combination of contacts and
constraints will define the degrees of freedom for components. Constraints can be frictionless,
pinned, displaced and fixed supports. Frictionless supports allow movement in the plane of
the face, but restricts movement normal to it. Pinned is used to prevent displacement of
cylindrical surfaces it can be fixed in the radial, axial and/or tangential direction, pinned
constraint is as an example used on bearings. Fixed constraints fixes a face or component
in place in the x, y, and/or z-direction, this is used when a component is fixed to a non
moving support.
A mesh is the name of the wire frame that the lines between the nodes make up. The
density of the mesh directly correlates with the accuracy of the result. A more dense mesh
will be a better approximation of the real geometry and on that basis give more accurate
results. The downside with a larger mesh is the addition in computing time needed to
calculate the solution. This is why mesh refinement is used. Mesh refinement is where the
mesh starts out coarse, and gets gradually increased to a point were further increase in
density have an insignificant increase in precision. In fusion 360 there is also a feature
called adaptive mesh refinement, this automates the process. It is set to refine the mesh
until the von mises stress result converges for each section of the object, meaning that
the result do not differentiate above a certain percentage from the previous result. This
feature will give a high density mesh where there is a larger uncertainty and a low density
were there is not. This will increase the precision without the large increase in computing
time. An example of this is shown in figure 1.1 where the concentration of mesh is greatest
at the stress peaks.
Figure 1.1: Mesh refinement
The result can now be checked for values that are out of the ordinary. Further a variety
of display options can be chosen to visualize the impact including animations. Displays
include but are not limited to safety factors, normal stress, shear stress, displacement
and von mises stress. These results can be used to better understand how the part work
under stress, and to see which component or feature that should be modified for better
optimizing.
7

1.9 Ball valve
A ball valve is used to permit flow across the pipeline. It has an inlet and an outlet with a
hollowed-out sphere/ball in the middle. When the hole is inline with the flow it is open and
closed when the hole is perpendicular to the flow. There is a slot machined in the top of
the ball which the stem fits in. The main components of a ball valve is shown in figure 1.2.
Figure 1.2: The main components of the TBV96S Valve.
The stem is the rotational loaded axle which opens and closes the valve. It connects to the
top of the ball and sits in the valve body which consists of two parts, the house that holds
all the components and the valve lid. The valve lid is the part that closes off the valve and
is bolted to the housing.
8

The seat is a metal ring which contains an inserted gasket in the front against the ball,
and o-rings against the body. The gasket is placed in a groove large enough to firmly hold
it in place.
Ball valves can be either full bore or reduced bore. In full bore ball valves the passage is
large enough to pass flow without reducing. This means that the inner diameter of the
bore is the same as the inner diameter of the pipes the valve is fitted to. Reduced bore has
as the name implies a reduced bore hole which also reduce the valve size and cost. These
are used when the pipes are not operating at full capacity.
There are two types of ball valves. Float and trunnion. In floating ball valves the ball is
supported by the stem and moves against the downstream seat when pressurized. The
higher the pressure the tighter the seal between the ball and seat. Floating ball valves
has seals in the stem to stop the fluid leaking at the top in addition to a seal in the valve
lid pressing against the housing. These have low weight and small cost because of the
simplicity. When the valve size increases the seats cant support the forces from the ball
and it is necessary with a new solution.
A trunnion ball valve has an additional mechanical anchoring of the ball at the bottom.
This makes the stem, ball and trunnion act as a single assembly and the bearings on both
the stem and trunnion does not allow for lateral movement. The bearings will therefore
take up all the force from the pressure on the ball. The seat will be pressed against the
ball with both a force from the work pressure on the back of the seat and several springs
working on the seat. This combined force will seal the valve. The trunnion ball valve have
in addition to the seals in the floating ball valve, seals in the trunnion to stop the fluid
leaking out the bottom. The advantage with the trunnion ball valve is that it can be scaled
up in both size and pressure. It also has a lower operating torque which reduces the size of
the actuator needed.
TBV96S
The TBV96S valve is a full bore trunnion ball valve. It has a house and valve lid that is
made of A479 UNS S31600, a molybdenum-containing austenitic stainless steel. As is the
trunnion, trunnion lid and the stem lid. The stem is made of the wrought nickel-base super
alloy steel UNS NO 7718, and the ball and seat is made of A276 UNS S32750, a super
duplex stainless steel. It is designed for use in a subsea environment, with a maximum
depth of 1000 m. The main components is shown in figure 1.2.
9

2 Gasket
Gaskets are an integral component of any device which requires the confinement of a gas or
liquid. They compensate for the irregularities of mating surfaces and creates a seal between
two mechanical components. It has to maintain the seal in the intended environment
and its conditions. The conditions are the pressure, temperature, fluid type and surface
roughness. The gasket is pushed by a force and squeezed such that the gasket surface
adapts and fills the imperfections in the mating surface.
To create a working seal three factors must be done correctly. The force must be sufficient
to press the gasket as mush as needed. The force must maintain a large enough internal
stress to ensure that the gasket will be in contact with the mating surface when pressure is
applied. Lastly the material must be chosen such that it will withstand the stress from the
pressure, temperature range it is affected by and the corrosive or eroding attack of the
fluid.
The fluid in this case is drilling mud. Drilling mud is a highly viscous fluid mixture that
is added to the well-bore to carry out cuttings, lubricate and cool down the drill bit.
The hydro-static pressure of the drilling mud also helps to prevent the well-bore walls in
collapsing, controlling pressure and provide buoyancy. It is water, brine, oil or synthetic
based. Water based fluids are used more in less demanding drilling jobs at medium depth.
Oil based are used in greater depths, and where the environmental concerns are important
the synthetic based is used.
In this valve the failing gasket is the one between the seat and ball. On the basis of the
relative small pressure and temperature size in the valve, it is assumed that the reason
for the leakage is eroding attack from the fluid. The mud is a corrosive fluid containing
hard particles which will erode a softer gasket at a fast phase. Figure 2.1 shows a picture
of the failed gasket. This shows that the inner gasket is ”eaten” up” which supports the
assumption. For this reason the gasket material is changed from the previously material
TFM with a hardness of around 59 to a new one.
Figure 2.1: The damaged gasket
10

2.1 Seal type
Metal to metal seals are used where the use of elastomeric and polymer seals are not an
option due to high wear, radiation, very high pressure, high vacuum or high temperature
in the application. It is especially useful where there is a need for fire-safe equipment. To
create a metal sealing often a tungsten carbide coating or electroless nickel plating is used
on the metal. The tolerances have improved greatly and the leakage rate for the ANSI
standards is down to an impressive 0.01% leakage of full open valve capacity. The major
down side with metal to metal seals is its cost.
Soft insert seal are made of softer and more flexible materials called elastomers. The
group of elastomers includes materials which can be stretched and bent without exerting
great force. Once the deforming force relaxes or no longer acts at all, the parts take their
original shape. The term elastomers, which is derived from elastic polymer, is often used
interchangeably with the term rubber. Elastomers can be made in varying degrees of
hardness, but are sensitive to light, ozone, high temperatures, a large number of fluids
and chemicals, and wear [2]. It is easier to obtain a near zero leakage for softer seals, but
they will be more vulnerable for corrosive mediums, chemical mediums, temperature and
pressures.
Hard insert seal are a middle ground of the two above. This solution uses a thermoplastic
ring inserted into the seat. The material used are mostly in the fluoropolymer group.
A fluoropolymer is a polymer that contains molecules of carbon and fluorine. They
are high-performance plastic materials used in harsh chemical and high-temperature
environments, primarily where a critical performance specification must be met.
The valve has a working pressure of 10 MPa and a temperature range from 0°C to 70°C
which all seal types can handle. The leakage rate allowed is not specified and is therefore
not assumed to be the highest priority. Mud is therefore seen as the main factor. It has
particles with a high hardness rating, meaning a high resistance of plastic deformation by
indentation, also referred to as the resistance to scratching/wear. The specific hardness is
unknown because of the variety of mud compositions but it is assumed that the metal to
metal seal is not needed. Hard insert is therefore chosen for its strength yet lower cost.
2.2 Material
In table 2.1 some of the most known fluoropolymers are compared to each other with focus
on the hardness shore [3–5]. Be aware that the hardness shore is an approximated value
because of small differentials depending on the volume of the materials mixed within the
polymer. This depends on the supplier and the number is only there to give an estimate
of the polymers hardness. The temperature range for the materials is also documented.
They are found at Boedeker Plastics webpage [6] and shows that all the fluoropolymers
withstand the required operating temperature range.
11

Table 2.1: Fluoropolymer materials for gaskets
Material Information Hardness
Shore
PTFE
(Polytetra -
fluoroethylene)
or TFE
(Teflon®
)
Is a fluorocarbon based polymer and typically is the
most chemically-resistant of all plastics while retaining
excellent thermal and electrical insulation properties.
TFE also has a low coefficient of friction so it is
ideal for many low torque applications. Although
TFEs mechanical properties are low compared to other
engineered plastics, its properties remain useful over a
wide temperature range (−73°C to 204°C)
56
TFM Is a modified second generation TFE polymer that
maintains the chemical and heat resistance properties
of first generation PTFE. It has a denser polymer
structure than standard PTFE with better stress
recovery. The temperature range is similar to PTFE
(−73°C to 204°C).
59
PCTFE
(polychlorotri-
fluoroethylene),
Kel-F®
or
Neoflon®
It combines physical and mechanical properties, non
flammability, chemical resistance, near-zero moisture
absorption and excellent electrical properties not found
in any other thermoplastic fluoropolymer that also
performs well in a wide temperature range (−240°C to
204°C).
75-85
PEEK (Poly
ether ether
ketone)
Is a high temperature, high performance engineered
thermoplastic offering an unique combination of
chemical, mechanical and thermal properties. PEEK is
considered a premium seat material with its excellent
water/chemical resistance and due to the fact it is
unaffected by continuous exposure to hot water/steam.
PEEK is non-porous, high strength for high pressure
(400 bar) applications and is suitable for high corrosion
environments. PEEK typically adds to the torque
requirement of the valve given the rigidity of the
material. It can be used up to 250°C.
99
12

2.3 Solution
When considering the gasket material the ones that can give an upgrade from the TFMs
hardness rating is PCTFE and PEEK. Although PEEK is a more cost heavy gasket, and
in addition it increases the required torque and thus the size of the actuator, it is chosen
for its hardness. PEEK has the highest hardness rating, and should therefore handle the
wear from the mud.
The increase in hardness from the gasket change gives an increase in the necessary preload
from the springs on the gasket. This preload is there to make the contact pressure between
the gasket and the ball large enough to prevent leakage. To make the changes on the valve
as small as possible, the same springs are used but the start length will be changed, giving
them a larger force. This is further explained and calculated in section 4.1. Because of the
several assumptions made in the selection of the gasket, the valve should be tested in a
testing facility according to the ISO 10423 standard.
In addition to the material selection, valves of Norway requested a more standardization
of the design for the gaskets. Today’s version of the valve is featured with two gaskets in
each seat. This is shown in ﬁgure 2.2a where the gasket is shown in blue. After research in
gasket manufacturers solutions, it is found that the most widely used solution is seats with
a single gasket. On that basis, the new valve will be manufactured with one gasket instead
of two. This is shown in ﬁgure 2.2b.
(a) The old double gasket solution (b) The new single gasket solution
Figure 2.2: Illustration of old and new gasket solution
It is preferable to chose a standard measurement of the gasket such that it is easily
manufactured or ordered quickly. In collaboration with the external supervisor, the
dimension for the new gasket will be 126 mm for the outer diameter, and 116.7 mm for the
inner diameter. 1
1
The values were found after consulting Vladimir Mitin, the engineer at Valves of Norway. The values is
based on long experience and testing.
13

3 Bearings
In contacting surfaces with high friction, bearings are implemented to avoid failure by
reducing friction. The relevant bearing type for this purpose is roller and plain bearings.
Roller bearings transfer forces with rollers in the shape of balls or needles in different forms,
placed between two metal cylinders. Plain bearings, also called bushings or sleeve bearings,
transfers forces by sliding. The different bearing types are illustrated in figure 3.1a and 3.1b.
Roller bearings are especially suitable for tasks that requires high precision combined with
high speeds and high dynamic loads. Roller bearings have an exceptionally low coefficient
of friction. If correct lubrication is regularly preformed they may have a longer life span
than the plain bearing and are also available with lower clearance. The downsides with
roller bearings are the high cost, low vibration damping and that most roller bearings must
have lubrication at all times. That said, there are some maintenance-free roller bearings
on the market, but they do not inherit the precision, speed and load handling capabilities.
However they are often better in corrosive environments.
Plain bearings can further be divided into three types, one for each type of motion that
the plain bearing could be subjected to. These are linear, thrust and journal. In linear
bearing the bearing slides in the length direction of the shaft it is mounted on. Thrust
bearings are placed between to mating surfaces to reduce friction when the surfaces rotates
in different directions or speeds. Journal bearing consist of a shaft that rotates inside the
bearing placed in a static housing. Plain bearings can be self-lubricating and are vibration
dampening. They can handle high static load, are cost and space efficient and have low
weight in comparison to roller bearings. Most plain bearings also have high resistant to
dirt pollution. The downside for plain bearings are the higher coefficient of friction and the
higher clearance which in turns reduces the capable speed and the achievable precision.
(a) Plain Bearing (b) Roller Bearing
Figure 3.1: The two main types of bearings
Because valve maintenance is both expensive and difficult to execute in subsea environments,
the bearings must be maintenance free and have a long life span. The fluid running through
the valve is mud and this increases the friction and wear of the bearing. It must also
handle low speeds and large forces. The type that meets most of the requirements is the
self-lubricating plain bearings.
14

3.1 Material
When choosing the materials for the bearings there are some factors that must be taken in
to account. These are load capacity, friction coefficient, velocity, temperature and required
life span. The wide range in material choices ranges from white metals, alloys to plastics.
In self-lubricating plain bearings the material choices are listed in table 3.1 [7, 8] :
Table 3.1: Material for bearings
Material Information Max. stress [N/mm2
]
Unfilled
thermoplastics
It has the lowest cost, is superior in corrosive
environments and has a high tolerance to
soft material in mating partner.
10
Filled/
reinforced
thermoplastics
Injection molded resin with additive fiber as
base material. Second lowest cost,
compatible with abrasives and adequate
toleration of soft counter-face.
Filled: 12
+ reinforced: 35
+ bounded to metal back:
140
Filled/
reinforced
PTFE
PTFE is one of the most versatile plastic
materials. It is used as coating, extrusion
and molding. The common reinforcements is
glass fiber, carbon, bronze and graphite.
Handles low temperature exceptionally.
Scores good inn handling corrosion, abrasives.
It also has low wear and low friction which in
turn means a longer life span.
Filled: 7
140
+ reinforced: 420
Filled/
reinforced
thermosetting
resins
Thermosetting is similar but stronger to
thermoplastics. It can handle higher loads,
but has some higher wear and cost.
Filled: 30
+ reinforced: 50
PTFE
impregnated
porous
metals
PTFE impregnates in the micro-pores in the
metals. Can handle higher speeds, high loads.
It also has low wear/long life and low cost.
350
Woven
PTFE/ glass
fibre
Can handle high loads. Has high stiffness
and low wear.
Woven
+ Reinforced
420
Carbons
-graphite
Handles high temperatures and corrosive
environments. Has low wear and is
Chemically inert
1.4 − 2
Metal
-graphite
mixtures
Good in high temperature and high speeds.
Scores high in dimensional stability and is
compatible with fluid lubricants.
3.4
Solid film
lubricants
Coating of fine particles of lubricating
pigments, binder and additives that cures in
to a solid film. Has low friction and scores
high in handling high and low temperatures,
high speeds and high loads. Good stiffness
and stability. Has high tolerance to soft
counter-faces.
Ceramics,
cermets,
hard metals
Tackles high temperatures, high loads fluid
lubrication and corrosion. It is also
dimensional stable.
15

3.2 Reaction forces
The most widely used method for calculating the materials ability to withstand the frictional
energy generated is the PV factor. This is the load capacity times the peripheral speed.
To high PV value will give unstable temperatures that leads to rapid failure. To get a
reasonable value, a safety factor of 2 is added.
The given work pressure p = 10 MPa = 10 N/mm2
pushes on the ball with a diameter
of Db = 176 mm. The stem and ball will be rotated at a maximum of c = 15 cycles/min
and have an oscillating angle of θ = 90° = 1/2 · π. The gaskets has an outer diameter of
Dg = 126 mm and an inner diameter of dg = 116.7 mm.
There is both a force directly on the ball, and a force on the seat from the working pressure.
This will in turn create reaction forces on the three bearings. With three unknown forces
and two usable equations, the system is unsolvable. Computer aided analysis is therefor
used to ﬁnd these forces. In this situation fusion 360 was used, see appendix M for the
report. The direct and indirect forces working on the bearings are calculated in equation 3.2
and 3.4. The total force on the ball from the work pressure is then calculated in equation 3.5.
The work pressure will be inside the valve house, with the stem and trunnion, stopping at
the trunnion seals and the stem seal/glide rings. This will cancel out most of the force
from the pressure on the seat. This means that the force from the work pressure on the
seat will only come from the area of the gasket. In addition, the force on the ball from the
work pressure will only aﬀect the area inside the seat gaskets.
Apg =
π · D2
g − d2
g
4
= 1773 mm2
(3.1)
Fpg = Apg · p = 17727 N (3.2)
Apb =
πd2
g
4
= 10696 mm2
(3.3)
Fpb = Apb · P = 106960 N (3.4)
Ftot = Fsp + Fp = 124689 N (3.5)
The results from the computer aided analysis shows that the reaction forces was for the
upper stem Fus = 3.6 kN in the negative direction, the lower stem Fls = 48 kN and for the
trunnion Ft = 73 kN.
16

With hand calculations the ratios can be compered. The net force and net moment adds
up to zero for static problems. L1 is the distance from the trunnion to the ball center,
where the force Fb acts. L4 is the distance from Fb to Fls and L5 is from Fls to Fus. This
is shown in figure 3.2. From this the equations below are found.
Figure 3.2: Free body diagram
Mt = Fb · L1 + (L1 + L4 + L5) · Fus − (L1 + L4) · Fls = 0 (3.6)
Fus =
Fls · (L1 + L4) − Fb · L1
L1 + L4 + L5
= −4 kN (3.7)
Fx = Ft − Fb + Fls = 0 (3.8)
Fus = Ft − Fb + Fls = −3.6 kN (3.9)
The hand calculations and the FEM result gave a difference of around 500 N in equation
3.7. This is a small difference and only support the conclusion that the finite element
analysis gave the correct values. These values will be used in further calculations.
17

3.3 PV factor
To find the PV factor, the rotational speed and working load need to be found. These are
dependent on the bearing dimensions which is commonly defined by the length-to diameter
ratio. The length-to diameter ratio is often set in the range from 0.5 to 1.5, and if the
diameter is known the ratio often set to 1.
In this case the dimensions from the previous version of the valve will be used for an
estimation. Values for each bearing is shown in table 3.2. The peripheral speed is calculated
with the formulas in appendix D. Since the valve operates with an oscillatory motion,
equation numbered two in the appendix is used.
P =
F
L · d
(3.10) V =
dcθ
103
(3.11)
Table 3.2: Data for calculating PV factor
d [mm] L [mm] F [kN] P [N/mm2
] V [m/s] PV·2
Upper stem 46 7.8 3.6 10 0.009 0.18
Lower Stem 49 30 48 33 0.0096 0.63
Trunnion 44 19 73 87 0.0086 1.50
Appendix B shows that the materials suitable for the lower stem- and trunnion bearings is
the PTFE with filler, bonded to steel backing with a PV value up to 1.75, or the woven
PTFE reinforced and bonded to metal backing with a PV value up to 1.60. Both can
handle the required loading. Other materials could be chosen for the upper stem bearing,
but with forces from the actuator this would also require the same material as the others.
18

3.4 Selection from supplier
Valves of Norway are now buying bearings from the company TENMAT. They want to
continue this. For use in valves, TENMAT recommends their FEROGLIDE material for
oil and gas applications that demands high strength and low friction. It is a woven PTFE
fabric with strengthening ﬁbers applied to a metal backing. The maximum loading in static
and dynamic loading situations are listed in table 3.3 and 3.4, these includes comments
about when these are valid.
Table 3.3: FEROGLIDE static load capacity
Maximum
static load [N/mm2
]
Backing metal
210 Mild steel
240 Bronze
420 Inconel
Table 3.4: FEROGLIDE dynamic load capacity
Maximum
dynamic load [N/mm2
]
Comments
14 − 28 Long life
140 Recommended Maximum Load
176 High Strength backing metals
Static is used where there is little to no movement. Dynamic loading is when there is
oscillating motion or linear movement below 10 m/min. In this case it is an oscillating
motion and the table for dynamic loading is used. The recommended maximum load
is chosen to give the bearings long life span, but still be capable of handling the forces
necessary.
19

(a) Wear (b) Friction
Figure 3.3: FEROGLIDE wear and friction graphs
Figure 3.3a and 3.3b shows two lines on both graphs. One in red and one in blue. The
FEROGLIDE report does not describe the diﬀerence of the two. For that reasons it is
assumed that the two lines indicate the minimum and maximum values. To be on the safe
side, the values for the blue lines are therefor used in further calculations.
Figure 3.3a shows the wear on a journal bearing with a load of 140 N/mm2
. The bearing
was ﬁxed and the shaft was oscillating with an amplitude of ±45°. The frequency was 10
cycles/min [9]. This is close to the values for this valve and is therefor a good representation
of what to expect when used. For the usage the valve is designed for the wear will be low
and hold for the life span of the valve.
The dimension choices for the FEROGLIDE journal coiled bearings are found in appendix
C and the dimensions closest to the previous version of the valve are chosen and listed in
table 3.5 with comments about the necessary changes of the valve. All dimensions are in
millimeters.
Table 3.5: FEROGLIDE bearing dimensions
d/D L Comment
Trunnion 40/44 20 The ball hole length must be machined 1 mm in the height (
The new length is 1 mm longer then the old), and 1 mm must
be added to ball hole diameter. In addition the trunnion
diameter must be machined down 4 mm beneath the bearing.
Lower stem 45/50 30 The stem diameter beneath the bearing must be machined
down 4 mm.
Upper stem 45/50 7.8 The stem diameter must be machined 2.52 mm beneath the
bearing, and the length must be cut or specially ordered.
20

4 Calculations
The changes made to the gasket and bearings will affect several components of the valve
which will have to be recalculated for the new version. This includes the spring force, friction
torques and the stress on the stem, trunnion and bearings, calculated in that order. For all
of them the safety factor is 1.5 from the standard, see section 1.7. The effect on the ball and
seat are small and the calculations of stress on these components are neglected in this thesis.
Some assumptions are made, this includes what height the mount is on the stem. This is
assumed to be 40 mm from the top, which is in the middle of the square section. This
height is adjustable and if the mount is placed higher the result in this thesis should be
viewed as an underestimate, and maximum allowed force should be reduced some.
Assuming that the highest stress occurs in either the fully opened or closed position, the
actuator force is calculated as a perpendicular force to the flat surface on the stem. The
stress on the stem and trunnion will be a combination of stresses from shear, torsion and
moments. Calculation of the stress on the bearings will be simplified to only radial force
and the area will be the projected area, this was agreed upon with the external supervisor.
The actuator can push on all four sides of the stem. The four directions is the positive
and negative x- and y-directions, when the pressure is working in the positive x-direction
also shown in figure 4.1. The worst case scenario is the one that is further checked. To
find the worst case scenario, several FEM analysis are done with the force from different
directions while changing the actuator forces. Through a large number of analysis, it was
found that when the actuator force worked in the opposite direction of the work pressure
it gave the highest reaction forces on the bearings. This is therefore the direction that will
be calculated, tested and designed for.
Figure 4.1: Illustration of the directions when the ball is in open position
21

Since both the stress on the bearings, trunnion and stem depend on the magnitude of the
actuator force, the maximum force from the actuator must be found to have as a limit.
This is done to have a value for the minimal arm length when designing the mount.
A parametric Soulver file will be created where all equations discussed in the next subsections
are implemented and the reaction forces and moments from the bearings are imported from
the finite element analysis. Soulver is a calculator software. It can remember variables,
which can later be used in formulas. This means that values can easily be changed and the
files can be saved unlike in a physical calculator.
A number of FEM analysis are then run with different actuator forces, checking each
one in soulver to find the force which will bring one of the components to its allowed
stress limit. The FEM setup is shown in figure 4.2. The complexity and amount of these
calculations is at a scale that makes it unnecessary to have in the body of this thesis and is
therefor placed in appendix K . Only the methods used and the results are shown in the body.
Figure 4.2: FEM setup
22

Figure 4.3: Notations used in calculations
The ﬁgure 4.3 shows the cross sections that the stress is calculated for, excluding the thrust
bearing. Also their notations is written in parenthesis, which is their subscript in the
calculations. These sections are chosen because of their change in geometry and some by
the local moment peaks. Figure 4.4 shows further how the stem is divided up.
Figure 4.4: Illustration of the components used on the stem
23

4.1 Friction torques from components
There are friction torques from several components. On the stem there is the upper and
lower stem bearing shown in red and the thrust bearing shown in green in figure 4.4.
Additionally there is the gasket and the trunnion bearing. These friction torques will in
turn create the torsion.
Friction torque from bearings
The friction coefficients for FEROGLIDE are found in the graph from TENMAT in figure
3.3b on page 20. For pressures above 40 MPa the coefficients will be rounded up to 0.03
because of the uncertainty with the friction graphs. This value is used for the upper stem
bearing. The lower stem bearing and the trunnion bearing will run partly in mud and this
increases the friction coefficient greatly and is therefore estimated to 0.1.2
If the seals were
to be placed below the lower bearing the friction coefficient increase would be avoided for
this bearing. On the other hand the distance from the force on the ball would increase the
force and stress on the trunnion bearing, which it would not withstand with the material
it is assigned.
The friction torque from the bearings are calculated with an equation from SKF [10], where
the arm is the radius of the bearing. The new friction coefficients and dimensions for each
bearing is listed in table 4.1. The frictional torque is dependent on the magnitude of the
actuator force and is therefore calculated for each FEA, where equation 4.1 is used.
M =
F
2
· µ · D (4.1)
Table 4.1: Friction torque on the bearings
µ L [mm] D [mm]
Upper stem 0.03 7.8 50
Lower stem 0.1 30 50
Trunnion 0.1 20 44
2
From meeting with Vladimir Mitin where the friction coefficients were discussed. 23/02/2016
24

Friction torque form glide rings
The friction torque from the glide rings depend on the projected area of the glide ring
on the stem hole, and the arm which is the radius from the stem center to the glide ring
edge. The glide rings are made of the material TFM [3], with a friction coefficient of 0.06.
This is increased to 0.1 to take the effect of the mud in to account. The friction torque
is calculated in equation 4.2, where the height of the glide ring is w = 3.9 mm and the
diameter of the stem hole is dsh = 50 mm.
Mgr = µ · p · w · dsh ·
dsh
2
= 4875 Nmm (4.2)
Friction torque from the thrust bearing
The thrust bearing shown in green in figure 4.4 takes up all the vertical force from the work
pressure. The pressure works on the area of the stem hole. The valve is designed for a depth
of 1000 m, which will also give it a pressure of L · ρ · g = 1000 m · 1000 kg/m3
· 9.81 m/s2
≈
10 MPa on the top of the stem directed downwards. This will create a reaction force
of equal size and they cancel each other out. To get a conservative result it is therefor
calculated for a vacuum. The thrust bearing is made from PETP and the friction coefficient
is µ = 0.18 [11]. The outer diameter and inner diameter is Dtb = 72.5 mm and dtb = 50.5
mm as in the previous version of the valve.
Mtb = µ · p · π · d2
sh ·
1
4
·
1
4
· (Dtb + dtb) = 60377.48 Nmm (4.3)
Friction torque from the gasket
The gasket gives a force from both the 40 springs and the resultant force from the work
pressure on the seat. The force from the work pressure was calculated in equation 3.2
on page 16. These will generate the friction torques from equation 4.6 and 4.7. These
equations are used by the engineers at Valves of Norway, found in [12]. The equation has a
correction factor to compensate for the inconsistent arm or the distance from the gasket
to the vertical center line of the ball. This correction factor is dependent on the slope or
incline of the seat, which in this case is β = 50°.
25

The spring force is in accordance with Vladimir Mitin determined by the suggested preload
of the gasket. The equation for preload is based on experience and testing. It is found to
be around 5% of the maximum allowed loading which is calculated as 60% of the tensile
strength σt = 110 N/mm2
[5]. 3
The same springs will be used in the valve as in the previous version. To get the correct
spring force the start length will be changed. The spring coefficient is set to k = 55 N/mm,
found in the data sheet for the springs which is located on page xxvi. The end length is
found in the 3D model to be l1 = 13.9 mm.
Fsg = 0.05 · 0.6 · σtensile ·
π
4
· D2
g − d2
g = 5850.01 N (4.4)
l0 =
F
nsg · k
+ l1 = 16.56 mm (4.5)
The gasket has an outer diameter of Dg = 126 mm and an inner diameter of dg = 116.7
mm. The friction coefficient is approximately 0.4, a high friction coefficient in comparison
to the bearings and seals. This along with the large force from the work pressure gives the
gasket the highest friction torque of all the components.
The friction torque from the gaskets will depend much on the position of the valve. If it
is closed there will only be pressure on one side of the ball. In this way only one of the
gaskets will generate friction torque. When the valve is opened the work pressure will
work on both sides and the friction doubles. The result from this on the stress analysis is
discussed further in the next section. The force from the springs will work on both sides
for both positions.
Mpg = Fpg · µ ·
1
2
· Db ·
(1 + sin(50°)) · 0.5
sin(50°) + 0.05 · cos(50°)
= 690325.60 Nmm (4.6)
Msg = 2 · Fsg · µ ·
1
2
· Db ·
(1 + sin(50°)) · 0.5
sin(50°) + 0.05 · cos(50°)
= 227807.45 Nmm (4.7)
3
From email correspondence with Vladimir Mitin where the formula is based on testing and long
experience in the valve business 01/03/2016, the e-mail is in appendix J.
26

4.2 Stress Calculations
The stress will be different for each section of the trunnion and stem. It will be a combination
of torsion, shear and moment stress. The stress will also be different for the two positions;
open and closed. When the valve is open, there will be a pressure working on the gasket
on both sides of the ball. In this position there will be no reaction forces from the bearings,
meaning no friction torques from the bearings, no shear force and no moment.
If the valve is closed, there will only be friction torque from one of the gasket, but the
friction torque, shear force and the moment from the bearings will have an effect here.
Because of the high friction torque of the gasket, both situations have to be checked to see
which gives the highest total stress.
The different sections are the trunnions circular section, the stems hexagonal section at
the bottom of the stem. Further the stems lower circular section in addition to its circular
sections below the two bearings and the seal/glide rings. These sections are chosen because
of the change in geometry and/or the size of the moment, shear and torsion.
The values shown are the final result with the maximum actuator force. The moment
of inertia and the first moment of area for the hexagon section is dependent on where
the neutral line will go. In this case it will go through the corners when closed. For the
hexagon the dimensions used are the width of w = 36 mm and side lengths of a = 21 mm.
For the circular section the diameters are listed in table 4.2.
Table 4.2: Diameters of circular sections
Trunnion Low circular Lower stem Seal Upper stem
Diameter [mm] 44 45 45 39 45
27

Torsion
The torsion will increase for each addition of frictional torque, starting with the trunnion
bearing and adding up to the total at the thrust bearing. The distribution is shown in the
torsional diagram in figure 4.5. The highest jump is from the gasket, which have a high
frictional torque because of its high friction coefficient, high force, and a long distance to
the center of the valve. Since this is above the trunnion, the trunnion itself can be made of
a material with a lower yield strength than the rest. At the top of the stem there will be
a total frictional torque that the actuator has to counter with a moment in the opposite
direction.
The friction torques depend heavily on the position of the valve. If it is open there will
be a force of zero on the trunnion bearing and the lower- and upper stem bearing, but
the pressure affect both gaskets. This significantly increases the torsion. In this position
the torsion will be constant at any actuator force. When the valve is closed on the other
hand, the friction torque from the bearings will vary with the actuator force as described
previously.
Figure 4.5: Torsion diagram
28

To calculate the stress from the torsion on circular sections which is a uniform cross-section,
equation 4.9 is used. Where T is the torsion, r is the radius and J is the torsion constant
which is equal to the polar moment of inertia, which is used to estimate the objects ability
to resist torsion.
J =
1
2
· π · r4
(4.8)
τt =
T · r
J
=
16 · T
π · d3
(4.9)
The equation above only applies for uniform cross-sections. Non uniform cross-sections such
as the hexagon behaves non-symmetrically when torque is exerted on them. In addition
to this the stress distribution is often non-linear. This means for hexagonal sections that
the equation is estimated to equation 4.11 [13], where the width (w) is measured from the
ﬂat sides of the hexagon. The results are listed in table 4.3 for both open and the closed
position.
J = 0.0649 · w4
(4.10)
τt = 5.297 ·
T
w3
(4.11)
Table 4.3: Torsional stress with closed and open valve
[Nm] Trunnion Hexagon Low
circular
Lower
stem
Seal Upper
stem
Closed 164 1310 1310 1410 1415 1436
Open 0 1836 1836 1836 1841 1841
29

Shear
Shear stress is caused by forces acting in a direction parallel to a surface or in other words,
a force acting along a plane that passes through a body. The forces in this case are the
ones from the work pressure, the actuator force and from the reaction forces in the bearings.
In ﬁgure 4.6 the shear diagram is illustrated. This is not to scale and only have the purpose
of illustrating how the shear forces is over the trunnion and stem. VA is from the trunnion
force. The force from the work pressure will then bring the shear force below zero to VB.
Further it will change direction for the lower and upper stem bearing. Finally ending at
zero. The calculation method for the shear forces are shown in equation 4.12.
VA = Ft (4.12)
VB = VA − (Fps + Fpb) (4.13)
VC = VB + Fls (4.14)
VD = VC − Fus (4.15)
Figure 4.6: Shear force diagram
30

In many cases, especially for long sections, the shear stress is neglected because of its
small size in comparison to the stress from moment. For small and thick beams the shear
stress has a much larger importance. The shear force that comes from bending is called
transverse shear and is at its highest at the neutral axis and at zero with the walls. The
geometric factors of the transverse shear is the width of the object, the moment of inertia
I, and the ﬁrst moment of area Q. First moment of area is the area of either the above or
below cross-section of the center line, times the distance from the center line to this areas
centroid. Equation 4.18 shows the method of calculating shear stress for circular sections.
Q = ai · yi =
π · r2
2
·
4r
3 · π
=
1
12
· d3
(4.16)
I =
π
64
· d4
(4.17)
τs =
V · Q
I · d
=
16
3
·
V
π · d2
(4.18)
To solve for Q for the hexagonal section, the hexagon is divided in two. A top and bottom
half. These are equal and which of them the area is found for is the same. To ﬁnd the
area the half is further divided into a square and two right angled triangles, one on each
side. The triangles and square will have a height of w/2 , and the length of the hypotenuse
of the triangles and the width of the square is a = 20.67 mm, which is the length of the
hexagon sides. From this the stress is found with the equations below.
Q = ai · yi =
area of square
w
2
· a ·
distance
w
4
+
area of triangles
2
1
2
·
w
2
· a2 −
w
2
2
·
distance
1
6
· w (4.19)
Q =
1
8
· a · w2
+
1
12
· w2
· a2 −
w
2
2
(4.20)
I =
5 ·
√
3 · a4
h
16
(4.21)
τs =
V · Q
I · w
(4.22)
The trunnion works more like a pin than a beam and is therefor calculated for average
shear stress. τs = V/A. And for the section with the lower stem bearing and the upper
stem bearing the shear force will be zero. The results are shown in table 4.4.
Table 4.4: Shear stress
Trunnion Hexagon Low circular Seal
τs [N/mm2
] 65 21 36 2
31

Moment
The moment is defined as the force multiplied with the lever arm which is the perpendicular
distance between the the center of moments and the force. In this case the shear force
above is used and the lever arms are listed in table 4.5.
Table 4.5: Lengths
Notation Length [mm] Description
L1 75 From trunnion to ball center
L2 56 From ball center to start of hexagon
L3 66 From ball center to start of lower circular
L4 100 From ball to mid lower stem bearing
L5 30 From mid lower stem bearing to mid second seal
L6 39 From lower to upper stem bearing
L7 63 From upper stem bearing to middle of square stem
The finite element analysis calculates that the bearings not only take up forces but also
some amount of moment. These moments are added as jumps in the moment diagram.
The moment diagram in figure 4.7 is not to scale and only there for illustration purposes.
Figure 4.7: Moment diagram
32

The moment of inertia (I) and the distance to the outer most fiber (y) is calculated in
different ways for the circular and hexagon cross-sections. For the circular sections the
outer most fiber is a distance of the radius from the neutral axis, and equations for moment
of inertia is shown below.
y =
d
2
(4.23)
Ix =
1
64
· d4
(4.24)
For the hexagon section, the outer most fiber is a distance of the width divided by two.
y =
w
2
(4.25)
I =
5 ·
√
3 · a4
h
16
(4.26)
With the moment, moment of inertia and the distance from the neutral axis to the outer
most fiber, the stress from moment can be found with equation 4.27. The results are listed
in table 4.6.
σ =
M
I
· y (4.27)
Table 4.6: Moment stress
Trunnion Hexagon Low
circular
lower
stem
seal upper
stem
τ [N/mm2
] 96 134 303 140 226 205
33

Total stress
The total stress is a combination of the torsion, shear and moment. To estimate the
equivalent stress, the von mises criteria is used. It is denoted as σv and this must be lower
then the total allowed stress for the material. The von mises stress result is listed in table 4.7.
The material used for the stem is a nickel based super alloy called UNS NO7718. It has a
yield strength of σy = 827 N/mm2
[14]. With a safety factor of 1.5 the allowed stress is
σa = σy/1.5 = 551.33 N/mm2
.
The trunnion is made of a weaker material since it experiences a smaller amount of stress.
It is made of the stainless steel A479 UNS S31600 with a yield strength of σy = 204 N/mm2
[15], giving it an allowed stress of σa = 136 N/mm2
.
σv = σ2 + 3 (τs + τt )2
(4.28)
Table 4.7: Von mises stress
circular
lower
stem
seal upper
stem
Closed 135 327 357 195 309 248
Open 0 367 178 178 278 178
The results are shown as utilization of the material and listed in table 4.8. If this is close
to hundred percent, the utilization is good and no changes need to be done. If it is above
hundred percent, the material can not be used or the actuator force should be reduced. If
it is well below, possible changes will be suggested but not implemented in this thesis.
Utilization =
σv
σa
· 100 % (4.29)
Table 4.8: Utilization with closed and open valve
circular
lower
stem
seal upper
stem
Closed 99 59 65 35 56 45
Open 0 67 32 32 50 32
34

4.3 Finite element analysis
After many trials and errors, tweaking the constraints and loads the ﬁnal result was found.
All bearings were constrained with pin constraints. For the lower and upper bearing only
the radial direction was locked. To take up the small vertical force the trunnion bearing was
locked in both the radial and the axial direction. This is in reality taken up by the thrust
bearing, but if the thrust bearings was included in the analysis they would in addition take
up a large amount of the moment. This may be the case if the deﬂection is large enough
but not likely, so these were excluded from the analysis.
The highest actuator force was found to be around 30 kN were the failing section was the
trunnion in addition to the bearing on the trunnion. By adding the torsion to the analysis,
the stress for each part was found and listed in table 4.9. There were singularities, but
these are ignored. The analysis also converged to a convergence rate of 2 − 4% with the
von mises stress results. The utilizations are listed in table 4.10.
A side note is that the peak stress for the hexagon and the lower circular part were at the
top of each section instead of at the bottom as in the hand calculations.
Table 4.9: Stress from FEM
circular
lower
stem
seal upper
stem
Closed 120 360 350 200 325 220
Open 20 400 220 150 400 230
Table 4.10: Utilization from FEM
circular
lower
stem
seal upper
stem
Closed 90 65 65 35 60 40
Open 15 75 40 30 70 40
35

4.4 Stress on bearings
The calculation for the bearing stress is simpliﬁed to be plain compression stress on the
projected area of the bearing. The projected area is the width times the diameter. The
safety factor is set to 1.5 from the ISO 10423 standard. With a maximum dynamic load
of σmax = 140 N/mm2
, the total allowed stress is σa = σmax/γ = 93.33 N/mm2
for the
FEROGLIDE bearings. The force acting on them is the reaction forces, these are found
with ﬁnite element analysis. The stress and utility results are listed in table 4.11 and 4.12
A = b · h (4.30)
σ =
F
A
(4.31)
Utilization =
σ
σallowed
· 100% (4.32)
Table 4.11: Stress on bearings
Trunnion lower stem upper stem
Stress [N/mm2
] 93 30 81
Table 4.12: Utilization on bearings
Trunnion lower stem upper stem
Utilization [%] 99 32 87
36

4.5 Conclusion stress calculations
There was a small gap between the FEM analysis and the hand calculations. The FEM
analysis showed a higher stress on especially the seal at the open position, which gave it
an utility of a little over 70%. The difference is of a small scale and most likely from the
complexity of the calculation. Therefore the results mostly support each other and the
conclusion will be that they are correct. In figure 4.8a the von mises stress is shown, where
the least affected areas are shown in blue and the more affected areas are shown in green.
The yellow and red areas from the scale at the top of the figure is not visible in the model
since they indicate the points with singularities and are to small to notice.
As the calculations shows, most sections only have a utilization of around 40 − 60%. This
means that the trunnion, trunnion bearing and the upper stem bearing could be scaled up
to accommodate more stress, or the other sections could be scaled down. The last solution
will reduce the material usage and size, and in turn the material cost on the valve. The
sections that should be reduced is shown in figure 4.8b.
The results gave an actuator force of 30 kN which gives a minimum arm length of
L = Ttot/Fact ≈ 50 mm. Since there is no safety factor for the friction torques, other than
the added factor for the friction coefficient on the bearings running in mud, the friction
torques may be a conservative value. If a safety factor of 1.5 is added the minimum arm
length will be 75 mm. Combined with a stem width of 40 mm, this is still a small arm
length and will not create any problem when designing the mount.
(a) Von Mises when closed (b) Possible modifications
Figure 4.8: Result stress analysis
37

5 Valve Actuator
For this valve the actuator shall meet the requirements of the ISO 10423, ch.10.16.3 [16].
An actuator is a motor that controls and moves mechanisms or systems by taking a power
source in the form of hydraulic fluid, pneumatic pressure or electric current to perform
work. This can be in the form of linear, rotary or oscillatory motion. They have a diverse
area of usage in many fields, including engineering.
Actuators are used when human operations are not desirable/possible, and when operations
are done either quickly or frequently. An advantage is also that the actuator and in turn
the valve can be controlled remotely. In this situation the actuator has a linear motion
and the motion is transformed to a rotary motion by the mount.
5.1 Types of actuators
There are three types of actuators that are mainly used, each type giving different mount
solutions. The actuator type that is most likely to be used with this valve is found, and
the mount is then designed with this in mind. The information about each type is found
in [17, 18].
Electric actuator
Electric actuator converts electrical energy into torque. An electric motor mechanically
connected turns a lead screw. It is very simple to connect and wire, also very easy to
use. Electrical actuators offers high precision-control positioning, they provide complete
control of motion and can include encoders to control velocity, position, torque, and
applied force. The initial unit cost of an electrical actuator is higher than pneumatic and
hydraulic actuators, but it has very low operating cost. The reliability is great because
of its repeatable, reproducible performance during the entire product life and very little
maintenance is required. It does not have as long life expectancy as the hydraulic actuator,
and it is not suited for all environments.
Pneumatic actuator
Pneumatic actuator consist of a piston inside a hollow cylinder. Pressurized air from an
external compressor or manual pump moves the piston inside the cylinder. Pneumatic
actuators generate precise linear motion. The cost of pneumatic actuators is low compared to
other actuators and pneumatic actuators are also lightweight, require minimal maintenance
and have durable components. This makes pneumatic actuators a cost-effective method
to produce linear motion. The pneumatic actuator is also safe to use in hazardous and
flammable areas. The main disadvantage with this type of actuator is its medium. Air is a
compressible medium which makes this a less efficient actuator.
38

Hydraulic actuator
Hydraulic actuators operates similarly to pneumatic actuators, but uses an in-compressible
fluid from a pump rather than pressurized air to move the cylinder. It requires plumbing,
filtering, pumps etc. and requires electronic/fluid interfaces and valve designs to be
operated. The speed is difficult to control accurately, but it gives virtually unlimited force
and is the most powerful choice. The actuator can produce forces 25 times greater than
pneumatic actuator of equal size.
A hydraulic actuator can hold the force and torque constant without the pump supplying
more fluid or pressure, this due to the in-compressible fluid. The pumps and motors
can be placed a considerable distance away with minimal loss of power. Hydraulics is
contamination sensitive and seals are prone to leak, but it has a longer life expectancy
than electrical when maintained regularly. Components often cost less than the electrical,
but installation and maintenance are high.
Figure 5.1 illustrates the principle of hydraulic actuators, where the fluid is pumped up
from a tank and delivered to one of the ends of the actuator pushing on the actuator piston
and shaft. The return is either done with springs or by changing the flow direction.
Figure 5.1: Illustration of how the fluid actuator work.
Choosing the right type of actuator
In a subsea environment there is most likely already installed a hydraulic power unit and
therefore this cost is not taken in to consideration. The electric actuator requires little
maintenance, but have shorter life expectancy. The pneumatic actuator are less efficient
and not suitable in subsea environment because of its compressible medium. In this task
precision is not important and for this and the reasons above it is assumed that a hydraulic
actuator is used. The mount will therefore be designed for this.
39

5.2 Commonly used mechanisms
The commonly used mounts are scotch yoke, slider-crank, lever and the rack and pinion.
The scotch yoke is a mechanism that converts linear motion to rotational motion by
having a slot in a sliding yoke. A pin on a off-center point on the rotational part is then
placed in the slot. When the yoke slides forward and backward the pin inside slides up
and down, resulting in a rotational motion, see figure 5.2. This is an efficient production of
rotational motion as it spends more time at the high points of the rotation then a piston
and has few parts. The setup is commonly used in control valve actuators in high-pressure
oil and gas pipeline [19].
Figure 5.2: Illustration of the scotch yoke.
The slider-crank mechanism transfers the linear motion from the actuator to a crank,
also called a lever, see figure 5.3. The mechanism contains two joints to maintain the
necessary freedom of motion. The lever is there to create a larger moment on the rotational
part. If the lever is longer, the moment gets larger or the force can be reduced. The
downside of scaling up is that the necessary travel length of the slider would increase. This
mechanism contains many parts and rotating joints that all need bearings.
Figure 5.3: Illustration of the sliding crank.
40

Lever mechanism is much like the slider-crank mechanism, but instead of two joints, it
uses a slot to maintain the necessary degrees of freedom, see figure 5.4. This is the opposite
method that the scotch yoke mechanism uses, where the slot is in the yoke. The advantage
is that it has few parts and uses a small amount of space. There are three types of lever:
First order, second order and third order. The first order levers works by placing the pivot
point in between the input force and output force. Since the forces are on different sides of
the pivot point, they will act in the opposite directions. In the second order lever, both
forces are placed on the same side of the pivot point. Were the input force is furthest away.
In this type of lever both forces will act in the same direction. In the third-order lever the
input and output forces are still on the same side, but now the output force is on the outer
end. This will also give the forces the same direction. The advantage here is the larger
travel distance of the output, but in turn the input force must be larger to counter the
moment.
Figure 5.4: Illustration of the lever solution.
The rack and pinion mechanism use gears to convert the motion. The pinion is the gear
on the rotating part, this is being rotated by a linear ”gear” bar that is called the rack.
The rack travels linearly on the edge of the pinion with the the teeth engaged. This pushes
the pinion in a rotational motion, see figure 5.5. The solutions contains very few parts,
and is therefor a simple and robust mechanism. It does not take up a large amount of
space but the gears may be sensitive to debris.
Figure 5.5: Illustration of the rack and pinion mechanism.
41

5.3 Design
When designing a mount a principle base is chosen from the most commonly used
mechanisms and then innovated on for this specific situation. The mechanism choice
is predominantly based on the simplicity, in addition to the material and space usage.
The slider-crank mechanism is not preferred here at all due to its large motion range, and
that it contains several parts, including several bearings which only adds to the costs and
complexity in this project. The rack and pinion does not take up a lot of space, but due to
the gears it could be more sensitive to debris, and therefore should be maintained regularly
to check for this, which is not possible in this case. Both the slider-crank and the rack
and pinion mechanisms have arms that must be an extension to the actuator arm so the
actuator may need to be attached further from the stem. For these reasons the slider-crank
and rack and pinion are discarded as a base mechanism.
The scotch yoke is small in size and easy to produce. The actuator can be placed closer to
the stem in both directions, making it a compact solution which can be mounted at the
center line of the valve if needed. This reduces the travel length of the actuator shaft. The
lever also has a small space usage and are easy to produce with little complexity and a
small material usage. Based on these advantages, further development will be based on
the scotch yoke and lever.
Pointers to take into consideration when designing the mount is that the stem end is a square
with a 39 × 39 mm2
area and the mount mechanism is being mounted to this. The height
on the stem should be adjustable to accommodate for different actuator models and sizes.
The actuator should also be able to be mounted to the valve, meaning that the distance
from the stem center to the actuator shaft should not be longer then around 200 mm, a
length estimated from studying the 3D model and the 2D assembly shown in appendix E.
The necessary turn angle is 90°, and the solution does not need to accommodate a larger one.
Furthermore the material chosen is a stainless steel A479 UNS S31600 with a maximum
allowed stress of 206 N/mm2
, same as the trunnion. The slot in the mount should also be
rounded in such a way that the resultant force from the actuator always is near tangential
to the stem. For this reason the actuator must deliver a higher force if it is at an angle to
the tangential of the stem to make the resultant force large enough to give the necessary
moment. This is further discussed in the calculations.
Trough brain storming some concepts were drawn up and presented in figure 5.6 with
notes about their positive and negative features. Some contains features from both the
scotch yoke and the lever, and some is based solely on the lever mechanism with different
connection methods. The mechanism numbered one uses a pin connected at both ends
trough a slot in the mount but the connection method could be changed to all of the three
in the top of the figure. The mechanism numbered two uses a linear bearing and a pin in a
square slot, and the mechanisms numbered three uses a slot in the shaft of the actuator to
push a pin fixed to the mount.
42

㄀
㈀
㌀
⬀ 匀洀愀氀氀攀爀搀椀猀琀愀渀挀攀琀漀愀挀琀甀愀琀漀爀
ⴀ 䰀愀爀最攀爀洀愀琀攀爀椀愀氀甀猀愀最攀
ⴀ 䠀愀爀搀攀爀琀漀瀀爀漀搀甀挀攀
⬀ 䈀攀琀琀攀爀琀漀氀攀爀愀渀挀攀
⬀ 匀椀洀瀀氀攀琀漀瀀爀漀搀甀挀攀
ⴀ 䴀漀爀攀挀漀洀瀀氀攀砀洀攀挀栀愀渀椀猀洀
⬀ 匀椀洀瀀氀攀洀攀挀栀愀渀椀猀洀
⬀ 匀椀洀瀀氀攀琀漀瀀爀漀搀甀挀攀
⬀ 匀洀愀氀氀䴀愀琀攀爀椀愀氀甀猀愀最攀
ⴀ 䰀愀挀欀椀渀瀀爀攀挀椀猀椀漀渀
Figure 5.6: Sketches of concepts
43

The features that is most important is the simplicity of the mechanism and production in
addition to a small material usage. This reduces the cost and at the same time increases the
rigidity, which is beneficial as it is placed in a harsh environment with limited possibilities
of maintenance. The solution shown as number one in figure 5.6 is the solution chosen for
its simplicity. The main disadvantage with this solution is its lack of precision. This is not
a major concern since the clearance when opening and closing a valve should be sufficient
and will not suffer for slack in the mount slot.
A slide is cut in the back of the mount with a M6 stainless steel socket head cap screw
through the gap, clamping the mount to the stem at the desired height. The tap- and
clearance drill dimensions is found in appendix F. The total friction torques is at its largest
when the valve is opened, and is therefore the position calculated for.
Finite element analysis is used to tweak the dimensions by strengthening the weak sections,
and reducing the strong section. To accommodate the changing intersection angle of the
actuator and the tangent to the stem, the slot is not made straight. Instead it is made
with an arc. In the first version this is a quarter of a circle, but in the second the slot is
only a portion of that, such that the it gives the end a more correct angle to the stem.
The resultant force has a more parallel direction to the tangent of the stem with these
curvatures. The portion of the force that is not parallel, does not contribute with moment
to the stem. Both arcs are estimated by testing the motion with a 3D modeled actuator
and contact sets to simulate their behavior.
The fist version, shown in figure 5.7 was designed for simplicity and for its low material
and space usage. This makes the production and material costs as low as possible. Since
the moment is at its highest at the stem center and decreases towards the force, version
one is designed with a larger amount of material around the stem, giving this section a
larger moment of inertia. Since the shaft can not be placed to close to the stem center,
this versions has to start and end at an angle. In the FEM analysis, an angle of 45° to
the valves x-direction is used (length axis). The actuator therefore has to deliver a larger
force to give the stem the necessary moment from the resultant force. This means that the
shear force will have a larger importance in the stress analysis, and more importantly that
the actuator has to be larger and in turn more costly. For these reasons the first version
was discarded over the second.
Figure 5.7: Render of mount - version one
44

The second version shown in figure 5.8 was made to improve and innovate on the problems
of the first version. The largest moment is as previously stated when the valve is fully
opened, it decreases as the valve closes and is at its lowest when the valve is closed. This
means that the actuator force should come at a 90° angle at the start, which makes a equal
but opposite directed resultant force parallel to the tangent of the stem. To achieve this
without making the shaft operate closer to the stem, a bend is made to the mount such
that the end of the mount is not aligned to the stem. This gives the mount the ability to
start perpendicular and end parallel to the valves x-direction. The slot was also narrowed,
giving it a small increase in precision. The only downside this version has to the first is its
larger material and space usage, but this increase is of a small scale, and will therefore be
neglected.
Figure 5.8: Render of mount - final version
45

Stress on mount
The total torsion is in the closed position approximately 1.5 · 106
Nmm and in the open
position 1.9 · 106
Nmm. Since the torsion is at its highest in the open position, this value
is used in further calculations.
The highest moment will occur with the center line of the stem. The moment of inertia is
calculated in equation 5.1 through 5.6, where r = 19 mm is the width of the stem divided
by two. a = 20 mm is the mount thickness and b = 26 mm is the width of each part in the
section. This is shown in ﬁgure 5.9. The maximum allowed stress σa = 136 N/mm2
.
Figure 5.9: Illustration of a and b
Ix = Ii + Ai · d2
i (5.1)
I1 = I2 = I =
1
12
· a1 · b3
1 (5.2)
A1 = A2 = A = a · b (5.3)
|y1| = |y2| = y = r +
b
2
(5.4)
¯y = 0 (5.5)
Ix = 2 · I + 2 · A(¯y − y)2
=
5
12
· a · b3
+ 722 · a · b + 38 · a · b2
≈ 1 · 106
mm4
(5.6)
σ =
Ttot
Ix
· (r + b) ≈ 86 N/mm2
(5.7)
σ · 100%
σa
≈ 63% (5.8)
46

The mount is also tested in FEM analysis and the von mises stress result is shown in ﬁgure
5.10. The peak stress was found to be in the slot corner. There were also stress peaks in
the corners of the square hole,which is known as stress concentration, an increase in stress
due to unevenly distributed forces where there is a sudden change in geometric shape.
With the length from the actuators start position to the center of the stem being
approximately 130 mm the force necessary is F = Ttot/L = 12 kN. In the FEM analysis on
the other hand, the necessary force to generate the right moment was found to be 14.5
kN. This may have a correlation with the oﬀset from the center line, or from the slight
curvature of the slot. The force in the FEA is the minimum force the actuator has to deliver.
When ignoring stress singularities, the von mises stress is around 70 − 100 MPa, which
gives a utility of around 80%, see appendix N. The mount could be better optimized, but
with the singularities a more conservative approach is taken. The conclusion is that this
mount will handle the forces applied.
Figure 5.10: Von mises FEM result for mount
47

Bolt stress
The bolt that is chosen for the mount is an M6 stainless steel socket head cap screw made of
Metric Type 316 Stainless Steel—DIN 912-A4, see 2D drawings on page xxii. The material
has a yield strength of σy = 205 MPa, see appendix G. This bolt is mostly loaded with
tension, and the tensile area is based on the the average of the minor and pitch diameter.
[20] The pitch is set to P = 1 mm giving the bolt a tensile stress area shown in equation
5.11.
dp = d − (0.649519 · P) = 5.35 mm (5.9)
dr = d − (1.226869 · P) = 4.77 mm (5.10)
At =
π
16
· (dp + dr)2
= 20.1 mm2
(5.11)
In most cases a preload is used to give the bolt a clamping force. The friction created
between the threads, head and the main component makes sure that the bolt do not come
lose under vibration. High preload tension also increases the strength of a joint. The
preload is different for reusable and permanent bolts. In this case the bolt is assumed to
be permanently installed.
The preload is calculated with the proof strength σp. For this material the proof strength
is unknown and is therefor estimated. This is done with equation 5.12, and further used to
find the preload for the permanently installed bolt in equation 5.13.[20]
σp = 0.85 · σy = 174.25 MPa (5.12)
Fi = 0.9 · At · σp = 3200 N (5.13)
The torque required to tighten the bolt is dependent on the constant K which depends
on the bolt material and size. Using the table of friction coefficients for stainless steels
in appendix H [21]. Assuming that the bolt and mount has equivalent properties to a A2
material grade, and a high resilient of connection the friction coefficients is found to be
µ = 0.26 in the thread and µ = 0.35 under the head. Although these values are specifically
for hex head screws, they are assumed to be close to the values for socket head cap screw.
The corresponding K value is found in appendix I [22] to be K ≈ 0.383. From this the
torque needed is calculated in equation 5.14.
T = K · Fi · d ≈ 7400 Nmm (5.14)
48

This clamping force gives a friction force between the mount and the stem. This will be
dependent on the friction coefficient between the two parts. In this case it is steel against
steel, giving it an approximated friction factor of µ = 0.5. The force needed to move the
mount can then be found using equation 5.15. Both forces are illustrated in figure 5.11.
F = Fi · µ = 1600 N (5.15)
The equivalent mass is m = F/g ≈ 160 kg, were g is the gravitational acceleration. This
is the weight that can be placed on the mount without displacement. This should be
sufficient to prevent any problems from its own weight and any debris hitting the mount in
subsea environments.
The bolt will only be affected by tension and the only force acting on it is the preload.
This means that the stress is simply force over tensile area, shown in equation 5.16. As
described in section 1.7, the safety factor is 1.2 for bolts. From this the allowed stress and
utilization is calculated in equation 5.17 and 5.18.
σ =
Fi
Atsa
= 160 MPa (5.16)
σa =
σy
γ
= 171 MPa (5.17)
σ
σa
· 100% = 93% (5.18)
Figure 5.11: Preload and friction force
49

5.4 Alternative design
If a situation requires a smaller arm and/or a higher safety factor, the solutions below will
cancel out the resultant force on the stem. The main disadvantages with these solutions is
the complexity and cost.
The goal will be to design a mount for only one actuator without exerting a resulting force
on the stem to reduce the stress. The mount should take up as little space as possible, and
be mounted to the valve itself. It has to work in the designated environment for the expected
lifespan of the valve. Three solutions are presented with their advantages and disadvantages.
The circular gear mechanism is based on the slider crank principle. The actuator with
its linear movement makes the crank rotate an outer gear. Inside the outer gear there
are two gears on each side running on the inner edge shown in figure 5.12a. These are
connected to the valve house and will derive the center gear with two equal forces, working
in opposite directions. These will cancel out and the center gear will only be affected by
the moment generated by the forces. The forces is shown in figure 5.12b.
This solution is self-aligning which makes it rigged. The actuator can be mounted close to
the gears which makes this a compact solution. The moment is scalable without increasing
the total size. This is done by increasing the size of the center gear while decreasing the
size of the two gears on the side. In addition it is simple to produce, or order from a
third party. The disadvantages with this solution is its need for several bearings including
a thrust bearing which can make it a more costly solution. It also is depending on the
tolerances sensitive to debris.
(a) Circular gear mechanism (b) Forces acting
Figure 5.12: Illustration of the circular gear mechanism
50

The linear gear mechanism is based on the ”rack and pinion” principle. The outer gear
is in this solution a ”rack” which is on the end of the actuator. The two side gears must be
larger then the center gear, or the rack and side gears must operate above it. Both gears
receive a force from the ”rack” which in turn works on opposite sides and directions on the
center gear, illustrated by figure 5.13.
This solution is one of the more simpler solutions with few parts. The downside is that
it still has to have two bearings, and it loses the rigidity of the circular gear mechanism.
Because hydraulic actuators have round shafts, the ”rack” can not be contracted inside the
actuator, resulting in a longer distance from the stem to the actuator. This may become a
problem since the actuator should be mounted to the valve.
(a) Linear gear mechanism (b) Forces acting
Figure 5.13: Illustration of the linear gear mechanism
The ”wrench” mechanism is based on the idea of an independent wrench. It will be
mounted to a support structure on the valve. The ”wrench” will be fitted in a bearing fixed
to the support connected to the valve body with the valve lid bolts. This will take up the
resulting forces from the actuator, and the ”wrench” will only deliver the moment through
to the stem. The concept is shown in figure 5.14. The advantages with this mechanism will
be that it takes up a small space and it contains few and simple parts. A downside is that
it requires a large bearing, and if enough slack is present the force can create a moment on
the stem.
Figure 5.14: Illustration of the ”wrench” mechanism
51

6 Discussion
The TBV96S trunnion ball valve is in this case used for controlling the flow of mud and
had problems in both the bearings and gaskets. Through examination of the failed valve,
the most likely cause was found to be faulty assembly and usage. The group was therefore
tasked to make the changes needed such that the valve could be used in the designated
assignment.
Mud is a substance with hard particles, this eroded the valves gaskets creating leakage
between the seat and ball. The material used was found to be of a lower hardness
rating and was for that reason changed. The new material assigned for this valve was
PEEK, a high performance thermoplastic. PEEK was chosen for its higher hardness
rating and should prevent the extensive wear from the fluid. Other fluoropolymers and
elastomers was to soft for the intended usage. The cost is higher for PEEK, and if through
physical testing it is found that the material do not need the highest hardness rating,
the slightly softer PCTFE should be used. Since the mud is made on site with different
formulas for each case and phase, the hardness and wear it produces can differ considerable.
Additionally the dimension selection is mostly based on the experience and knowledge from
the engineers at Valves of Norway. For these reasons the valve and gasket should be tested
according to the ISO 10423 standard discussed in section 1.7. For situations where a low
leakage rate is key, it is important to note that softer materials commonly gives a better seal.
The other challenge in this valve was the bearings, they were broken and this was the
main reason that the valve were returned to the company. The bearing failure was
probably caused by a the faulty assembly in addition to a large actuator force. Since the
actuator force was unknown, the problem was solved through a material change. Plain
bearing was chosen over roller bearing especially for its maintenance free options which
is demanded in subsea environments. Some innovation is done with maintenance free
roller bearings, but these loose many of the positive aspects due to the material that is used.
The chosen material was the woven PTFE reinforced and bonded to a metal backing
from TENMAT called FEROGLIDE. TENMAT was the preferred supplier for Valves
of Norway, and others where therefore not researched. PTFE was chosen because it
had the necessary allowed PV factor, meaning it could handle the load and velocity on
the bearings, in addition to its temperature range and wear rate. The wear rate and
friction coefficients from TENMAT is from an unknown sample size and with an insufficient
description. The values are therefore rounded up to accommodate for the uncertainties.
The dimensions was chosen to make the changes on the stem and trunnion as subtle as
possible. Only small changes was done on the diameters of the stem, trunnion and ball hole.
The PV factor is dependent on the forces on the bearings. Identifying the magnitude of the
force on each bearing was challenging due to the complexity. There were more unknowns
than equations, and finite element analysis was therefore used to find the reaction forces
from the bearings. The result was then compared with the hand calculated ratios, the
difference was small and the results was taken as correct values. The analysis and PV
factor was only calculated for the previous bearing dimensions which could give an error,
but because of the minor changes this is likely irrelevant.
52

The changes made on the bearings and gasket affects both the trunnion, stem and springs.
In advance the actuator force was undefined and the maximum allowed force was therefore
found. The worst case scenario was chosen for the direction of the actuator force (against
flow), if other directions are used the actuator force could be of a larger magnitude and
the stress distribution could be different.
Friction torques and stresses on the different sections was calculated with hand and FEA
calculations. Unconventional formulas has been used for many of the friction torque
calculations, this in agreement with the engineer at Valves of Norway. Although the
formulas used are believed to be of a sufficient accuracy, a suggestion for further research is
to find more well documented and reasoned formulas. Additionally the friction coefficients
from the bearings is assumed to be 0.1 because they run in mud. This is possibly a high
value, and if testing is done it may be reduced significantly.
The result showed that only the trunnion, trunnion bearing and upper stem bearing was
close to their material limit and these should either be increased or the other sections
should be reduced to make the design more optimized. Increasing the parts will give the
valve the ability to handle a larger actuator force and reduce the arm length if this is
desirable. The valve as is will need a minimum arm length of 75 mm with a safety factor
of 1.5 on the friction torques. When designing the mouth this value did not create any
problems. On that basis a suggestion might be to reduce all components to optimize the
material usage of the valve.
Several solutions for the mount were proposed, including mounts that will reduce the
resultant force on the stem to zero and thereby increasing the safety factor drastically.
The concept that was further designed on was the lever principle for a hydraulic actuator,
assuming access to an hydraulic systems. In later years electric actuators have been
manufactured for subsea use. These actuators deliver a different set of solutions for mounts,
and if this is used in the application, a new mount should be designed.
A bend was added to the mount to give the ability to start perpendicular to the stem,
such that the force tangent to the stem was highest at the starting point were the friction
torques was largest. The part was 3D modeled along with an actuator, and these were
used to simulate their behavior. This type of simulations is cost efficient, and can reduce
the time used for prototyping to a great extent.
When stress analyzing the mount, the height was set to 40 mm from the top as an estimate.
The height is actually adjustable and if placed higher the result should be re assessed. The
mount was made an amount to strong, and further optimizing could be done, but because
of singularities it then should be physically tested before use.
53

7 Conclusion
This report is the bachelor thesis in the mechanical engineering study with specialization
in marine structural engineering, and addresses bearing and gasket failures with the valve
TBV96S which is constructed by Valves of Norway. The task is given by Valves of Norway
a valve producing company located by the west coast of Norway, outside Bergen.
A new material is chosen for both the gaskets and bearings in the valve. This change should
be suﬃcient in avoiding a repeated failure of the valve. The valve is stress analyzed and
all components are established to be capable of handling the work pressure of 10 MPa in
addition to an actuator force of up to 30 kN. All calculations follows the recommendations
from ISO 10423. In case of a situations where a higher safety factor or a smaller arm is
needed several proposals for mounts that will cancel out the actuator force are suggested.
This thesis should give a clear and reasoned description of the selection process for the
gasket and bearing which can be used in all types of applications, aiding engineers to take
a good material choice for further projects. Although other projects may have diﬀerent
priorities, all aspects should have been mentioned. The proposed mount solutions is also
believed to be an innovative solutions giving the highest force were needed, but still be
simple in both the mechanism and production. The thesis also addresses how to calculate
torsion for both uniform and un-uniform geometry, in addition to average and shear stress
and lastly moment stress.
A large amount of time should be given to the selection and calculation process, which
will greatly reduce the cost in physical prototypes and returns. Digital tools should also
be taken to use, they can improve the optimization of a product even before a physical
model is made as well as calculating problems with considerable complexity. This thesis
shows how digital tools can be used as an extension of the hand calculations to achieve
more accurate results and solutions.
Even though with the high complexity of the task and the lack of well documented formulas,
the results shows great precision and the group considers the task to be achieved.
54

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56

A-A ( 1 : 5 )
A A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Weight
Material
Helene F. S.
Date
Unless otherwise specified:
Break sharp corners 3 Fillet radii 0.8
Surface Ra 1.6 Tolerances ISO 2768-f
This document contains Norske Ventiler AS
proprietary and confidential information. It is loaned
for limited purposes only and remains the property
of Norske Ventiler AS. It may not be re-produced in
whole or part or disclosed to third parties without
the prior writen consent of Norske Ventiler AS. The
document is to be returned to Norske Ventiler AS
upon request and in all events upon completion of
the use for which it was loaned.
Originator
Checked by
Helene F. S.
Design state
Date
02.05.2016
02.05.2016
Date
Date
02.05.2016
H kon M.
A479 UNS S31600
N/A
02.05.2016
TITEL
Scale:
DOCUMENT NO.
Valve House
1:5
NV 200009
V A L V E S OF NOR WA Y
NORSKEVENTILER AS
Size:
A3
REV.
www.norskeventiler.no
Release Date
Approved by
Modified by
Helene
2
M24 40 (x 8)
5 12 (x 40)
M12 40 (x 8)
100
215,9
114
25 (x 8)
215,9
R5
1 X 45
50
51,46
72
94
BOTTOM
6,65 12 (x 10)
TOP
6
9
55,54
178
180,91
296
36
145
147
161
163
96
126,2
127,66
136
140
150
275
160
20
34
7
45,1
29,50
283
186
200,5
4
R89,01
R150,02
72

A-A ( 1 : 2 )
A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Release Date
02.05.2016
Date
Helene F. S.
Design state
Originator
Weight
Date
N/A
02.05.2016
Material
Scale:
Helene F. S.
A479 UNS S31600
Date
02.05.2016
02.05.2016
TITEL
Valve Lid
1:2
NORSKEVENTILER AS
Checked by
H kon M.
Modified by
Helene
DOCUMENT NO.
NV 200010
Approved by
Break sharp corners 0.2 Fillet radii 0.8
Date
Size:
A3
REV.
2
25
5 12 (x 40)
96
97
126,2
127,66
136
150
57,532
275 7
45,1
82,5
106,5
97
5
7
12,1
17,1
24,2
149
28
32
3 X 45
R5
R1,2
R1,2
1 X 45
1 X 45
1 X 45114
215,9

A-A ( 1 : 1 )
A A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Design state
Approved by
Checked by
Einar K.
Release Date
Date
30.04.2016
DOCUMENT NO.
the use for which it was loaned. Scale:
1,8kg
Helene F. S.
Originator
Date
29.04.2016
29.04.2016
TITEL
Date
Material
1:1
Trunnion Lid
H kon M.
Helene F. S.
A479 UNS S31600
NV200001
NORSKEVENTILER AS
Modified by
15.04.2016
2
A3 Sheet 1 of 1Size:
REV.
Weight
Date
92
19
28
3 X 45
0,5 X 45
40h7
0,5 X 45 (x 2)
4,8 4,8
36,8
64
84
44,8
50h7
72
12
14
6,65
9
15
9
20

A-A ( 1 : 2 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Date
Date
01.05.2016
-
TITEL
01.05.2016
01.05.2016
A3Size:
NV 200000
Date
A479 UNS S31600
Helene
REV.
NORSKEVENTILER AS
1:2
Stem Lid
01.05.2016
DOCUMENT NO.
Scale:
Approved by
H kon M.
Checked by
Helene F. S.
Modified by
Originator
Weight
Material
Design state
Date
Release Date
Helene F. S.
1
20 13 (x 8)
10 13 (x 2)
20 (x 8)
10,11 (x 12)
180
51
73
74
52
4 8
28
0,5 X 45 (x 3)
1 X 45 (x 3)
152
100

B-B ( 1 : 2 )
B B
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Design state
Approved by
Modified by
NORSKEVENTILER AS
20367 1
Checked by
Helene
Ball
Date
Helene
Date
1:2www.norskeventiler.no Scale:
15.04.2016
Weight
12,5kg
Material
A276 S32750
Originator
Sheet 1 of 1
Date
Release Date
A3
DOCUMENT NO.
TITEL
Size:
REV.Date
36,00
- 0,01
0,03+
R43x
44 H7
2019
29
96
8484
103,95()
R2,5
2,5
R88
101
17652,46
45 - 0
0,07+
142,99( )
0,4

A ( 5:1 )
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Date
Material
Design state
Helene F.S.
Modified by
H kon M.
Checked by
Einar K.
Scale:
Approved by
DOCUMENT NO.
Release Date
1:2
30.04.2016
Originator
Weight
Stem
NORSKEVENTILER AS
NV200012
A3
2
Size:
REV.
Date
15.04.2016Helene F.S.
2,5kg
29.04.2016
Date
29.04.2016
TITEL
A479 UNS S31600
Sheet 1 of 1
Date
38-0,01
0,05-
38 - 0,01
0,05-
2
50
35,4
- 0,01
0,05-
50
72
49,8
45 h7
38,8 0,05
0,5 X 45 0,3
1
0,5 X 45
0,5 X 45
2
80
89,5
10
29,5
61
65
69,2
73,8
78
81
89
188,5
92,5
49,4
- 0,02
0,01+
R4

A-A ( 1 : 1 )
C ( 2 : 1 )
A
A
C
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Release Date
30.04.2016
Date
29.04.2016
TITEL
1,3kg
Helene F. S.
Date
09.04.2016
DOCUMENT NO.
Date
29.04.2016
H kon M.
Originator
NV200163
Seat
NORSKEVENTILER AS
Modified by
Helene F. S.
A276 S32750
Approved by
Material
Scale:
Design state
Einar K.
Checked by
2
Sheet 1 of 1A3Size:
REV.
Weight
Date
38,5 0,05
97
0,5 X 45
0,5 X 45
3,7 (x 2)
7,8
12,8
1:1
126,0
-0,01
0,06+
0,65
0,5 X 45 (x 2)R1,6 11
14
4,5
12,72( )
96
115,6
1260,05
135,6
116,90,05
5,3
10,45

A-A ( 1 : 1 )
A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
29.04.2016
Date
Date
30.04.2016H kon M.
Release Date
Date
Date
Helene F. S.
29.04.2016
Checked by
Einar K.
Approved by
Helene F. S.
Modified by
Material
Originator
Design state
19.04.2016
Weight
N/A
TITEL
PEEK
REV.DOCUMENT NO.
Size:
A3
the use for which it was loaned. Scale: 1:1 Sheet 1 of 1
Gasket
2NV200164
NORSKEVENTILER AS
R0,75 (x 4)
116,7
11,5
1
0,27
126,2

A-A ( 1:1 )
A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
30.04.2016
Release Date
DOCUMENT NO.
23.04.2016
PTFEThe bearing is bought by TENMAT and therefore
this is a dummy 2D drawing.
29.04.2016
Bearing Upper Stem
TITEL
REV.
NORSKEVENTILER AS
1
Sheet 1 of 1
-
Scale: 1:1
29.04.2016
Date
A3Size:
Helene F. S.
Einar K.
Originator
H kon M.
Modified by
Checked by
Approved by
Design state
Material
Date
Helene F. S.
Date
Weight
Date
N/A
50
45
7,8
Bearing Upper Stem
Type: Plain linear self-lubricating
FEROGLIDE
PA1 JOURNAL BEARINGS COILED
Standard Backing Metals:
1) Zinc coated mild steel
2) Stainless Steel (AISI 316)
3) Inconel 625 id
Sizes and Tolerances to DIN1494
Inner diameter tolerance H10 when
installed in housing.
Material: Woven PTFE fabric with
strengthening fibers applied to a metal
backing.
Part nr.: On request
Supplier: TENMAT Ltd.
For more information see
http://www.tenmat.com/Content/Feroglide

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Date
Date
N/A
Date
Helene F. S.
Design state
Approved by
Modified by
Checked by
Weight
Material
The bearing is bought by TENMAT and therefore
Originator
H kon M.
Einar K.
Helene F. S.
29.04.2016
23.04.2016
30.04.2016
PTFE
-
NORSKEVENTILER AS
Bearing Lower Stem
A3
1
Size:
TITEL
Scale:
DOCUMENT NO.
Release Date
29.04.2016
Date
Sheet 1 of 11:1
REV.
50
45
30
Bearing Lower Stem
FEROGLIDE
3) Inconel 625 id
backing.
Part nr.: P-A1-45.30

A-A ( 1:1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Date
Date
N/A
Date
Helene F. S.
Design state
Approved by
Modified by
Checked by
Weight
Material
The bearing is bought by TENMAT and therefore
Originator
H kon M
Einar K.
Helene F. S.
29.04.2016
23.04.2016
30.04.2016
PTFE
-
NORSKEVENTILER AS
Bearing Trunnion
A3
1
Size:
TITEL
Scale:
DOCUMENT NO.
Release Date
29.04.2016
Date
Sheet 1 of 11:1
REV.
44
40
20
Bearing Trunnion Lid
FEROGLIDE
3) Inconel 625 id
backing.
Part nr.: P-A1-45.20

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
N/A
Date
Date
Helene
Date
Weight
Material
This is a dummy 2D drawing.
Originator
Design state
Approved by
Checked by
Modified by
Helene F. S.
H kon M.
Helene F. S.
02.05.2016
02.05.2016
02.05.2016
PETP
1
Thrust Bearing
Size:
A3 Sheet 1 of 1
NORSKEVENTILER AS
20562
TITEL
DOCUMENT NO.
Release Date
02.05.2016
Date
1:1Scale:
REV.
72,5
50,5
0,5 X 45
2

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
TFM
NORSKEVENTILER ASOriginator
Weight
Material
Helene F. S.
H kon M.
Helene F. S.
Seat Glide Ring
1:1
20358
Scale:
01.05.2016
01.05.2016
01.05.2016
01.05.2016
Release Date
Date
A3Size: Sheet 1 of 1
Checked by
1
REV.
Design state
Modified by
Approved by
TITEL
N/A
Date
Date
Helene
DOCUMENT NO.
Date
0,4 X 45
3,4
121
126,2

A-A ( 1 : 1 )
A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
TFMUnless otherwise specified:
Date
Date
N/A
Date
01.05.2016
DOCUMENT NO.
TITEL
Date
01.05.2016
Stem Glide Ring
Scale: 1:1
Release Date
Material
Helene
Approved by
Helene F. S.
Checked by
Modified by
Originator
Weight
NORSKEVENTILER AS
Design state
Helene F. S.
H kon M.
20561
01.05.2016
01.05.2016
REV.
1
Sheet 1 of 1Size:
A3
44,6
50
3,9
0,4 X 45

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Originator
Weight
Material
NBR 70
Approved by
Checked by
Helene F. S.
H kon M.
Helene F. S.
Modified by
Date
Helene
N/A
Date
NORSKEVENTILER AS
Design state -
O-Ring 170,82 x 5,33
DOCUMENT NO.
A31:1 Size:
REV.
1
Sheet 1 of 1
01.05.2016
TITEL
01.05.2016
01.05.2016
Release Date
Date
Date
Scale:
181,48
170,82
5,33

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Date
DOCUMENT NO.
01.05.2016
Scale:
Date
01.05.2016
01.05.2016
TITEL
01.05.2016
O-Ring 113,89 x 3,53
-
1:1
1
Sheet 1 of 1Size:
Release Date
NORSKEVENTILER AS
A3
REV.
Date
Date
N/A
Helene
Modified by
Helene F. S.
Helene F. S.
Design state
H kon M.
Checked by
Approved by
NBR 70
Material
Weight
Originator
120,95
113,89
3,53

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
Scale:www.norskeventiler.no
01.05.2016
-
Date
Date
01.05.2016
01.05.2016
Release Date
O-Ring 44,45 x 3,53
Date
1:1 A3
DOCUMENT NO.
Size:
1
01.05.2016
TITEL
NORSKEVENTILER AS
REV.
Helene F. S.
Design state
Approved by
N/A
H kon M.
Weight
Material
Modified by
Checked by
Date
Helene
Helene F. S.
Sheet 1 of 1
Originator
NBR 70
51,51
44,45
3,53

A-A ( 1 : 1 )A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
-
Size: Sheet 1 of 1A3
O-Ring 37,69 x 3,53
REV.
Scale:
01.05.2016
01.05.2016
Date
Date
01.05.2016
www.norskeventiler.no 1:1
Release Date
01.05.2016
Date
Date
TITEL
Helene
H kon M.
Originator
Helene F. S.
Checked by
Approved by
Helene F. S.
Material
N/A
Weight
Design state
NORSKEVENTILER AS
Modified by
DOCUMENT NO.
NBR 70
44,75
37,69
3,53

A-A ( 1 : 2 )
B ( 1 : 1 )
A A
B
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
A479 UNS S31600
H kon M.
Date
2,3 kg
Date
Checked by
Modified by
Design state
Approved by
Material
Weight
Helene F. S.
Originator
H kon M.
19.04.2016
Date
Einar K.
30.04.2016
DOCUMENT NO.
TITEL
1:2Scale:
1
NORSKEVENTILER AS
001
Sheet 1 of 1
Mount
A3Size:
REV.
Release Date
Date
29.04.2016
29.04.2016
25
2x 1 X 45
R45
R170
38 + 0,01
0,05+
38 + 0,01
0,05+
269,93
180
R117,95
R132,12
120 60
R80
R45
R54x
15
15
21,62
34,24
0,82
1,57
13,35
12
6,6
0,15
3
21,5
6,15 M6
90
1 X 45 (x 2)
0,96 X 46,27 (x 2)

A-A ( 1 : 4 )
PARTS LIST
MATERIALTITLEPART NUMQTYITEM
A479 UNS S31600Valve HouseNV 20000911
A479 UNS S31600Valve LidNV 20001012
A276 S32750Ball2036713
UNS NO7718StemNV 20001214
A479 UNS S31600Stem LidNV 20000015
A479 UNS S31600Trunnion LidNV 20000116
A276 S32750SeatNV 20016327
PEEKSeat gasketNV 20016428
FEROGLIDE woven PTFEBearing Lower Stem19
FEROGLIDE woven PTFEBearing Upper Stem110
PETPThrust Bearing20562211
FEROGLIDE woven PTFEBearing Trunnion112
TFMSeat Glide Ring20358413
TFMStem Glide Ring20561214
UNS NO6625Spring103732B8015
NBR 70O-Ring 170,82 x 5,33216
NBR 70O-Ring 113,89 x 3,53417
NBR 70O-Ring 44,45 x 3,53218
NBR 70O-Ring 37,69 x 3,53219
A4-80ISO 4762 - M8 x 201020
A4-80ISO 4762 - M24 x 40821
A4-80ISO 4762 - M12 x 40822
A479 UNS S31600Mount001123
A4-80ISO 4762 - M6 x 35124
A
A
1
1
2
2
3
3
4
4
5
5
6
6
A A
B B
C C
D D
TVB96S
Sheet 1 of 1
NORSKEVENTILER AS
A3
Modified by
Checked by
Originator Date
19.04.2016
124 kg
Date
Helene F. S.
Helene F. S.
Einar K.
H kon M.
Approved by
Design state
Weight
Material
4
29.04.2016
Date
TITEL
DOCUMENT NO.
30.04.2016
Size:
Date
29.04.2016
REV.
1:4
Release Date
Scale:
258x
Design Code: ISO 13628
Working Pressure: 10 MPa
Operating Pressure: 20 MPa
Body Test Pressure: 15 MPa
Design Temperature: 0 to 70 C
Maximum Depth: 1000 m
Required Operating Torque: 4000 Nm
Interface: 4" ASME B16,5 CL600 RF
M12 X 1,75 (x 12)
10 H12 -10 DEEP (x 2)
22
21
215,9
300
1 5 11 4 23 10 19 14
9
16
8
13
17
15
7
2
1820612
3
7924
244
165
150
394
152
100
Actuator pin
275
300
375
738,1037,4
400
96
24

NUMBER
PART
Information in this drawing is provided for reference only.
http://www.mcmaster.com
10 mm
5 mm
Hex
6 mm
6 mm 35 mm
Thread length may vary from
24 mm to fully threaded in length.
M6 x 1 mm Thread
92290A334
Stainless Steel
Socket Head Cap Screw
© 2014 McMaster-Carr Supply Company

NUMBER
PART
36 mm
19 mm
Hex
24 mm
24 mm 50 mm
M24 x 3 mm Thread
91292A282
Stainless Steel

NUMBER
PART
18 mm
10 mm
Hex
12 mm
12 mm 40 mm
M12 x 1.75 mm Thread
92290A622
Stainless Steel

NUMBER
PART
13 mm
6 mm
Hex
8 mm
8 mm 20 mm
M8 x 1.25 mm Thread
92290A426
Stainless Steel

Appendices
A Tolerances from ISO-2768
General Tolerances to DIN ISO 2768
• The latest DIN standard sheet version applies to all parts made to DIN standards.
• Variations on dimensions without tolerance values are according to "DIN ISO 2768- mk".
GENERAL TOLERANCES FOR LINEAR AND ANGULAR DIMENSIONS (DIN ISO 2768 T1)
LINEAR DIMENSIONS:
Tolerance class
designation (description)
Permissible deviations
in mm for ranges in
nominal lengths f (fine)
m (medium) c (coarse)
v (very coarse)
0.5 up to 3 ±0.05 ±0.1 ±0.2 -
over 3 up to 6 ±0.05 ±0.1 ±0.3 ±0.5
over 6 up to 30 ±0.1 ±0.2 ±0.5 ±1.0
over 30 up to 120 ±0.15 ±0.3 ±0.8 ±1.5
over 120 up to 400 ±0.2 ±0.5 ±1.2 ±2.5
over 400 up to 1000 ±0.3 ±0.8 ±2.0 ±4.0
over 1000 up to 2000 ±0.5 ±1.2 ±3.0 ±6.0
over 2000 up to 4000 - ±2.0 ±4.0 ±8.0
EXTERNAL RADIUS AND CHAMFER HEIGHTS
Tolerance class
in mm for ranges in
nominal lengths f (fine)
m (middle) c (coarse)
v (very coarse)
0.5 up to 3 ±0.2 ±0.2 ±0.4 ±0.4
over 3 up to 6 ±0.5 ±0.5 ±1.0 ±1.0
over 6 ±1.0 ±1.0 ±2.0 ±2.0
ANGULAR DIMENSIONS
Tolerance class
in degrees and minutes
for ranges in nominal
lengths f (fine) m (middle) c (coarse) v (very coarse)
up to 10 ±1º ±1º ±1º30' ±3º
over 10 up to 50 ±0º30' ±0º30' ±1º ±2º
over 50 up to 120 ±0º20' ±0º20' ±0º30' ±1º
over 120 up to 400 ±0º10' ±0º10' ±0º15' ±0º30'
over 400 ±0º5' ±0º5' ±0º10' ±0º20'
xxvi

GENERAL TOLERANCES FOR FORM AND POSITION (DIN ISO 2768 T2)
STRAIGHTNESS AND FLATNESS
Tolerance classRanges in nominal
lengths in mm
H K L
up to 10 0.02 0.05 0.1
over 10 up to 30 0.05 0.1 0.2
over 30 up to 100 0.1 0.2 0.4
over 100 up to 300 0.2 0.4 0.8
over 300 up to 1000 0.3 0.6 1.2
over 1000 up to 3000 0.4 0.8 1.6
PERPENDICULARITY
lengths in mm
H K L
up to 100 0.2 0.4 0.6
over 100 up to 300 0.3 0.6 1
over 300 up to 1000 0.4 0.8 1.5
over 1000 up to 3000 0.5 0.8 2
SYMMETRY
lengths in mm
H K L
up to 100 0.5 0.6 0.6
over 100 up to 300 0.5 0.6 1
over 300 up to 1000 0.5 0.8 1.5
over 1000 up to 3000 0.5 1 2
RUN-OUT
Tolerance class
H K L
0.1 0.2 0.5
xxvii

B Bearing material specs
xxviii

C FEROGLIDE dimensions
FEROGLIDE
3) Inconel 625
Inner diameter tolerance H10 when installed in housing. Parts maybe out of round in their free state.
-----------------------------------------------------------------------------------------------------------------------------------------------
Dash Inner Outer Length Weight
No. d D L g
-----------------------------------------------------------------------------------------------------------------------------------------------
P-A1-24.25 24 27 25 21.9
P-A1-24.30 24 27 30 26.3
P-A1-25.15 25 28 15 13.1
P-A1-25.20 25 28 20 18.3
P-A1-25.25 25 28 25 22.9
P-A1-25.30 25 28 30 27.6
P-A1-25.50 25 28 50 46.0
P-A1-28.20 28 32 20 27.7
P-A1-28.30 28 32 30 41.9
P-A1-30.15 30 34 15 22.4
P-A1-30.20 30 34 20 29.6
P-A1-30.25 30 34 25 37.0
P-A1-30.30 30 34 30 44.7
P-A1-30.40 30 34 40 59.7
P-A1-32.20 32 36 20 31.4
P-A1-32.30 32 36 30 47.5
P-A1-32.40 32 36 40 63.5
P-A1-35.20 35 39 20 34.2
P-A1-35.30 35 39 30 51.6
P-A1-35.40 35 39 40 68.4
P-A1-35.50 35 39 50 86.5
P-A1-40.20 40 44 20 38.8
P-A1-40.30 40 44 30 58.6
P-A1-40.40 40 44 40 77.6
P-A1-40.50 40 44 50 98.2
P-A1-45.20 45 50 20 55.1
P-A1-45.30 45 50 30 83
P-A1-45.40 45 50 40 111
P-A1-45.50 45 50 50 139
P-A1-50.20 50 55 20 62
P-A1-50.30 50 55 30 93
P-A1-50.40 50 55 40 123
P-A1-50.60 50 55 60 185
P-A1-55.40 55 60 40 137
P-A1-55.60 55 60 60 202
P-A1-60.30 60 65 30 109
P-A1-60.40 60 65 40 149
P-A1-60.60 60 65 60 220
P-A1-60.70 60 65 70 265
P-A1-65.50 65 70 50 198
P-A1-65.70 65 70 70 277
P-A1-70.40 70 75 40 170
P-A1-70.50 70 75 50 212
TENMAT LTD, Ashburton Road West, Trafford Park, Manchester, M17 1RU, 0161-872-2181 www.tenmat.com
Issue 2
5
D
d
FEROGLIDE LINER
L
xxix

E Previous 2D assembly drawing
xxxi

F Metric tap- and clearance drill dimensions
Tap Drill Clearance Drill
Metric Tap &
Clearance Drill
Sizes
75% Thread for
Aluminum, Brass,
& Plastics
50% Thread for
Steel, Stainless,
& Iron
Close Fit Standard Fit
Screw Size
(mm)
Thread
Pitch (mm)
Drill Size
(mm)
Closest
American
Drill
Drill Size
(mm)
Closest
American
Drill
Drill Size
(mm)
Closest
American
Drill
Drill Size
(mm)
Closest
American
Drill
M1.5 0.35 1.15 56 1.25 55 1.60 1/16 1.65 52
M1.6 0.35 1.25 55 1.35 54 1.70 51 1.75 50
M 1.8 0.35 1.45 53 1.55 1/16 1.90 49 2.00 5/64
0.45 1.55 1/16 1.70 51
M 2
0.40 1.60 52 1.75 50
2.10 45 2.20 44
M 2.2 0.45 1.75 50 1.90 48 2.30 3/32 2.40 41
M 2.5 0.45 2.05 46 2.20 44 2.65 37 2.75 7/64
0.60 2.40 41 2.60 37
M 3
0.50 2.50 39 2.70 36
3.15 1/8 3.30 30
M 3.5 0.60 2.90 32 3.10 31 3.70 27 3.85 24
0.75 3.25 30 3.50 28
M 4
0.70 3.30 30 3.50 28
4.20 19 4.40 17
M 4.5 0.75 3.75 25 4.00 22 4.75 13 5.00 9
1.00 4.00 21 4.40 11/64
0.90 4.10 20 4.40 17M 5
0.80 4.20 19 4.50 16
5.25 5 5.50 7/32
M 5.5 0.90 4.60 14 4.90 10 5.80 1 6.10 B
1.00 5.00 8 5.40 4
M 6
0.75 5.25 4 5.50 7/32
6.30 E 6.60 G
1.00 6.00 B 6.40 E
M 7
0.75 6.25 D 6.50 F
7.40 L 7.70 N
1.25 6.80 H 7.20 J
M 8
1.00 7.00 J 7.40 L
8.40 Q 8.80 S
1.25 7.80 N 8.20 P
M 9
1.00 8.00 O 8.40 21/64
9.50 3/8 9.90 25/64
1.50 8.50 R 9.00 T
1.25 8.80 11/32 9.20 23/64M 10
1.00 9.00 T 9.40 U
10.50 Z 11.00 7/16
M 11 1.50 9.50 3/8 10.00 X 11.60 29/64 12.10 15/32
1.75 10.30 13/32 10.90 27/64
1.50 10.50 Z 11.00 7/16M 12
1.25 10.80 27/64 11.20 7/16
12.60 1/2 13.20 33/64
2.00 12.10 15/32 12.70 1/2
1.50 12.50 1/2 13.00 33/64M 14
1.25 12.80 1/2 13.20 33/64
14.75 37/64 15.50 39/64
M 15 1.50 13.50 17/32 14.00 35/64 15.75 5/8 16.50 21/32
2.00 14.00 35/64 14.75 37/64
M 16
1.50 14.50 37/64 15.00 19/32
16.75 21/32 17.50 11/16
M 17 1.50 15.50 39/64 16.00 5/8 18.00 45/64 18.50 47/64
2.50 15.50 39/64 16.50 41/64
2.00 16.00 5/8 16.75 21/32M 18
1.50 16.50 21/32 17.00 43/64
19.00 3/4 20.00 25/32
M 19 2.50 16.50 21/32 17.50 11/16 20.00 25/32 21.00 53/64
2.50 17.50 11/16 18.50 23/32
2.00 18.00 45/64 18.50 47/64M 20
1.50 18.50 47/64 19.00 3/4
21.00 53/64 22.00 55/64
the premier source of parts and accessories for mini lathes and mini mills
LittleMachineShop.com • (800) 981-9663
xxxii

G 316 Stainless steel properties
http://www.azom.com/article.aspx?ArticleID=863
xxxiii

H Friction coeﬃcient for stainless steel bolts
I Torque Constant K
xxxiv

J Email regarding PEEK
Vennlig hilsen/ Best regards
Vladimir Mitin
Design engineer
Mob.
xxxv

K Calculations torsion, shear and moment.
Calculations of the stresses in each section were Q1 and U1 is the von mises and utility for
the closed position, and Q2 and U2 is for the open position.
K.1 Reaction forces and moments
Fact = 30000 N (K.1)
Fus = 31599 N (K.2)
Fls = 44438 N (K.3)
Ft = 81852 N (K.4)
Fb = 124689 N (K.5)
MusF EM = 8890 Nmm (K.6)
MlsF EM
= 584308 Nmm (K.7)
MtF EM = −604929 Nmm (K.8)
K.2 Shear forces and moments
VA = Ft = 81852 N (K.9)
VB = VA − Fb = −42837 N (K.10)
VC = VB + Fls = 1601 N (K.11)
VD = VC − Fus = −29998 N (K.12)
VF = VD + Fact = 2 N (K.13)
M1 = VA · L1 + MtF EM = 5533971 Nmm (K.14)
M2 = M1 + VB · L4 = 1250271 Nmm (K.15)
M3 = M2 + MlsF EM
= 1834579 Nmm (K.16)
M4 = M3 + VC · L5 = 1896858 Nmm (K.17)
M5 = M4 + MusF EM = 1905748 Nmm (K.18)
M6 = M5 + VD · L6 = 3875 Nmm (K.19)
Mh = M1 + VB · L2 = 3135099 Nmm (K.20)
Mlc = M1 + VB · L3 = 2706729 Nmm (K.21)
Mcs = M2 + VC · Lcs = 1297901 Nmm (K.22)
xxxvi

K.3 Friction torques
dus = 50 mm (K.23)
dls = 50 mm (K.24)
dt = 44 mm (K.25)
µus = 0.03 (K.26)
µls = 0.1 (K.27)
µt = 0.1 (K.28)
Dg = 126 mm (K.29)
dg = 116.7 mm (K.30)
p = 10 N/mm2
(K.31)
Apg =
π · D2
g − d2
g
4
= 1773 mm2
(K.32)
Fpg = Apg · p = 17727 N (K.33)
Fsg = 0.05 · 0.6 · 110 N/mm2
· Apg = 5850 N (K.34)
Db = 176 mm (K.35)
y = 0.4 (K.36)
Mus =
Fus
2
· uus · dus = 21329 Nmm (K.37)
Mls =
Fls
2
· uls · dls = 99986 Nmm (K.38)
Mt =
Ft
2
· ut · dt = 163704 Nmm (K.39)
Mss = 4875 Nmm (K.40)
Mtb = 60377 Nmm (K.41)
Mpg = Fpg · y ·
1
2
· Db ·
(1 + sin(50)) · 0.5)
sin(50) + 0.05 · cos(50)
= 690326 Nmm (K.42)
M2pg = 2 · Mpg = 1380651 Nmm (K.43)
Msg = 2 · Fsg · y ·
1
2
· Db ·
(1 + sin(50)) · 0.5)
sin(50) + 0.05 · cos(50)
= 455615 Nmm (K.44)
xxxvii

K.4 Allowed stress
σstemallowed
=
827
1.5
= 551.33 N/mm2
(K.45)
σtrunnionallowed
=
204
1.5
= 136 N/mm2
(K.46)
K.5 Torsion
Torsion for the diﬀerent sections that are checked. For both the open and closed position.
The values for moments used in calculations below is from section 4.1.
Tt = Mt = 163704 Nmm (K.47)
Th = Mt + Mpg + Msg = 1308645 Nmm (K.48)
Tlc = Mt + Mpg + Msg = 1308645 Nmm (K.49)
Tls = Mt + Mls + Mpg + Msg = 1409630 Nmm (K.50)
Tus = Mt + Mls + Mpg + Msg + MgrMus = 1435834 Nmm (K.51)
Ttot = Mt + Mpg + Msq + Mls + Mgr + Mus + Mtb = 1496211 Nmm (K.52)
Tt = 0 Nmm
Th = 2 (Mpg + Msg) = 1836266 Nmm
Tlc = 2 (Mpg + Msg) = 1836266 Nmm
Tls = 2 (Mpg + Msg) = 1836266 Nmm
Tus = 2 (Mpg + Msg) + Mgr = 1841141 Nmm
Ttot = 2 (Mpg + Msg) + Mgr + Mtb = 1901518 Nmm
xxxviii

K.6 Stress lower Hexagon
wh = 35.8 mm (K.53)
ah = 29.67 mm (K.54)
yh =
w
2
= 17.9 mm (K.55)
Ih =
5 ·
√
3 · a4
h
16
= 419451 mm4
(K.56)
Qh =
1
8
· ah · w2
h +
1
12
· w2
h · a2
h −
wh
2
2
= 7280 mm3
(K.57)
qh =
Mh
Ih
· yh = 134 N/mm2
(K.58)
ths =
VB · Qh
Ih · wh
= −21 N/mm2
(K.59)
tht = 5.297 ·
Thx
w3
h
= 151 N/mm2
(K.60)
thtot = tht − ths = 172 N/mm2
(K.61)
Qh1 = (qh)2 + 3 · (thtot)2 = 327 N/mm2
(K.62)
Uh1 = Qh1/Qstemallowed
· 100% = 67% (K.63)
thtot2 = 5.297 ·
Thx2
w3
h
= 212 N/mm2
(K.64)
Qh2 = 3 · (thtot2)2 = 367 N/mm2
(K.65)
Uh2 = Qh2/Qstemallowed
· 100% = 67% (K.66)
xxxix

K.7 Stress lower circular section
dlc = 45 mm (K.67)
Wlc =
pi
32
· (dlcmm)3
= 8946 mm3
(K.68)
Alc = pi ·
1
4
· d2
lc = 1590 mm2
(K.69)
qlc =
Mlc
Wlc
= 303 N/mm2
(K.70)
tlcs =
16
3
·
VB
π · d2
lc
= −36 N/mm2
(K.71)
tlct =
16 · Tlc
π · d3
lc
= 73 N/mm2
(K.72)
tlctot = tlct − tlcs = 109 N/mm2
(K.73)
Qlc1 = (qlc)2 + 3 · (tlctot)2 = 357 N/mm2
(K.74)
Ulc1 =
Qlc1
Qstemallowed
· 100 % = 65% (K.75)
tlct2 =
16 · Tlc2
π · d3
lc
= 103 N/mm2
(K.76)
Qlc2 = 3 · (tlct2)2 = 178 N/mm2
(K.77)
Ulc2 =
Qlc2
Qstemallowed
· 100% = 32% (K.78)
xl

K.8 Stress lower stem
dls = 45 mm (K.79)
Wls =
pi
32
· (dls)3
= 8946 mm3
(K.80)
Als = pi ·
1
4
· d2
ls1590 mm2
(K.81)
qls =
M2
Wls
= 140 N/mm2
(K.82)
tlst =
16 · Tls
π · d3
ls
= 79 N/mm2
(K.83)
Qls1 = (qls)2 + 3 · (tlst)2 = 195 N/mm2
(K.84)
Uls1 =
Qls1
Qstemallowed
· 100% = 35% (K.85)
tlst2 =
16 · Tls2
π · d3
ls
= 103 N/mm2
(K.86)
Qls2 = 3 · (tlst2)2 = 178 N/mm2
(K.87)
Uls2 =
Qls2
Qstemallowed
· 100% = 32% (K.88)
xli

K.9 Stress mid upper seal/glide ring
dcs = 38.8 mm (K.89)
Wcs =
pi
32
· (dcsmm)3
= 5734 mm4
(K.90)
Acs = pi ·
1
4
· d2
cs1182 mm2
(K.91)
qcs =
Mcs
Wcs
= 226 N/mm2
(K.92)
tcss =
16
3
·
VC
π · d2
cs
= 2 N/mm2
(K.93)
tcst =
16 · Tcs
π · d3
cs
= 123 N/mm2
(K.94)
tcstot = tcst − tcss = 122 N/mm2
(K.95)
Qcs1 = (qcs)2 + 3 · (tcstot)2 = 309 N/mm2
(K.96)
Ucs1 =
Qcs1
Qstemallowed
· 100% = 56% (K.97)
tcst2 =
16 · Tcs2
π · d3
cs
= 161 N/mm2
(K.98)
Qcs2 = 3 · (tcst2)2 = 278 N/mm2
(K.99)
Ucs2 =
Qcs2
Qstemallowed
· 100% = 50% (K.100)
xlii

K.10 Stress upper stem
dus = 45 mm (K.101)
Wus =
pi
32
· (dusmm)3
= 8946 mm3
(K.102)
Aus = pi ·
1
4
· d2
us = 1590 mm2
(K.103)
qus =
M3
Wus
= 205 N/mm2
(K.104)
tust =
16 · Tus
π · d3
us
= 80 N/mm2
(K.105)
Qus1 = (qus)2 + 3 · (tust)2 = 248 N/mm2
(K.106)
Uus1 =
Qus1
Qstemallowed
· 100% = 45% (K.107)
tust2 =
16 · Tus2
π · d3
us
= 103 N/mm2
(K.108)
Qus2 = 3 · (tust2)2 = 178 N/mm2
(K.109)
Uus2 =
Qus2
Qstemallowed
· 100% = 32% (K.110)
xliii

K.11 Stress trunnion
dt = 40 mm (K.111)
Wt =
pi
32
· (dtmm)3
= 1257 mm3
(K.112)
At = pi ·
1
4
· d2
t = 1257 mm2
(K.113)
qt =
MtF EM
Wt
= −96 N/mm2
(K.114)
tts =
VA
At
= 65 N/mm2
(K.115)
ttt =
16 · Tt
π · d3
t
= 13 N/mm2
(K.116)
tttot = ttt + tts = 78 N/mm2
(K.117)
Qt1 = 3 · (tttot)2 = 135 N/mm2
(K.118)
Ut1 =
Qt1
Qtrunnionallowed
· 100% = 100% (K.119)
ttt2 =
16 · Tt2
π · d3
t
= 0 N/mm2
(K.120)
Qt2 = 3 · (tt2)2 = 0 N/mm2
(K.121)
Ut2 =
Qt2
Qtrunnionallowed
· 100% = 0% (K.122)
xliv

K.12 Allowed stress bearings
σy = 140 N/mm2
(K.123)
γ = 1.5 (K.124)
σbearingallowed
=
sigmay
γ
= 93.33 N/mm2
(K.125)
K.13 stress lower stem bearing
blsb = 50 mm (K.126)
hlsb = 30 mm (K.127)
Alsb = blsb · hlsb = 1500 mm2
(K.128)
qlsb = Fls/Alsb = 30 N/mm2
(K.129)
Ulsb =
qlsb
qtotb
· 100% = 32% (K.130)
K.14 stress upper stem bearing
busb = 50 mm (K.131)
husb = 30 mm (K.132)
Ausb = busb · husb = 390 mm2
(K.133)
qusb = Fus/Ausb = 81 N/mm2
(K.134)
Uusb =
qusb
qtotb
· 100% = 87% (K.135)
K.15 stress trunnion bearing
btb = 50 mm (K.136)
htb = 30 mm (K.137)
Atb = btb · htb = 880 mm2
(K.138)
qtb = Ft/Atb93 N/mm2
(K.139)
Utb =
qtb
qtotb
· 100% = 100% (K.140)
Lact =
Ttot
Fact
= 50 mm (K.141)
Lacta =
Ttot · γ
Fact
= 75 mm (K.142)
xlv

L Calculation method for distributed loads
This section shows the calculation method to ﬁnd the moment if the reaction forces from
the bearings was calculated as distributed loads. This will give a lower moment then for
point loads, and for that reason not used.
The lengths that the force is distributed on is trunnion, gasket and the lower and upper
bearing, denoted as Lt, Lg, Lls, Lus. For these lengths there will be a linearly inclining or
declining shear force. The lengths Lg−ls, Lls−us and Lus−act is the lengths in between the
loads, where there will be a constant shear force.
Section Length [mm]
Lt 20
Lg 65
Lg−ls 23.3
Lls 30
Lls−us 19.8
Lus 7.8
Lus−act 59.5
To ﬁnd the points where the declining shear force intercepts the neutral line, which there
is three of. The relationship A/a = B/b is used, where A and B is the two legs of the full
triangle, and a and b is the legs of the triangle above or below the neutral line, inside the
full triangle. The length from the end of the trunnion to the interception is denoted as
Lzero−g, And for the length from start of the lower bearing to the interception, and from
the start of the lower bearing to the interception are denoted as Lzero−ls and Lzero−us.
Lzero−g =
|VA| · Lg
|VA| + |VB|
(L.1)
Lrest−g = Lg − Lzero−g (L.2)
Lzero−ls =
|VB| · Lls
|VB| + |VC|
(L.3)
Lrest−ls = Lls − Lzero−ls (L.4)
Lzero−us =
|VC| · Lus
|VC| + |VD|
(L.5)
Lrest−us = Lus − Lzero−us (L.6)
xlvi

The moments at the diﬀerent lengths are then:
M1 = Mt−FEM +
VA · Lt
2
(L.7)
M2 = M1 +
VA · Lzero−g
2
(L.8)
M3 = M2 +
VB · Lrest−g
2
(L.9)
M4 = M3 + VB · Lg (L.10)
M5 = M4 + Mls−FEM + Lzero−ls · VB (L.11)
M6 = M5 + Lrest−ls · VC (L.12)
M7 = M6 + Lls−us · VC (L.13)
M8 = M7 + Mus−FEM + Lzero−us · VC (L.14)
M9 = M8 + Lrest−us · VD (L.15)
M10 = M9 + Lus−act · VD (L.16)
In addition to this, the moments at the interesting sections are as followed. Where Mh is
the hexagon, Mc is the lower circular, and Ms is the circular section under the seal/glide
ring. Mtc, Mlsc and Musc is the circular sections beneath the trunnion, lower stem bearing
and the upper stem bearing.
Mtc = M1 (L.17)
Mh = M3 + Lh · VB (L.18)
Mc = M3 + Lc · VB (L.19)
Mlsc = M5 (L.20)
Ms = M5 + Ls · VC (L.21)
Musc = M8 (L.22)
xlvii

Real simulation v8:1
As before
Study Properties
Study Type Static Stress
Last Modification Date 2016-04-13, 11:39:09
Constraints
Upper
Type Pinned
Radial Yes
Axial No
Tangential No
Lower
Type Pinned
Radial Yes
Axial No
Tangential No
Trunnion
Type Pinned
Radial Yes
Axial Yes
Tangential No
Loads
Force Ball
Type Force
X Value 0 N
Y Value 124689 N
Z Value 0 N
Radius 70.59 mm
Contacts
Bonded
M FEM report for the Bearing forces
xlviii

Name
Bonded9 [20367 - Ball:1||NV 200012 - Stem:1]
Separation + Sliding
Name
Separation + Sliding1 [NV 200012 - Stem:1||20894 - Bearing:1]
Separation + Sliding24 [20367 - Ball:1||NV 200014 - Bearing:1]
Results
Reaction Forces
Constraint Name
Reaction Force Reaction Moment
Magnitude Component (X,Y,Z) Magnitude Component (X,Y,Z)
Upper 3586 N
0 N
1155 N mm
-1154 N mm
-3586 N -37.36 N mm
6.436 N 0 N mm
Lower 48349 N
0 N
481827 N mm
-481827 N mm
-48349 N 0 N mm
0 N 0 N mm
Trunnion 72756 N
0 N
548649 N mm
548647 N mm
-72756 N 1420 N mm
0 N 0 N mm
xlix

Mount bend new v5:1
Bend new
Study Properties
Constraints
Stem
Type Fixed
Ux Yes
Uy Yes
Uz Yes
Loads
Actuator
Type Force
X Value -4.702E-11 N
Y Value -28000 N
Z Value 0 N
Radius 12.5 mm
Selected Entities
N FEM report for mount with bend
l

Results
Result Summary
Name Minimum Maximum
Safety Factor
Per Body 0.9125 15
Stress
Von Mises 0.02678 MPa 224.7 MPa
1st Principal -59.31 MPa 272.9 MPa
3rd Principal -219.4 MPa 36.52 MPa
Normal XX -216.6 MPa 193.1 MPa
Normal YY -186.9 MPa 199.9 MPa
Normal ZZ -70.52 MPa 68.65 MPa
Shear XY -56.48 MPa 119.8 MPa
Shear XZ -15.19 MPa 24.39 MPa
Shear YZ -31.02 MPa 25.54 MPa
Displacement
Total 0 mm 0.1161 mm
X -0.06764 mm 0.02008 mm
Y -0.105 mm 0.007412 mm
Z -0.01042 mm 0.004867 mm
Strain
Equivalent 1.163E-07 0.001035
1st Principal -1.975E-05 0.001213
3rd Principal -9.627E-04 5.666E-06
Normal XX -9.611E-04 8.728E-04
Normal YY -8.684E-04 8.72E-04
Normal ZZ -2.637E-04 1.726E-04
Shear XY -3.671E-04 7.788E-04
Shear XZ -9.872E-05 1.585E-04
Shear YZ -1.92E-04 1.581E-04
Contact Pressure
Total 0 MPa 55.52 MPa
X -32.43 MPa 10.53 MPa
Y -55.43 MPa 34.33 MPa
Z -23.39 MPa 17.68 MPa
li

Stress
Von Mises
[MPa] 0 224.7, Threshold: 0 - 130.6
Displacement
Total
[mm] 0 0.1161
lii

Negative Y + torsion Closed
Study Properties
Constraints
Upper
Type Pinned
Radial Yes
Axial No
Tangential No
Lower
Type Pinned
Radial Yes
Axial No
Tangential No
Trunnion
Type Pinned
Radial Yes
Axial Yes
Tangential No
Actuator
Type Pinned
Radial No
Axial No
Tangential Yes
Loads
Force Ball
Type Force
X Value 0 N
Y Value 124689 N
Z Value 0 N
Radius 70.59 mm
Force actuator
Type Force
X Value 0 N
Y Value -30000 N
Z Value 3.886E-12 N
Trunnion friction
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 163704 N mm
Gasket friction
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 1.146E+06 N mm
Lower bearing friction
Type Moment
X Value 0.1 N mm
Y Value 0 N mm
Z Value 99986 N mm
Glide ring
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 4875 N mm
Upper bearing friction
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 21329 N mm
Thrust bearing friction
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 60377 N mm
O FEM report for closed valve
liii

Contacts
Bonded
Name
Bonded11 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded12 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded13 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded14 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded15 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded16 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded17 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded18 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded19 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded20 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded21 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded23 [20367 - Ball:1||NV 200012 -
Stem:1]
Name
Separation + Sliding1 [NV 200012 -
Stem:1||20894 - Bearing:1]
Separation + Sliding24 [20367 - Ball:1||NV
200014 - Bearing:1]
liv

Results
Stress
Von Mises
[MPa] 0.3 887.2
Shear XY
[MPa] -180 166
Shear XZ
[MPa] -195.8 184.3
Shear YZ
[MPa] -399.6 236.4
lv

Displacement
Total
[mm] 0.0203 0.133
X
[mm] -0.1085 0.1088
Y
[mm] -0.1066 0.1257
Z
[mm] -0.00035 0.04357
lvi

Negative Y + torsion Open
Study Properties
Constraints
Upper
Type Pinned
Radial Yes
Axial No
Tangential No
Lower
Type Pinned
Radial Yes
Axial No
Tangential No
Trunnion
Type Pinned
Radial Yes
Axial Yes
Tangential No
Actuator
Type Pinned
Radial No
Axial No
Tangential Yes
Loads
Gasket friction
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 1.836E+06 N mm
Glide ring
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 4875 N mm
Thrust bearing friction
Type Moment
X Value 0 N mm
Y Value 0 N mm
Z Value 60377 N mm
P FEM report for open valve
lvii

Contacts
Bonded
Name
Bonded11 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded12 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded13 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded14 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded15 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded16 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded17 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded18 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded19 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded20 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded21 [20367 - Ball:1||NV 200012 -
Stem:1]
Bonded23 [20367 - Ball:1||NV 200012 -
Stem:1]
Name
Separation + Sliding24 [20367 - Ball:1||NV
200014 - Bearing:1]
lviii

Results
Stress
Von Mises
[MPa] 0 1319
Shear XY
[MPa] -341.1 219.9
Shear XZ
[MPa] -453.1 556.8
Shear YZ
[MPa] -592.7 725.4
lix

Displacement
Total
[mm] 0.0021 0.1724
X
[mm] -0.1723 0.1723
Y
[mm] -0.1724 0.1724
Z
[mm] -0.003128 0.000045
lx

Martin_Ness_Bachelor_Thesis

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