Environmentally acceptable lubricants (EAL) in the maritime industry
Christer Finnøy
Subsea Technology
Supervisor: Nuria Espallargas, IPM Submission date: June 2016
Preface
This Master Thesis is submitted to the Norwegian University of Science and Technology (NTNU) as a part of the Master’s degree program Subsea Technology at the Department of Engineering Design and Materials. The project has been a collaboration between NTNU and SINTEF, with Nuria Espallargas as the main supervisor together with Sergio Armada and Angelika Brink as co-supervisors.
Acknowledgement
I would like to thank my supervisor and co-supervisors for all help and guidance during the project period. I would also like to thank the industrial partner for valuable input.
Trondheim, 10.06.2016
Christer Finnøy
Abstract
All dynamic mechanical equipment needs lubrication to perform in a best possible way. In the maritime industry, e.g. ships, many of the most important dynamic mechanical systems are located beneath the surface of the sea. Examples of equipment are thrusters, stern tubes and ship stabilizers. All these components are dependent on proper lubrication to perform optimally and secure a long service-life. In the interface between the seawater moving parts, a seal is required to separate the oil and seawater from each other. Due to new requirements, conventional mineral oil shall be exchanged with an environmentally acceptable lubricant (EAL).
In this study different elastomers were tested against different steels (S355-steel and NiCr- steel) with two different oils (Shell Omala 68 and Kl¨uberbio EG 2-68). The motivation for performing the tests was to identify the tribological properties of the elastomers in the required EAL compared to the oil in use, and occurring mechanisms and the chemomechan- ical effects of the lubricants in interaction with the seal. The testing was performed at lab scale using a ”Pin-on-plate” configuration in a tribomachine. To investigate the effect of the lubricants on the elastomers two soaking tests were performed, and soaked samples were tribologically evaluated.
The result from this study showed that the conventional mineral oil give the lowest coefficient of friction (COF) when comparing the two oils. When comparing the elastomers, NBR showed a lower COF than FKM. It was concluded that all the elastomers are operating in a boundary/mixed lubrication regime. NBR exhibit higher wear compared to FKM in most situations, exceptions are when adding abrasive particles (Silicon Carbides) to the lubricant and the contact pressure (CP) is low. Transfer of polymer from adhesive wear towards the metal surface could be observed for all elastomers. The results from the soaking tests showed that EAL degrade NBR more than conventional oil and that higher temperatures accelerate the degradation. In the soaking test FKM gained mass from aging. This study gives pointers of expected behaviour in the sealing system.
Sammendrag
Alt dynamisk mekanisk utstyr trenger smøring for ˚a prestere best mulig. I maritim indus- tri, f.eks. skip, er mange av de viktigste dynamiske mekaniske systemene lokalisert under sjølinjen. Eksempler p˚a dette er sidepropellere, propellhylser og skips-stabilisatorer. Alle disse komponentene er avhengig av tilstrekkelig smøring for ˚a fungere optimalt og sikre en lang levetid. I grensesnittet mellom sjøvann og bevegelige deler er det nødvendig med en tetning for ˚a holde sjøvann og smøremiddelet avskilt. Grunnet nye krav, skal konvensjonell mineralolje byttes ut med et miljøvennlig smøremiddel.
I denne studien ble forskjellige elastomerer testet mot forskjellige st˚al (S355-st˚al og NiCr- st˚al) med to forskjellige oljer (Shell Omala 68 and Kl¨uberbio EG 2-68). Motivasjonen for testingen var ˚a identifisere de tribologiske egenskapene til elastomerene i det miljøvennlige smøremiddelet sammenlignet med dagens olje, og karakterisere hvilke mekanismer som opp- stod og studere smøreoljenes kjemi-mekaniske effekter. Testingen ble utført p˚a lab-niv˚a ved bruk av en ”Pinne-p˚a-plate” konfigurasjon i en tribo-maskin. For ˚a undersøke effekten av smøremiddelet p˚a elastomerene ble en neddykket test ogs˚a utført, i tillegg ble de nedykkede elastomerende friksjons-testet.
Resultatet fra denne studien viste at mineralolje ga den laveste friksjonskoeffisienten (COF) hvis man sammenligner de to oljene. Sammenligning av elastomerene viste at NBR gav en lavere COF enn FKM. Det ble konkludert med at begge elastomerene opererte i ”grensesjik- tet” og ”blandasjiktet”. NBR ble slitt mer sammenlignet med FKM i de fleste situasjoner, unntaket var n˚ar det var abrasive partikler i kontaktpunktet og n˚ar den ble testet p˚a la- vere kontakttrykk (CP). Overføring av polymer fra adhesiv slitasje mot metalloverflaten kunne observeres for alle elastomerene. Resultatet fra de neddykkede testene viste at EAL degraderer NBR mer enn konvensjonell mineralolje og at en økning i temperaturen vil ak- selerere degraderingen. Den samme testen viste at NBR blir utsatt for nedbrytning, og at en høy temperatur av smøremidlet vil akselerere denne prosessen. I de neddykkede testene økte FKM i vekt p˚a grunn av eksponeringen og effekten var størst i konvensjonell min- eralolje ved høyere temperaturer. Denne studien gir pekepinner om forventet oppførsel i tetningene.
Contents
Preface i
Abstract iii
Sammendrag v
1 Introduction 1
1.1 Aim of work . . . 3
2 Theory 5 2.1 Tribology . . . 5
2.1.1 Friction . . . 5
2.1.2 Laws of sliding friction . . . 6
2.1.3 Theory of elastomer friction . . . 6
2.1.4 Lubricated contact . . . 8
2.1.5 Wear . . . 9
2.1.6 Wear of elastomers . . . 10
2.1.7 Classification of metallic wear . . . 12
2.2 Sealing systems . . . 12
2.2.1 Elastomers used in sealing systems . . . 13
2.2.2 Acrylonitrile-butadiene rubber . . . 14
2.2.3 Fluorocarbon rubber . . . 14
2.2.4 Lip seals . . . 15
2.2.5 Sealing surface . . . 16
2.2.6 Surface roughness and real contact area . . . 16
2.3 Lubricants and lubrication . . . 17
2.3.1 Lubrication regimes . . . 17
2.3.2 Mineral oil . . . 19
2.3.3 Environmentally Acceptable Lubrication . . . 19
2.3.4 Different types of environmentally acceptable lubricants . . . 20
2.3.5 Aging in lubricant . . . 22
2.3.7 Water in the lubricant . . . 23
2.3.8 Viscosity . . . 24
2.3.9 Viscosity index . . . 24
3 Experimental setup 25 3.1 Sample preparation . . . 25
3.1.1 Metal samples . . . 25
3.1.2 Elastomer pins . . . 26
3.1.3 Environmentally Acceptable Lubricant . . . 26
3.1.4 Conventional Mineral Oil . . . 27
3.1.5 Contamination . . . 27
3.2 Pin on plate . . . 29
3.3 Soaking-tests . . . 32
4 Results 33 4.1 Elastomers . . . 33
4.1.1 Alternative elastomers . . . 38
4.2 Different materials as counter surface . . . 40
4.3 Different contact pressure . . . 43
4.3.1 Friction Force . . . 46
4.4 Lubricants . . . 46
4.4.1 Mixed lubricant . . . 47
4.4.2 Lubricant contaminated with water . . . 48
4.4.3 Lubricant contaminated with SiC . . . 51
4.4.4 No lubricant . . . 55
4.5 Wear of elastomer in friction tests . . . 58
4.6 Friction tests with aged elastomers . . . 63
4.7 Soaking tests . . . 66
4.8 Hydrolysis of Kluber . . . 68
5 Discussion 71 5.1 Influence of elastomer . . . 71
5.2 Influence of counter-surface . . . 73
5.3 Influence of contact pressure . . . 74
5.4 Influence of lubricant . . . 76
5.5 Influence of water contamination . . . 76
5.5.1 Hydrolysis of ester oil . . . 77
5.6 Influence of no lubricant . . . 77
5.7 Influence of aging . . . 78
5.8 Influence of SiC particles . . . 79
5.9 Wear of metal . . . 80
5.10 Experimental approach versus real application . . . 82
6 Conclusions 85
7 Further work 87
Appendix A Appendix B Appendix C
Abbreviatons
EAL Environmental acceptable lubricant VGP Vessel General Permit
NBR Nitrile-buatadiene rubber FKM Fluorocarbon rubber NBR-1 72 NBR 902
FKM-1 75 FKM 585 Shell Shell Omala S2 G68 Kluber Kl¨uberbio EG 2-68 Sand Sika WS - F800 ACN Acryonitrile
HRC Hardness Rockwell, Scale C CP Contact pressure
COF Coefficient of friction HD Hydrodynamic
EHD Elastohydrodynamic
EHL Elastohydrodynamic lubrication PAG Polyakylene glycols
VI Viscosity index
SEM Scanning electron microscope
EDS Energy-dispersive X-ray spectroscopy PPM Parts per million
SiC Silicon Carbide Approx. Approximately
List of Figures
1 Different abrasive mechanisms. [28] . . . 10
2 Lip seal functionality with pressure profile . . . 15
3 Stribeck curve with lubrication regimes . . . 18
4 Example of an environmentally acceptable lubricant . . . 21
5 Two fluids with different VI . . . 24
6 Picture of metal sample used in tests . . . 25
7 Picture of a pin used in the tests . . . 25
8 Pin on plate . . . 29
9 The TE88 . . . 29
10 Trace/retrace . . . 30
11 Test-setups for friction tests . . . 31
12 NBR-1 and FKM-1 in Shell . . . 33
13 NBR-1 and FKM-1 in Kluber . . . 33
14 NBR-1 tested in Kluber . . . 34
15 FKM-1 tested in Kluber . . . 34
16 NBR-1 and FKM-1 in 75% Shell and 25% Kluber . . . 34
17 NBR-1 and FKM-1 in 50% Shell and 50% Kluber . . . 34
18 NBR-1 and FKM-1 in 25% Shell and 75% Kluber . . . 35
19 NBR-1 and FKM-1 in 75% Shell and 25% water . . . 36
20 NBR-1 and FKM-1 in 75% Kluber and 25% water . . . 36
21 NBR-1 and FKM-1 in 50% Shell and 50% water . . . 36
22 NBR-1 and FKM-1 in 50% Kluber and 50% water . . . 36
23 NBR-1 and FKM-1 in 25% Shell and 75% water . . . 37
24 NBR-1 and FKM-1 in 25% Kluber and 75% water . . . 37
25 SEM-image outside the wear track of S355 tested with NBR-1 and Shell. . . 37
26 SEM-image inside wear track of S355 tested with NBR-1 and Shell. . . 37
27 SEM-image outside the wear track of S355 tested with FKM-1 and Shell. . . 38
28 SEM-image inside wear track of S355 tested with FKM-1 and Shell. . . 38
29 NBR-1 and NBR-2 compared in Shell . . . 39
30 NBR-1 and NBR-2 compared in Kluber . . . 39
31 FKM-1 and FKM-2 compared in Shell . . . 39
32 FKM-1 and FKM-2 compared in Kluber . . . 39
33 Point analysis of S355 with FKM-2 . . . 40
34 Point analysis of S355 with NBR-2 . . . 40
35 NBR-1 tested on different steels with 1 MPa nominal CP . . . 41
36 FKM-1 tested on different steels with 1 MPa nominal CP . . . 41
37 NBR-1 tested on different steels with 2 MPa nominal CP . . . 41
38 FKM-1 tested on different steels with 2 MPa nominal CP . . . 41
39 Grooves at surface at S355 . . . 42
40 Grooves at surface of NiCr . . . 42
41 SEM-image outside the wear track of S355 tested with NBR-1 and Shell. . . 43
44 SEM-image inside the wear track of NiCr tested with NBR-1 and Shell. . . . 43
45 NBR-1 in Shell, different nominal CP . . . 44
46 NBR-1 in Kluber, different nominal CP . . . 44
47 FKM-1 in Shell, different nominal CP . . . 45
48 FKM-1 in Kluber, different nominal CP . . . 45
49 NBR-1 in Kluber on NiCr, different nominal CP . . . 45
50 FKM-1 in Kluber on NiCr, different nominal CP . . . 45
51 SEM-image inside the wear track of NiCr tested at 1 MPa. Direction of sliding was along the diagonal between top-left corner and bottom-right corner. . . . 46
52 SEM-image inside wear track of NiCr tested at 2 MPa. Direction of sliding was along the diagonal between top-left corner and bottom-right corner. . . . 46
53 Friction force for different nominal CP . . . 47
54 Friction force for different nominal CP . . . 47
55 NBR-1 in Shell and Kluber mixtures . . . 48
56 FKM-1 in Shell and Kluber mixtures . . . 48
57 NBR-1 in Shell and water mixtures . . . 49
58 FKM-1 in Shell and water mixtures . . . 49
59 NBR-1 in Kluber and water mixtures . . . 50
60 FKM-1 in Kluber and water mixtures . . . 50
61 New FKM-1 pin . . . 50
62 FKM-1 tested in 75% Shell and 25% water . . . 50
63 FKM-1 tested in 50% Shell and 50% water . . . 50
64 FKM-1 tested in 25% Shell and 75% water . . . 50
65 NBR-1 in SiC contaminated Shell . . . 51
66 NBR-1 in SiC contaminated Kluber . . . 51
67 FKM-1 in SiC contaminated Shell . . . 52
68 FKM-1 in SiC contaminated Kluber . . . 52
69 NBR-1 in SiC contaminated Shell and Kluber . . . 52
70 FKM-1 in SiC contaminated Shell and Kluber . . . 52
71 NBR-1 and FKM-1 in SiC contaminated Shell . . . 53
72 NBR-1 and FKM-1 in SiC contaminated Kluber . . . 53
73 NBR-1 and FKM-1 in Shell and Kluber contaminated with SiC . . . 53
74 3D-image of NBR-1 pin after tested in SiC contaminated Shell . . . 54
75 3D-image of NBR-1 pin after tested in SiC contaminated Kluber . . . 54
76 3D-image of FKM-1 pin after tested in SiC contaminated Shell . . . 54
77 3D-image of FKM-1 pin after tested in SiC contaminated Kluber . . . 54
78 SEM-image in the wear track after tested with NBR-1 in Shell (without SiC). 55 79 SEM-image in the wear track after tested with NBR-1 in SiC contaminated Shell. . . 55
80 SEM-image in the wear track after tested with FKM-1 in Shell (without SiC). 55 81 SEM-image in the wear track after tested with FKM-1 in SiC contaminated Shell. . . 55
82 NBR-1 in dry contact . . . 56
83 FKM-1 in dry contact . . . 56
84 Macroscopic-picture of the wear track after test of FKM-1 in dry contact . . 56
85 IFM-image of the wear track after test of NBR-1 in dry contact . . . 56
86 IFM-picture of NBR-1 tested in dry contact . . . 57
87 IFM-picture of FKM-1 tested in dry contact . . . 57
88 SEM-image outside the wear track. . . 57
89 SEM-image of the wear track for S355 with NBR-1 tested under dry condition. 57 90 Mass loss in the basis tests . . . 58
91 Mass loss in different oil mixtures . . . 59
92 Mass loss in different oil and water mixtures . . . 60
93 Mass loss in the oil contaminated by SiC-sand . . . 61
94 Mass loss at NiCr . . . 62
95 Wear in the dry conditions . . . 62
96 NBR-1 in Shell . . . 63
97 NBR-1 in Kluber . . . 63
98 FKM-1 in Shell . . . 64
99 FKM-1 in Kluber . . . 64
100 NBR-1 aged at 60◦C in Shell . . . 64
101 NBR-1 aged at 60◦C in Kluber . . . 64
102 FKM-1 aged at RT in Shell . . . 65
103 FKM-1 aged at RT in Kluber . . . 65
104 Plate-track after tested with NBR-1 aged at 60◦C in Shell. Sliding direction was along left to right. . . 65
105 Plate-track after tested with NBR-1 aged at 60◦C in Kluber. Sliding direction was along left to right. . . 65
106 Plate-track after tested with FKM-1 aged at 60◦C in Shell. Sliding direction was along left to right. . . 65
107 Plate-track after tested with FKM-1 aged at 60◦C in Kluber. Sliding direction was along left to right. . . 65
108 NBR-1 aged at RT . . . 66
109 FKM-1 aged at RT . . . 66
110 NBR-1 aged at 60◦C . . . 67
111 FKM-1 aged at 60◦C . . . 67
112 NBR-1 aged at RT . . . 68
113 FKM-1 aged at RT . . . 68
114 NBR-1 aged at 60◦C . . . 68
115 FKM-1 aged at 60◦C . . . 68
116 Typical surface of elastomer after test . . . 73
117 Illustration of probable wear-rate . . . 75
118 Wear-track in real application . . . 81
119 Temperature dependent COF . . . 82
120 Load-profile over wear track . . . 83
List of Tables
1 Mild and severe wear, gathered from [31] . . . 12
2 Most common used elastomers . . . 13
3 Roughnessdata from Simrit Technical Manual 2007 . . . 16
4 Physical properties, gathered from data-sheet of the elastomers . . . 26
5 Acronym of the lubricant mixtures . . . 28
6 pH-values for different mediums . . . 69 7 Test-matrix of soaking-test . . . . 8 Test-matrix of second soaking-test . . . .
1 Introduction
Due to new and stricter regulations, the use of Environmentally Acceptable Lubricants (EAL) becomes more and more relevant in Maritime and Offshore industries to replace more pollutant lubricants used today. The main limitation of EAL that avoids its deployment in the maritime is that EALs are still not proven as good as conventional lubricants. One issue regarding the EAL is the probability of increased wear, with the economic consequences, such as; Replacing parts, expenses due to downtime of the machine, low efficiency and increased energy consumption.
The 2013 Vessel General Permit (VGP) states that ”All vessels must use an EAL in all oil to sea interfaces, unless technically infeasible” [2]. This means that if a vessel longer than 79 feet, wants to enter the coastline of the United States (US), the vessel shall have EAL in all oil-to-sea interfaces. An oil-to-sea interface is a general term for equipment that is subjected to immersion in seawater, included stern tubes, stabilizers and thrusters. [2]
The background for the new regulations are among other factors the environment. While the number of ships sailing on the ocean is increasing, the pollution from each ship is still the same and therefore the pollution of the ocean is growing. As an attempt to downsize this pollution, drastic action needs to be implemented. According to an analysis of Etkin D.S. [8] the amount of vessels operational discharges covered 51 percent of all oil input to the marine environment. This is by far the biggest input compared to accidents, ship spills and so on. This leakage is considered as a part of normal consumption of the oil. With this large leakage to the sea during ”normal operations” the regulations are longed for.
The industrial partner in this project and thesis, is a world leading manufacturer of a me- chanical product. Due to confidentiality the product and name for the manufacturer can not be revealed. After the new VGP in 2013 there has been an increasing demand for a solution that could use EAL. Todays solution uses a lubricant-filled product. Everything inside the product is immersed in mineral oil, included a gearbox. The oil in the product have several tasks, cooling, lubricating, corrosive protection and act as a barrier to seawa- ter among others. A mineral oil obtains several of the required demands in the product.
conventional oil used in the product. Some concerns when changing the oil are the sealing system and the gearbox. The gearbox needs an oil that can lubricate the high load contact in the gear wheels. The problems concerning the sealing system are more complex. Today the sealing system consists of a elastomer seal and a metallic counter sealing surface. Both components may be changed easily and this is often done when the product is scheduled for service. There is no exact knowledge of how often the seal or the steel counter sealing surface need to be changed, and it is decided by the serviceman when it is necessary to do so. So far, there are several proposals how the industrial partner will solve this. One solution is to change all the oil to EAL, and change the components that does not handle the new oil e.g.
the seals. Another solution is to divide the product into several separate rooms and fill the gearbox with mineral oil and the other with EAL. This way the gear box will be lubricated by mineral oil which is a self proven good solution, and the EAL will lubricate the sealing systems and leak into the mineral oil and the sea. This solution requires a new design of the product and a way to separate the oils so they don’t mix. The pros with replacing the existing mineral oil with EAL in the original design are that there is no need for re-design and that there is no need to separate two oils. The con is that this will lead to uncertainties of the wear and protection of the gear wheels.
The used sealing system today is a nitrile rubber (NBR) - lip seal against a mild-steel counter surface lubricated with mineral oil. This solution has been used for decades and is a self proven good solution. There are still uncertainties how this sealing system will react with a new type lubricant. The supplier of lip seals has recommended to switch from NBR to a fluorocarbon (FKM) lip seal, which is a more robust material in harsh environments. The counter sealing surface also show signs of wear when operated for a while. The wear is believed to be higher with a FKM lip seal, so therefore it is discussed if the counter surface metal should also be changed. It would be of great value to know the limitations and lifetime of the components in the sealing-system before installing it in the product. Even though the lubricant is an EAL, it is desired that the system do not leak.
When the product is operating there are many factors which can influence the tribological performance of the seal e.g. temperature, velocity in the interface, type of lubricant, contact pressure, sealing counterface, operating hours and elastomer type.
1.1 Aim of work
The aim of this Master Thesis will be to investigate the tribological properties of different elastomers and two lubricants. One of the lubricants will be a conventional mineral oil, the other an EAL. Research within elastomers, environmentally acceptable lubricants, wear mechanisms and general tribology must be performed.
Simplified tests will be performed at NTNU’s Tribology-lab. The first set of tests will be to create a basis to compare other tests with, where the elastomers and lubricants will be tested at the original steel used in the application. A contact pressure, temperature and speed will be chosen as reference values. When this basis is established, the factors will be changed individually to investigate the influence they may achieve.
Some constraints appeared in this project. It was planned to perform tribological tests in a test rig with the industrial partner. This was not performed due to limited access to the test rig. Stribeck-curves for the different interactions were not investigated due to a broken
”Pin-on-disc”-machine.
The motivation to perform this project was to investigate if the EAL was compatible in the application for the industrial partner and what physical changes have to be done to make it optimal. A full investigation of the problem could possible take years and therefore it was necessary to narrow the problems down and investigate selected conditions that may occur in the system.
2 Theory
2.1 Tribology
Tribology is the science and technology of interacting surfaces in relative motion and of re- lated subjects and practices. The mechanisms occurring on the surface regarding tribology are highly complex, and to fully understand it, knowledge of various disciplines are acquired, such as; Physics, chemistry, applied mathematics, solid and fluid mechanics, thermodynam- ics, heat transfer, materials science, and others. Tribology is divided into macro tribology and micro/nano tribology. In macro tribology, tests are performed with a relative large mass and high normal forces. Under these conditions, wear is inevitable and the tribological properties are dominated by the properties of the mating test objects. Micro/nano tribology was developed because when two surfaces are forced against each other, contact between them occurs at numerous asperities, and there is of great interest to investigate the contact between single asperities. Testing of micro/nano tribology is performed by measuring the performance of at least one component in a system existing of two components mated against each other with low mass and low normal forces. This test will result in negligible wear and the tribological performance will be dominated by the surface properties. [4]
2.1.1 Friction
The friction force always act against the sliding and reduces the performance of the system.
Friction can be divided into dry friction and fluid friction. The dry friction is the contact force that occurs when two dry surfaces moving relative to each other. Fluid friction is the contact force that occurs adjacent to the fluid layer that is between two surfaces moving relative to each other. The friction can also be divided into static and dynamic friction. The static friction force, Fs is the force required to initiate motion between two bodies mated with a nominal force. The dynamic friction force (also called kinematic friction force), Fk is the required force to maintain the motion between two bodies with a nominal load.
The friction is a system response and not a material property. Friction depends among others
that is needed to overcome the friction force will be partly dissipated as plastic deformation, waste heat and wear. In most systems a tribologist will try to keep the friction low to save the system from damage and wear. Although, in some systems it is desirable having a high friction force with low wear rates e.g. automobile tyres, clutches and vehicle brakes. [31, 4]
2.1.2 Laws of sliding friction
There are three laws of sliding friction that are most commonly used. The first and second are often referred to as Amontons’ equations and the third often referred to as Coulomb law.
The three laws state the following:
• F=µW. The friction F is directly proportioned to the normal force W. µis the coeffi- cient of friction (COF).
• The COF is independent on the apparent area of contact between the contacting bodies.
• The COF is independent on the sliding velocity once motion start.
[4]
2.1.3 Theory of elastomer friction
There are two commonly described contributions to the friction force between rubber and a hard and rough surface, adhesion and the hysteresis components. The hysteresis component is also commonly known as deformation. [21]
A large fraction of the elastomer friction is due to a large energy dissipation occurring in the rubber bulk material, this is the hysteresis component. This occurs when the surface stresses created between the hard asperities of the metal and the softer elastomer, deform the elastomer surface and the energy created travel into the elastomer bulk material. Moore [21] describe that hysteresis is a bulk phenomenon dependent on the viscoelastic properties of the elastomer. The adhesion component on the other hand is occurring on the surface.
[24, 22]
The adhesion is often describes as a ”thermally activated molecular stick-slip process” [22].
The chains in the elastomer structure are flexible and in a constant state of thermal motion.
When the elastomer moves over the hard surface, the flexible chains in the elastomer wants to attach to the molecules on the hard surface’s asperities. Some of these chains links to the hard surface for a moment before this connection fails at the weakest point. This is called the stick-slip process and is responsible for adhesion. [22]
The visco-elastic properties in the elastomer has a relation to both adhesion and hysteresis.
By defining a complex E* as the ratio of stress to strain for a visco-elastic body, the theory of visco-elastic define:
E∗ =E0+jE00
E0 is the storage moduls or stress-strain ratio for the component of strain in phase with the applied stress, E00 is the loss modulus or stress-strain ratio for the component of strain 90◦ out of phase of applied stress. The tangent modulus is defined as the ratio of energy dissipated to energy stored per cycle:
tan δ= E00 E0
For an elastomer the adhesional and hysteresis component of the COF is given by:
fadhesion=K2(E0 Pr)tan δ
fhysteresis =K3(P
E0)ntan δ
K2 and K3 are constants dependent of particular sliding combinations, P is the applied contact pressure and r is an exponent with a value around 0.2. Through this it is shown that both adhesion and hysteresis depends on the viscoelastic mechanisms.
in a self-making contact or against metal or ceramics. The ratio E/H (E is the Young’s modulus and H is the hardness) determines the extent of plasticity in the contact area. For most metals sliding against each other, this ratio is 100 or greater, while in soft polymers often around 10. The elastomer therefor almost completely deforms elastic when there is metal against polytetrafluoroethylene, nylons and rubbers.
Elastomers sliding against rigid and hard counterfaces often transfer a film of polymer over to the counterface. The formation and behaviour of these transfer films are important factors in the friction and wear of these elastomers. Most of the elastomers creates this transfer film by a ”lumpy transfer” process. The mechanism of lumpy transfer is that lumps of elastomer are removed by abrasion or adhesion from the elastomer and adhere to the counterface. The film transfer will continue on further sliding, adding more material to the transfer film since the interfacial bond to the counterface is often stronger than the elastomer itself. The transfer film will also wear through generation of wear particles adhered to the counter surface in addition to wear the elastomer. The elastomer and counter surface will gradually reach a steady interface with low friction and wear when the transfer layer is created. If the COF between an elastomer and metal is recorded over a wear track, a sudden peak in the graph is the sign of a lumpy film transfer. [12, 27]
2.1.4 Lubricated contact
Using a lubricant between two rough surfaces in relative motion will decrease the adhesion between the surfaces, and the main contribution to friction will be due to deformation (hysteresis) of the elastomer. [21]
In a lubricated contact the friction depends on several factors. The geometry, surface to- pography, velocity, cleanliness of the surface, temperature and the material and surface of the elastomer are among others important factors influencing the friction. Predicting the friction is challenging due to the complexity of the physical processes involved, especially in elastomer. There are several theories how the friction in lubricated contacts are achieved.
One theory is that the friction is mainly developed by resistance between the interactions of solids and the liquid, the resistance is achieved by shearing a thin layer of lubricant separat- ing the surfaces. If so, the friction is governed by the shear strength of the lubricant. When
operating the contact in a thin lubricant layer, which elastomer seals do, the thin layer can allow the solids asperities to come in contact and adhere.
Recently experimental work (2015) by Goda T.J. [10] indicates that NBR sliding on a lu- bricated smooth surface is dominated by hysteresis. It was also shown that the COF was decreasing with increasing temperatures in the lubricant. This was explained to be due to decreasing internal friction in the elastomer with increasing temperature. [10]
Persson [24] describes that experiments performed with elastomers on glass and silicon car- bide paper gives a friction coefficient with the same temperature dependency as the complex elastic modulus of elastomer. This prove that the friction force mainly is a bulk property of the rubber. In the study it is also shown that at low velocity the rubber will completely deform into the surface roughness profile. In another study Persson [23] describes ”Surface roughness of different length scale contribute equally to the friction force if the ratio between the amplitude and wavelength is constant”. This means, that if enough pressure is applied to the elastomer it will deform and be squeezed into a complete contact with the substrate.
The magnitude of the hysteresis contribution to the COF will depend on the ratio between wavelength and amplitude of the surface roughness. [24, 23]
2.1.5 Wear
Wear is the damage and material loss at the surface of a component due to friction. Most systems are lubricated to reduce wear that would occur in a dry contact. The wear rate ω is defined as the volume lost over time, and can be denoted therefore ms3. The wear rate for an un-lubricated surface is relative to speed, normal load, temperature, and the thermal/mechanical/chemical properties. Wear is categorized by different mechanisms such as; adhesive, abrasive, fatigue, erosion, chemical and electro arcinduced.[31, 4]
Adhesive wear Adhesion occurs at the contact in at the interface, which are sheared by sliding This may result in the detachment of a fragment from one of the surfaces and lead to an attachment onto the other surface. [4]
Abrasive wear ”Abrasive wear occurs when asperities of a rough, hard surface or hard particles slide on a softer surface and damage the interface by plastic deformation or frac- ture” [4], This mechanisms is also called ploughing, since the hard metal ”plough” into the softer material, and leaves a characteristic wear track. Abrasive wear can be divided into subcategories of mechanisms. The different mechanisms are cutting, fracture, fatigue by re- peated ploughing and grain pull-out. Cutting is the classic model when a soft surface is cut by either a hard asperity or sharp grit as seen in fig: 1 (a). If the abraded material is brittle, the wear will result in cracking of the surface. Example for this are ceramics, illustrated in fig: 1 (b). If the material is ductile and the grit is blunt, then cutting is unlikely and the result will be that the surface will be deformed as shown in fig: 1 (c). The last mech- anism is the “grain pull-out” which will mainly occur in ceramics with a weak intergranual boundary. If there are particles in the contact, either lubricated or dry, two-body abrasion or three-body abrasion may occur. With two-body abrasion the particles will embed into one of the surfaces and scratch against the other one, resulting in ploughing. With three-body abrasion, the particles does not embed and will move, roll and slide in the contact. [28, 12]
Figure 1: Different abrasive mechanisms. [28]
2.1.6 Wear of elastomers
Seals who are designed to run in a lubricated contact operate the majority of its lifetime with a lubricant. However, sometimes it has to be operated when there is no lubricant in the seal. This can be when starting the machine or in a system where there is redundancy with two seals in a row. With a double-barrier system (two seals) the outer seal can experience oil
starvation. Therefore, the friction and wear in dry conditions may be crucial for the sealing system. [17]
There are several factors influencing the wear, such as: pressure, sliding velocity, coefficient of friction, surface texture, elastic modulus, strength and fatigue resistance of the elastomer.
Wear of elastomers are due to tearing - local mechanical rapture, and smearing - decom- position of the molecular network to a lower molecular weight. Tearing of elastomer on a smooth surface can be due to either fatigue or frictional wear. These mechanisms can also be divided into adhesive, abrasive and fatigue wear. As smoother the surfaces are, more ad- hesive wear occurs compared to abrasive. As stated earlier, many of the elastomers creates film transfer due to abrasion and adhesion, this is also a wear mechanism. Initially when an elastomer runs dry against a hard surface repeatedly, a decrease in friction due to the transfer film formation until the layer detaches may occur until a complete layer of material is deposited from the elastomer to the hard surface [17]. Abrasive wear will occur when there is a sharp texture in the base surface tearing up the sliding elastomer and resulting in micro-cutting and longitudinal scratches can be observed on the elastomer. Fatigue wear occur when the texture of the base surface has projections that are blunt rather than sharp.
The surface of the elastomer will undergo cyclic deformation and eventually fail as a result of the fatigue. Fatigue wear occurs in most sliding operations, and requires a relative low coefficient of friction. Elastomers which have a low abrasion resistance have a relative high COF, which support the idea that an increased friction will increase the wear. With fatigue wear it will have pitting marks in the elastomer surface. The strength of the elastomer has a large influence on the wear resistance. For each elastomer there is a critical value of shear stress. If the shear stress exceed this value, abrasion wear will occur, and with a value below the critical the wear will mainly be due to fatigue. [17] [21]
When sliding against rough surfaces the abrasive mechanism may be the dominant wear.
Generally, elastomers have a high tolerance against abrasive particles, since the relative soft bulk-material can embed the particles easy. This can in opposite lead to an increase of wear of the hard phase counter metal. Many of the elastomers react with different fluid, and may swell and weaken their mechanical properties. [4]
2.1.7 Classification of metallic wear
The simplest way to classify metallic wear is either mild or severe wear. This classification does not say anything about the given wear mechanism, but is based on visual inspection and can be summarized in table 1. [31]
Mild wear Severe wear
Results in extremely smooth sur- faces - often smoother than origi- nal
Results in rough, deeply torn sur- faces - much rougher than the original
Debris extremely small, typically only 100 nm diameter
Large metallic wear debris, typi- cally up to 0.01 mm diameter High electrical contact resistance,
little true metallic contact
Low electrical contact resistance, true metal junctions formed Table 1: Mild and severe wear, gathered from [31]
2.2 Sealing systems
Almost always when there is a bearing function present, some kind of sealing is used. The sealing will reduce the loss of lubricant and also deny entry of contamination in form of particles and other liquids. There are two categories of seals; Static and dynamic seals.
Static seals are designed to produce leak-tight joint between two static surfaces. Dynamic seals are designed to create a sealing joint between a static surface and a surface in relative motion. Most commonly seals are the ”O-ring”, ”U-ring” and the lip seal. The O-ring is the most widely used seal in static situations. The o-ring is low in cost, reliable, easy to install and re-install. For all oil-to-sea interfaces there is need for a sealing to prevent seawater to enter the system and to keep the lubricant inside. The lip-seal is a common used sealing- solution when a rotating shaft is involved. The lip-seal creates a barrier between two sides to avoid to have contamination into the system. This is done by creating a pressure difference over the seal. The pressure in the system should be higher than the environment outside.
[31]
2.2.1 Elastomers used in sealing systems
Elastomers are categorized on the material they are based on. Table 2 shows the most common elastomers used in seals.
Chemical description Abbreviation
Acrylonitrile-butadiene rubber NBR Hydrogeneated acrylonitrile-butadiene rubber HNBR
Fluorocarbon rubber FKM
Perfluoroelastomer FFKM
Table 2: Most common used elastomers
Each category of elastomer can further be divided by their fillers, softeners (plasticizers), processing aids, curing agents, accelerators and other additives they contain.[30]
The basic polymer alone is not very useful. In order to be functional it has to be cured. This involve creation of cross-links between the polymer molecules. A small quantity of curing agent is mixed to the polymer and is subjected to heat and pressure to create a chemical bond between the polymer molecules.[9]
The reason why the use of elastomers in seals are so popular, are because they have some unique properties:
• Since the elastomers have such a low modulus of elasticity (E) as down to 5 MPa, compared with typical engineering metals which have a modulus of elasticity between 50−200 GPa, it will require a relative low stress to deflect. Also the elongation to break is normally above 100%. Seals are often designed to have a design strain between 10- 30%. Due to the elongation to break there is little chance of damaging the seal when installing it.
• Elastomers provide a high degree of resilience with low hysteresis. This provides the elastomer the ability to respond rapidly to pressure changes and vibrations.
• Low creep. The tensile strength of the material is relatively low, but when strained within the material properties they do not suffer excessive creep.
• Elastomers have a high Poisson Ratio, very close to 0.5. This means that the material has a increased resistance against in-compressibility. With a low elastic modulus and a high Poisson Ratio, this gives a material that is easy to deformed and incompressible.
[9]
2.2.2 Acrylonitrile-butadiene rubber
Acrylonitrile-butadiene rubber (NBR) is a popular sealing material because of the general good mechanical properties and its relatively low cost. Properties of NBR regarding chemical and physical resistance are determined by the acrylonitrile (ACN) content which can vary between 18% and 50%. At low ACN-values the elastomer has good flexibility at low temper- atures, but has a low resistance against oils and fluids. At higher ACN-values the flexibility at low temperature is reduced but the resistance against oil and fluid is improved. NBR is typically resistant against mineral oil-base lubricants and typically not resistant against chlorinated hydrocarbons and glycol-based brake fluids. [30]
2.2.3 Fluorocarbon rubber
Fluorocarbon rubber (FKM) material is known for its high resistance to heat and a wide variety of chemicals. The standard FKM materials have excellent resistance to mineral oils, greases, chlorinated hydrocarbons and many organic solvents and chemicals. The differ- ent FKM types can be categorized into five main types. They are categorized for fluorine percentage and what they consists of. The main types is listed up below.[14]
• A Type Di - polymer, it has a fluorine content of 65-66% and consists of vinylidene fluoride and hexafluoropropylene. This is the most common used FKM and it is highly resistant to oils, steam and acids.
• B Type Ter-polymer, has a fluorine content of 67-68% and consists of vinylidene fluo- ride, hexafluoropropylene and tetrafluoroethylene. This elastomer has a slightly better resistance against fluids than Type A.
• F Type Ter-polymer, fluorine content of 69% and consists of vinylidene fluoride, hex- afluoropropylene and tetrafluoroethylene. This has the best fluid resistance but is not ideal for low temperature use.
• GF Type Ter-polymer, fluorine content of 70% and consists of vinylidene fluoride, hexafluoropropylene, tetrafluoroethylene and a cure site monomer. This type has extra resistance to fluids, steam and acids. However still not ideal for low temperature uses.
• GFLT Type Ter-polymer, fluorine content of 67% and consists of vinylidene fluoride, hexafluoropropylene, tetrafluoroethylene and a cure site monomer. This elastomer has the lowest swell rate and best performance for low temperature uses. LT in GFLT stands for low temperature.
2.2.4 Lip seals
The sealing effect is achieved by elastomer that connects the shaft geometry to the housing.
This is done by a radial force created by a garter spring (see figure 2) and an internal overpressure. For a lip seal to perform optimally, there shall be a very thin film of lubricant, often no more than a micron in thickness, formed at the interface between sealing and counter-surface. This is a result of both hydrodynamic and surface tension effects. The lubricant-film must be thick enough to physically separate the surfaces, and at the same time be as small as possible to prevent unacceptable volumes of leakage. A brand new lip seal will have a sharp trimmed or moulded lip. During the first hours of running the lip beds in on the shaft creating a contact zone that is typically 0.2 – 0.3 mm wide. This number will grow after operational wear. In fig: 2 a typical lip seal is shown with the entrained pressure profile developed on the shaft. [9]
When choosing an elastomer for a sealing system, the temperature has a major influence in the choice. The temperature between the shaft and the seal is approximately up to 40◦C warmer than the bulk oil temperature. In viscous oil, such as in transmission applications, the difference in bulk temperature and local heat may be even more. [9]
Figure 2: Lip seal functionality with pressure profile
2.2.5 Sealing surface
The sealing system is dependent of the counter-surface the seal is sliding against. The absolute surface roughness is important for the sealing function (a general description of surface roughness is introduced later in this chapter). With too large surface roughness values there is a high possibility of premature wear of the sealing lip and the lifetime will be shortened and allow leakage. If the surface roughness value is too low, there is a risk that the transport of the lubricating oil fails to reach the sealing edge. The sealing edge will become hardened the formation of cracks are eased and the flexibility is decreased. The temperature will increase and the sealing edge will show signs of burning. [9] Table 3 shows the surface roughness guidelines from Freudenberg, which is one of the largest manufacturer of lip seals.[29] Freudenberg also recommend a specified hardness of the surface depending on the relative speed. With low speed ( lower then 5m/s) it is possible to run on untempered surfaces. With increased operating velocity the surface has to be at least 45 Hardness Rockwell, scale C (HRC). If the operating media has dirt present or the circumferential velocity in the contact is higher than 12 m/s, then 60 HRC with a depth hardening of at least 0.3 mm is required. [29]
Type Permissible values At operating pressure >0.1 MPa
Rz [µm] 1.0-5.0 1.0-3.0
Ra [µm] 0.2-0.8 0.2-0.4
Rmax [µm] ≤6.3 6.3
Table 3: Roughnessdata from Simrit Technical Manual 2007
2.2.6 Surface roughness and real contact area
There are several ways to describe the roughness of a surface. The two simplest and com- monly used in the industry are the Ra value (Centre-line average) and the Rq value (root mean square). The Ra value has the disadvantage of not distinguishing between a relative spiky profile and a surface that is much more gently. The Rq value slightly overcome this problem. Another value commonly used is the Rz, which is the average of the five highest peaks and five lowest valleys within a single sampling length.[31]
It is well known that due to the surface roughness, the real contact area is below the apparent
area and dependent on the roughness profile. The contact pressure between asperities in contacting surfaces is therefore dependent on the surface roughness. With a smooth surface and a lower asperity angle the contact area will be large compared with a rough surface and a higher asperity angle. With two completely smooth surfaces, the real contact area will be close to the apparent area. Since the contact pressure is force divided by real contact area, a smooth surface will have lower contact pressure compared to a rough surface.
Black et al. [5] showed that by resolving the forces acting at the interface between contacting asperities of metal in boundary regime the COF can be given as µ = tan(α+arcant(c)), where α is the asperity angle and cis a constant. This means, that a higher asperity angle will lead to a higher COF. [5].
2.3 Lubricants and lubrication
Lubricants are used to reduce the frictional force between surfaces. The coefficient of friction µ is often higher then 0.5 in dry metal and polymer contacts. This high frictional force will lead to a high energy loss, high temperature and wear. A wide variety of greases, oils and solids may be used as lubricants. [12]
2.3.1 Lubrication regimes
In this Master Thesis the lubricant used were various oils, and therefor only research within lubricating oils were performed. When using lubricating oils there are various regimes of lubrication; Hydrodynamic (HD) lubrication, elastohydrodynamic (EHD) lubrication (EHL), mixed lubrication and boundary lubrication. All this regimes are shown in the Fig. 3.
The Stribeck curve in Fig. 3 explains how the friction coefficient µdepends on the viscosity η of the fluid, the relative speedv between the surfaces and the nominal bearing pressure P.
This figure shows that a higher viscosity and higher speed, will lead to a lower COF, while a higher pressure will result in a higher COF. [31]
Figure 3: Stribeck curve with lubrication regimes
Hydrodynamic lubrication: Is also called thick-film lubrication with a thickness of typ- ically 5-500µm. When the sliding surfaces are separated fully by a relative thick fluid film, the regime is in the state of HD. The surfaces are separated by a layer of fluid. The fluid is compressed between the surfaces and creates a pressure sufficient to support the load. HD lubrication is often referred to as the ideal lubricating regime because the fluid film is often thick enough to secure that no solid/asperity contact occur. The COF can be as small as 0.001 in this regime but with higher speed the COF will increase due to viscous drag of the fluid. In this regime the wear and friction is low, and the wear that may occur is often due to start and stop operations. [12, 4]
Elastohydrodynamic lubrication: If the surfaces is counter-conformal then a local pres- sure will be generated in the contact area. The film thickness is thinner (0.5-5µm) than HD and in local areas asperities are in contact. In this condition the lubricating viscosity plays an important role. EHL is occurring in heavily loaded contacts of low geometrical confor- mity, but also in some low elastic modulus contacts of high geometrical conformity, such as lip seals. [12, 4]
Mixed lubrication This region is characterized as the transition between the boundary lubrication regime with the full film lubrication regime. The surfaces starts to separate but is still in contact at some asperities. This may lead to adhesion and wear-particles formation.
The fluid film is typically 0.025-2.5µm thick. [4]
Boundary lubrication: With high pressure, low speed or low viscosity the hydrodynamic forces are insufficient to maintain a lubricating film to separate the surfaces, and direct contact will occur between the asperities. In the boundary lubrication regime there is often relative high wear and high friction. [12, 4]
2.3.2 Mineral oil
Mineral oil is the most used lubricant in the industry. Also in the marine sector mineral oil has been the leading lubricant. It is relative cheap and performs good as a lubricating medium. Mineral oil is produced by distillation of crude petroleum. When mineral oil was first used as lubricant it exhibited little difference in performance in comparison with heavy fuel oils. Past the last decades a lot of improvements has been done, and more sophisticated mineral oils have been created with increased performance. Examples of improvements are removing unwanted elements, like excessive sulphur compounds and waxy elements, and adding additives that improves natural lubricity and oiliness.[31, 28]
2.3.3 Environmentally Acceptable Lubrication
The Vessel General Permit (VGP) for 2013 states that “All vessels must use an EAL in all oil to sea interfaces, unless technically infeasible”. Environmentally acceptable lubricants (EAL) are defined as lubricants that are biodegradable, minimally-toxic and not bioaccumu- lative.[2]
Biodegradable – Is the ability of an oil to be breakdown chemically by micro-organisms.
It is divided into two main types. Primary biodegradation is the loss of one or more active
This can result in a transformation of the compound from a toxic, to a less toxic or non-toxic.
Ultimate biodegradation is a process where the chemical compound is converted to carbon dioxide, water and mineral salt.[1]
Minimally-toxic – EAL must also be of low toxicity to aquatic organisms. There is a lot of different tests to demonstrate if the EAL is minimally-toxic.[1]
Not bioaccumulative – Bioaccumulation is defined as the build-up of chemicals in a tissue of an organism over time. If the organism is prone to bioaccumulation, it will accumulate more chemical the longer it is exposed for it. The concentration of the chemical will build up in the organism if the degradation rate is slow. Therefore, it is desirable to use “Not bioaccumultive” compounds in the EAL. [1]
Additives The EAL can include hundreds of additives which are are added to increase the performance. Adding the right additives improves the protection against oxidative aging, cor- rosion, high pressure, low or high temperature conditions, phase transition, shear, foaming, and hydrolysis. Additives in EAL can be problematic since it must also be biodegradable, minimally-toxic and not bioaccumulative to be approved as an EAL.
2.3.4 Different types of environmentally acceptable lubricants
Fig: 4 shows a typical composition of EAL. EALs are commonly classified according to the base oil used in their formulation. Some of the different EAL’s are:
Vegetable oil Main components are triglycerides, which are natural esters. The chemical compound depends on plant species and how the oil is obtained. Examples of plant species are; Rapeseed, canola, soyabeans or sunflowers. Vegetable oil has a low share in the bio- lubricant market. This is due to the performance issues related to low thermo-oxidative stability and poor cold flow behavior. The most common application is in hydraulic fluids and wire rope grease.[2, 1]
Figure 4: Example of an environmentally acceptable lubricant
Synthetic esters By esterification of bio-based materials, such as a combination of mod- ified animal fat or vegetable oil, synthetic esters can be made. Synthetic esters can be tailor made for their intended application and have therefore many advantages in performance over pure vegetable oils. This is why they are used as a base oil for many vessel applications, such as hydraulic oil, stern tube oil, thruster oil, gear lubricant and grease. Several major oil companies develop synthetic ester-based EALs. [1]
Polyakylene glycols (PAG) PAG are synthetic lubricant base oils. They are typically made by the polymerization of ethylene or propylene oxide. They can be soluble in either oil or water. They are made from petroleum-based materials and can be highly biodegrad- able.[1]
Water Water is also mentioned in the “EAL VGP” as an EAL. At least one company has developed a seawater-lubricated stern tube system.[1]
2.3.5 Aging in lubricant
Since the intention for most seals is to retain a fluid, the interaction of the seal material and the fluid plays an important role. Different oils will influence the performance of the elastomers, such as the friction and wear properties. The elastomer may swell and degrade in the contact of different oils. If the elastomer absorbs the lubricant, the wear will increase.
Modifi M. et al [18] explained that this was due to aggravated cracking of the solvent of the elastomer. There are several possibilities to moderate the swelling and absorption of the oils.
Increasing the presence of the polar side-groups in the backbone chain and cross-linking the elastomer (constraints will limit the amount of solvent absorbed into the elastomer). Nitrile rubber (NBR) contains of acrylonite and butadiene. As higher the content of acrylonite as lower will the swelling be in the fluid. The disadvantage using a high acrylonite-content is that the elasticity and temperature flexability will become poor. Fluoroelastomers (FKM) are known for good resistance against heat and aggressive fluids. However, there is not much literature available on studies within the topic. The different oils will affect the elastomer in different ways. In general, an ester base oil will be more aggressive on elastomers compared to mineral oils. The ester based oil is more aggressive due to the existence of carboxylic groups which are polar and will easier be absorbed into the elastomer. [20] [18]
Mofidi M. et al [20] performed experiments with NBR with an acrylonite content of 28%. The elastomer was tested in different oils, including ester based oil and mineral oil. The elastomer was soaked in the different oils for one week at 125◦C. The elastomer aged in synthetic ester showed a decreased COF after soaking. Also the abrasive wear was investigated, and the result showed that aging in any base fluid, especially in ester base fluids leads to more abrasive wear. Dong C.L. et al [7] also performed experiments with dry aged NBR. The most interesting findings in this study were that with increased aging time and temperature, the hardness of the elastomer and the COF decreased. Even thought the COF was reduced the wear increased as a result of decrease in tensile and tear strength. [7]
2.3.6 Particles in the lubricant
A sealing-system in operation will probably be contaminated from wear, sand, corrosion, dust and seawater. This is therefore an important factor regarding the life and sealing ability of the seal. Such particles may lead to an increase in abrasive wear. Additional, they can be embedded in the soft surface of the elastomer resulting in abrasion wear of the counter surface. [18]
2.3.7 Water in the lubricant
When operating thrusters and stern-tubes there will always be a possibility that water is entering the system. Often these systems has a filtering device to filter the water from the lubricant. A common method is to settle the water by gravity to the bottom of the system and than drain it with a small pump.
Kl¨uber Lubrication released a ”White paper” [11] on the problematic on hydrolysis in ester based lubricants. The problem with water in the EAL is that it is designed to dissolve in water, this is why it is environmentally acceptable. The synthetic ester oils are made by an organic or inorganic acid and alcohol. Water is created as an bi-product in the fabrication of the ester based oil.
The opposite reaction is called hydrolysis. When water comes in contact with the ester based oil, there is a chance that it will react and create acid and alcohol. Kluber Lubrication recommend that the gear oil should not exceed 200-1000 PPM (Parts per million). In stern tube oil the maximum amount of water is five%. The speed of the hydrolysis is determined by several factors [11]:
• The chemistry of the base oil
• Percentage of water in the oil
• Temperature which will accelerate or deaccelrate the reaction
• Formation of reaction products that support further hydrolysis
• Additives in the oil that support hydrolysis
2.3.8 Viscosity
The most important rheological property of an oil is the viscosity. Viscosity describes the resistance of a fluid to share and is defined as the shear stress on a plane within the fluid per unit velocity gradient normal to the plane. [12]
The viscosity of the oil will influence the heat generation because it will directly influence the thickness of the lubricating film between the seal lip and shaft at a given pressure. A high viscosity will lead to a thicker oil film and decrease the friction. A to low viscosity will allow the lubricant film to be so thin that it will not manage to separate the lip from the shaft. If so, the system will have increased friction and wear. [9]
2.3.9 Viscosity index
To characterize the oil for their temperature-dependency, the viscosity index (VI) is used.
An oil with a low VI will be highly influenced by the temperature, and the viscosity will be far lower when increasing the temperature. An oil with high VI on the other hand will not be significant influenced by the temperature, and the viscosity of the oil will remain high even with high temperatures. Two different oils can have the same viscosity at e.g. room temperature, but different at a higher/lower temperature as shown in Fig: 5. It is therefore important to choose the correct VI when selecting a lubricant for a system.[12]
Figure 5: Two fluids with different VI
3 Experimental setup
3.1 Sample preparation
In this Master thesis different types of elastomers were tested against different steel surfaces.
The two lubricants used were a conventional mineral oil and an EAL.
Figure 6: Picture of metal sample used in tests
Figure 7: Picture of a pin used in the tests
3.1.1 Metal samples
The metal samples (Fig. 6) were provided by the industrial partner. They were manufactured in an equally size; 58 mm x 38 mm x 4 mm. With the two holes, the plates could be fixed to the tribomachine. Two plate materials were tested. S355J2G3-steel (S355) and NiCr-steel (NiCr), both are used in the real application. The surface of the S355-samples was grinded to the desired roughnessRa= 0.4µm. The grinding was performed by manual grinding with sandpaper. The samples were moved around while grinding to secure a random direction of the surface roughness as in the application. The NiCr-samples were finished before delivering with a surface roughness of Ra = 0.4 µm. To confirm the roughness a profilometer was used to measure each sample three times. Samples were cleaned before and after use in a ultrasonic bath, using ethanol.
3.1.2 Elastomer pins
The elastomer samples (Fig. 7) were provided by the industrial partner. 72 NBR 902 (NBR- 1), 75 FKM 585 (FKM-1), an unspecified NBR (NBR-2) and an unspecified FKM (FKM-2) were investigated. NBR-1 and FKM-1 are the materials used in the real application. The most relevant information from the data-sheets are summarized in Table 4. The NBR-2 and FKM-2 samples came as o-rings which already had the correct pin-diameter. This material were only identified as ”NBR” and ”FKM”, and therefore no exact information about the physical properties were available. The elastomers were cut to a length of 7 mm and were afterwards grinded manually until the pins were 6 mm and the ends were smooth and planar.
After this, the elastomer was glued to a metal pin, to increase the stability in the sample holder. All the pins were stored for at least 24 hours to secure that the glue was dry. The pins were weighted and investigated with 3D confocal microscope before and after testing.
The pins were cleaned in an ultrasonic bath before and after testing, using ethanol.
Properties 72 NBR 902 75 FKM 585 Unit
Density 1.43 2.05 g/cm3
Hardness 75 75 Shore
Rebound resilience 26 6 %
Modulus 7.2 6.9 MPa
Tensile strength 13.8 12.8 MPa
Elongation at break 360 250 %
Compression set 30 28 %
High temperature 100 200 ◦C
Low temperature -29 -16 ◦C
Table 4: Physical properties, gathered from data-sheet of the elastomers
3.1.3 Environmentally Acceptable Lubricant
Kl¨uberbio EG 2-68 (Kluber) was chosen as EAL for this thesis, since it is one of the approved oils of the industrial partner. According to the product information at Kl¨uber Lubrication’s homepage, the benefits of this oil are:
• Comply with the requirements for Environmentally Acceptable Lubricants as defined in Appendix A of the EPA 2013 VPG (Vessel General Permit)
• Readily biodegradable and not toxic to marine organisms reducing environmental im- pact in the event of leakage
• Kl¨uberbio EG 2 oils have a high scuffing resistance, protecting gear teeth reliably against fretting damage even at high peak loads
• Standard NBR and FKM elastomers of the leading propeller shaft seal manufacturers are resistant to and approved for use with Kl¨uberbio EG 2 oils, preventing leakages and impurities
The Kl¨uberbio EG 2-68 is readily biodegradable, based on synthetic ester oil, minimally toxic, non-bioaccumulating. It contains less than 90% of renewable raw materials and complies with the European Ecolabel. The oil is developed for lubrication of gearboxes in ships, particularly for thruster and rudder propellers. The viscosity index is approximately 140, the viscosity grade is 68 and the max service temperature is 100◦C. With the given temperature of 60◦C and VI of 140, the viscosity is 32.2 mm2/2.[16]
3.1.4 Conventional Mineral Oil
Shell Omala S2 G 68 (Shell) was used as conventional oil for testing and is in use in the application. The oil is described as a high quality, extreme-pressure oil, designed for the lubrication of heavy-duty industrial gears. It has a high load carrying capacity and an anti-friction characteristic which gives its good performance in gears and other industrial applications. The oil helps reducing wear at steel and bronze components. It also has an excellent water separation property, so that water in the system can be drained easily. The viscosity index is approximately 100, the viscosity grade is 68 and it can be used up to 100◦C.
At the given temperature of 60◦C the viscosity is 28.66 mm2/2. [25]
3.1.5 Contamination
The tests performed with particle contamination in the lubricant used ”Sika WS - F800” as abrasive particles [26]. This is Silicon Carbide particles (SiC) and were chosen because of the defined particle-size available. Approximately 1 ml SiC particles were added to 10 ml lubricant.
When testing the lubricant contaminated with water, the oil and water was put in a sample- glass and thoroughly stirred and immediately poured in the lubricant-retainer. The water and oil did not emulsify. Same procedure was used when mixing the lubricants with each other. Apparently the oils was mixed into an homogeneous mixture. Through the results acronyms will be used for the different mixtures of oil+oil and oil+water. The acronyms are listed in Table 5.
Acronym Lubricant mixture 75Shell/Kluber 75% Shell + 25% Kluber 50Shell/Kluber 50% Shell + 50% Kluber 25Shell/Kluber 25% Shell + 75% Kluber 75Shell/Water 75% Shell + 25% Water 50Shell/Water 50% Shell + 50% Water 25Shell/Water 25% Shell + 75% Water 75Kluber/Water 75% Kluber + 25% Water 50Kluber/Water 50% Kluber + 50% Water 25Kluber/Water 25% Kluber + 75% Water Table 5: Acronym of the lubricant mixtures
3.2 Pin on plate
Figure 8: Pin on plate Figure 9: The TE88
To perform the friction testing a ”TE88 Multi-Station Friction and Wear Test Machine (TE88)” (Phoenix) pin on plate (Fig. 8) configuration was used. In Fig. 9 a typical setup can be seen. The steel sample was mounted in a station (Nr.1 on Fig. 9) with a reservoir to retain the lubricant. The station is equipped with electric heating and a thermocouple to control the temperature. The station is mounted on a second plate (Nr.2 on Fig. 9), which is located by ball bushings and a linear bearing assembly to secure linear movement of the station. The plate is reciprocated by a variable throw crank(Nr.3 on Fig. 9). The elastomer pin is mounted in an arm (Nr.4 on Fig. 9) which is static during testing. The loading of the pin is achieved by a pneumatic piston (Nr.5 on Fig. 9) and is applied at the end of the arm while the other end of the arm is hinged. The arm is locked and the pressure on the pads are measuring force. The temperature, frictional load, load at pin and the velocity is recorded.
The velocity follows a sinusoidal function and is maximal in the middle of the stroke and zero at the turning points. All the test were done with a frequency of 2 Hz on the fly wheel and a wear track of 40 mm. [15]
Finding the exact nominal contact pressure (CP) from the real application was not achieved.
It was decided to use two different contact pressure; 1 MPa and 2 MPa which might be very
All the tests were repeated at least once. Different contact pressure, metal counter-surfaces, elastomers, lubricants, oil-oil mixtures and oil-water mixtures and different contamination were tested. The variations of tests can be observed in Fig. 11, where all the different test setups for the friction tests are illustrated. The mean coefficient of friction for every test was calculated. Since the velocity in the pin on plate is not constant, an area with a distance of 4 mm, in the middle of the wear-track was investigated. The maximum velocity between the metal-sample and the elastomer-pin is between 250 mm/s and 230 mm/s, which can be observed at Fig. 10.
Figure 10: Trace/retrace
S355
2 MPa
NBR-1
Shell/Kluber mixtures (25/75%, 50/50%, 75/25%) Shell/water mixtures (25/75%, 50/50%, 75/25%) Kluber/water mixtures (25/75%, 50/50%, 75/25%)
Shell Kluber
Shell contaminated with sand Kluber contaminated with sand
FKM-1
Shell/Kluber mixtures (25/75%, 50/50%, 75/25%) Shell/water mixtures (25/75%, 50/50%, 75/25%) Kluber/water mixtures (25/75%, 50/50%, 75/25%)
Shell Kluber
Shell contaminated with sand Kluber contaminated with sand
NBR-2
ShellKluber
FKM-2
ShellKluber
1 MPa
NBR-1
ShellKluber
FKM-1
ShellKluber
NiCr
2 MPa
NBR-1
KluberFKM-1
Kluber1 MPa
NBR-1
KluberFKM-1
Figure 11: Test-setups for friction tests